Basics of Vibration

Vibration theory amp analysis

What is Vibration

Vibration Terms

Time Waveform Analysis

complex time waveform

individual vibration signalscombine to form a complextime waveform showing overallvibration

frequency

low freq

high fre

q

timeoverall vibration

Scale Factorsndash When comparing overall vibration signals it is

imperative that both signals be measured on the same frequency range and with the samescale factors

Measurements amp Units

Displacement (Distance)mils or micrometer mm

Velocity (Speed - Rate of change of displcmt)insec or mmsec

Acceleration (Rate of change of velocity)Grsquos or insec2 or mmsec2

10 100 1000 10000Frequency (Hz)

10

10

01

1

001

100

Displacement (microns)Acceleration(gs - 981msec2)

Velocity (mmsec)

Common MachineryOperating Range

Amplitude(microns

mmsec grsquos

Sensor Relationships

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

What is Vibration

Vibration Terms

Time Waveform Analysis

complex time waveform

individual vibration signalscombine to form a complextime waveform showing overallvibration

frequency

low freq

high fre

q

timeoverall vibration

Scale Factorsndash When comparing overall vibration signals it is

imperative that both signals be measured on the same frequency range and with the samescale factors

Measurements amp Units

Displacement (Distance)mils or micrometer mm

Velocity (Speed - Rate of change of displcmt)insec or mmsec

Acceleration (Rate of change of velocity)Grsquos or insec2 or mmsec2

10 100 1000 10000Frequency (Hz)

10

10

01

1

001

100

Displacement (microns)Acceleration(gs - 981msec2)

Velocity (mmsec)

Common MachineryOperating Range

Amplitude(microns

mmsec grsquos

Sensor Relationships

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration Terms

Time Waveform Analysis

complex time waveform

individual vibration signalscombine to form a complextime waveform showing overallvibration

frequency

low freq

high fre

q

timeoverall vibration

Scale Factorsndash When comparing overall vibration signals it is

imperative that both signals be measured on the same frequency range and with the samescale factors

Measurements amp Units

Displacement (Distance)mils or micrometer mm

Velocity (Speed - Rate of change of displcmt)insec or mmsec

Acceleration (Rate of change of velocity)Grsquos or insec2 or mmsec2

10 100 1000 10000Frequency (Hz)

10

10

01

1

001

100

Displacement (microns)Acceleration(gs - 981msec2)

Velocity (mmsec)

Common MachineryOperating Range

Amplitude(microns

mmsec grsquos

Sensor Relationships

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Time Waveform Analysis

complex time waveform

individual vibration signalscombine to form a complextime waveform showing overallvibration

frequency

low freq

high fre

q

timeoverall vibration

Scale Factorsndash When comparing overall vibration signals it is

imperative that both signals be measured on the same frequency range and with the samescale factors

Measurements amp Units

Displacement (Distance)mils or micrometer mm

Velocity (Speed - Rate of change of displcmt)insec or mmsec

Acceleration (Rate of change of velocity)Grsquos or insec2 or mmsec2

10 100 1000 10000Frequency (Hz)

10

10

01

1

001

100

Displacement (microns)Acceleration(gs - 981msec2)

Velocity (mmsec)

Common MachineryOperating Range

Amplitude(microns

mmsec grsquos

Sensor Relationships

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Scale Factorsndash When comparing overall vibration signals it is

imperative that both signals be measured on the same frequency range and with the samescale factors

Measurements amp Units

Displacement (Distance)mils or micrometer mm

Velocity (Speed - Rate of change of displcmt)insec or mmsec

Acceleration (Rate of change of velocity)Grsquos or insec2 or mmsec2

10 100 1000 10000Frequency (Hz)

10

10

01

1

001

100

Displacement (microns)Acceleration(gs - 981msec2)

Velocity (mmsec)

Common MachineryOperating Range

Amplitude(microns

mmsec grsquos

Sensor Relationships

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Measurements amp Units

Displacement (Distance)mils or micrometer mm

Velocity (Speed - Rate of change of displcmt)insec or mmsec

Acceleration (Rate of change of velocity)Grsquos or insec2 or mmsec2

10 100 1000 10000Frequency (Hz)

10

10

01

1

001

100

Displacement (microns)Acceleration(gs - 981msec2)

Velocity (mmsec)

Common MachineryOperating Range

Amplitude(microns

mmsec grsquos

Sensor Relationships

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

10 100 1000 10000Frequency (Hz)

10

10

01

1

001

100

Displacement (microns)Acceleration(gs - 981msec2)

Velocity (mmsec)

Common MachineryOperating Range

Amplitude(microns

mmsec grsquos

Sensor Relationships

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Multi-Parameter MonitoringSame Data in Velocity and Acceleration

VelocitySpectrum

AccelerationSpectrum

On the same bearing cap low freq events (imbalance misalignment etc) show best in the velocity spectrum while high freq events (bearing faults gearmesh) show best in the acceleration spectrum

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Accelerometers

bull Rugged Devicesbull Operate in Wide Frequency

Range (Near 0 to above 40 kHz)bull Good High Frequency Responsebull Some Models Suitable For High

Temperaturebull Require Additional Electronics

(may be built into the sensor housing)

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

What is vibrationComplex signal

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

FFT Signal Processing

Time

Am

plitu

de

T ime

Am

plitu

de

Frequency

Am

plitu

de

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Narrow Bands with trend

T re n d o fB a la n c e

A la rm

Ampl

itude

S u b -H a rm o n ic 1 X 2 X B e a r in g B e a r in g G e a rs B e a r in g

1 x 2 x

3in s e c

1in s e cT im e

(D a y s )T im e

(D a y s )

T re n d o f B e a r in g s

1 0 x

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Alarm Types ndash Narrow Bands

A2 - 824 BPFI Pomp PNVP1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

2

4

6

8

10

12

14

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

47500 9542 06356

Imba

lanc

eM

isal

ignm

ent

Loos

enes

s

Bea

ring

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Overall Vibrationbull The total vibration energy

measured within a specific frequency range

ndash includes a combination of all vibration signals within measured frequency range

ndash does not include vibration signals outside measured frequency range

ndash produces a numerical value

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Alarm Types ndash Overall Alarm

bull Look to the global vibration levelA2 - 824 BPFI Pomp PNV

P1K10 -PNV POMP NIET-KOPP VERTIKAAL

Label BPFI with 1xrPM modulations

Route Spectrum 30-jan-96 151451

OVERALL= 1352 V-DG RMS = 1346 LOAD = 1000 RPM = 2987 (4978 Hz)

0 500 1000 1500 2000 2500

0

3

6

9

12

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Fault Limit

Freq Ordr Spec

13219 2655 119

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Analyse of data Spectra Waveform and Trends

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibrationraquo -Imbalanceraquo -Misalignmentraquo -Loosenessraquo -Bearing problemsraquo -Belt problemsraquo -Gear problemsraquo -Lubrificationraquo -Electrical problemsraquo -Resonanceraquo -Sleeve Bearing problemsraquo -Other

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysis

bull Of all the parameters that can be measured non intrusively in industry today the one containing the most information is the vibration signature Art Crawford

bull Vibration Analysis is the foundation of a predictive maintenance program

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

SIGNATURE ANALYSISSIGNATURE ANALYSIS

bull Which frequencies exist and what are the relationships to the fundamental exciting frequencies

bull What are the amplitudes of each peakbull How do the peaks relate to each otherbull If there are significant peaks what are their

source

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysis

Unbalance

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

COUPLE UNBALANCECOUPLE UNBALANCE

bull 1800 out of phase on the same shaftbull 1X RPM always present and normally dominatesbull Amplitude varies with square of increasing speedbull Can cause high axial as well as radial amplitudesbull Balancing requires Correction in two planes at 180o

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

OVERHUNG ROTOR UNBALANCE

OVERHUNG ROTOR UNBALANCE

bull 1X RPM present in radial and axial directionsbull Axial readings tend to be in-phase but radial readings

might be unsteadybull Overhung rotors often have both force and couple

unbalance each of which may require correction

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Diagnosing UnbalanceDiagnosing Unbalancebull Vibration frequency equals rotor

speed

bull Vibration predominantly RADIAL in direction

bull Stable vibration phase measurement

bull Vibration increases as square of speed

bull Vibration phase shifts in direct proportion to measurement direction

900

900

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysis

MisalignmentBent shaft

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ANGULAR MISALIGNMENT

ANGULAR MISALIGNMENT

bull Characterized by high axial vibrationbull 1800 phase change across the couplingbull Typically high 1 and 2 times axial vibrationbull Not unusual for 1 2 or 3X RPM to dominatebull Symptoms could indicate coupling problems

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

PARALLEL MISALIGNMENT

PARALLEL MISALIGNMENT

bull High radial vibration 1800 out of phasebull Severe conditions give higher harmonicsbull 2X RPM often larger than 1X RPMbull Similar symptoms to angular misalignmentbull Coupling design can influence spectrum

shape and amplitude

RadialRadial

1x1x 2x2x4x4x

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

MISALIGNED BEARINGMISALIGNED BEARING

bull Vibration symptoms similar to angular misalignment

bull Attempts to realign coupling or balance the rotor will not alleviate the problem

bull Will cause a twisting motion with approximately 1800 phase shift side to side or top to bottom

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

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ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

BENT SHAFTBENT SHAFT

bull Bent shaft problems cause high axial vibrationbull 1X RPM dominant if bend is near shaft centerbull 2X RPM dominant if bend is near shaft endsbull Phase difference in the axial direction will tend

towards 1800 difference

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

OTHER SOURCES OF HIGH AXIAL VIBRATION

OTHER SOURCES OF HIGH AXIAL VIBRATION

a Bent Shafts

b Shafts in Resonant Whirl

c Bearings Cocked on the Shaft

d Resonance of Some Component in the Axial Direction

e Worn Thrust Bearings

f Worn Helical or Bevel Gears

g A Sleeve Bearing Motor Hunting for its Magnetic Center

h Couple Component of a Dynamic Unbalance

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysis

Mechanical looseness

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

MECHANICAL LOOSENESS (A)

MECHANICAL LOOSENESS (A)

bull Caused by structural looseness of machine feetbull Distortion of the base will cause ldquosoft footrdquo

problemsbull Phase analysis will reveal aprox 1800 phase

shift in the vertical direction between the base plate components of the machine

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

MECHANICAL LOOSENESS (B)

MECHANICAL LOOSENESS (B)

bull Caused by loose pillow block boltsbull Can cause 05 1 2 and 3X RPMbull Sometimes caused by cracked frame structure

or bearing block

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

MECHANICAL LOOSENESS (C)

MECHANICAL LOOSENESS (C)

bull Phase is often unstablebull Will have many harmonicsbull Can be caused by a loose bearing liner excessive

bearing clearance or a loose impeller on a shaft

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysis

Sleeve bearingRotor rub

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

SLEEVE BEARINGWEAR CLEARANCE

PROBLEMS

bull Later stages of sleeve bearing wear will give a large family of harmonics of running speed

bull A minor unbalance or misalignment will cause high amplitudes when excessive bearing clearances are present

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ROTOR RUBROTOR RUB

bull Similar spectrum to mechanical loosenessbull Usually generates a series of frequencies which

may excite natural frequenciesbull Sub harmonic frequencies may be presentbull Rub may be partial or through a complete

revolution

Truncated waveform

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

OIL WHIP INSTABILITYOIL WHIP INSTABILITY

bull Oil whip may occur if a machine is operated at 2X the rotor critical frequency

bull When the rotor drives up to 2X critical whirl is close to critical and excessive vibration will stop the oil film from supporting the shaft

bull Whirl speed will lock onto rotor critical If the speed is increased the whip frequency will not increase

oil whirl

oil whip

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

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ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

OIL WHIRL INSTABILITYOIL WHIRL

INSTABILITY

bull Usually occurs at 42 - 48 of running speedbull Vibration amplitudes are sometimes severebull Whirl is inherently unstable since it increases centrifugal

forces therefore increasing whirl forces

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

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ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

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ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

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ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Resonance

typically 10 or greater

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

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ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

RESONANCERESONANCE

bull Resonance occurs when the Forcing Frequency coincides with a Natural Frequency

bull 1800 phase change occurs when shaft speed passes through resonance

bull High amplitudes of vibration will be present when a system is in resonance

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

BELT PROBLEMS (A)BELT PROBLEMS (A)

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

bull Often 2X RPM is dominantbull Amplitudes are normally unsteady sometimes pulsing with either

driver or driven RPMbull Wear or misalignment in timing belt drives will give high amplitudes

at the timing belt frequencybull Belt frequencies are below the RPM of either the driver or the driven

WORN LOOSE OR MISMATCHED BELTSWORN LOOSE OR MISMATCHED BELTS

BELT FREQUENCYHARMONICS

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

BELT PROBLEMS (D)BELT PROBLEMS (D)

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

bull High amplitudes can be present if the belt natural frequency coincides with driver or driven RPM

bull Belt natural frequency can be changed by altering the belt tension

BELT RESONANCEBELT RESONANCE

RADIAL

1X RPM

BELT RESONANCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

bull If gap between vanes and casing is not equal Blade Pass Frequency may have high amplitude

bull High BPF may be present if impeller wear ring seizes on shaft

bull Eccentric rotor can cause amplitude at BPF to be excessive

BPF = BLADE PASS FREQUENCY

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

bull Flow turbulence often occurs in blowers due to variations in pressure or velocity of air in ducts

bull Random low frequency vibration will be generated possibly in the 50 - 2000 CPM range

FLOW TURBULENCEFLOW TURBULENCE

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

HYDRAULIC AND AERODYNAMIC FORCES

HYDRAULIC AND AERODYNAMIC FORCES

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

bull Cavitations will generate random high frequency broadband energy superimposed with BPF harmonics

bull Normally indicates inadequate suction pressurebull Erosion of impeller vanes and pump casings may occur if

left uncheckedbull Sounds like gravel passing through pump

CAVITATIONCAVITATION

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

BEAT VIBRATIONBEAT VIBRATION

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

bull A beat is the result of two closely spaced frequencies going into and out of phase

bull The wideband spectrum will show one peak pulsating up and down

bull The difference between the peaks is the beat frequency which itself will be present in the wideband spectrum

WIDEBAND SPECTRUM

ZOOMSPECTRUM

F1 F2

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysis

Electrical

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Stator problems generate high amplitudes at 2FL (2X line frequency )

bull Stator eccentricity produces uneven stationary air gap vibration is very directional

bull Soft foot can produce an eccentric stator

STATOR ECCENTRICITY SHORTED LAMINATIONSSTATOR ECCENTRICITY SHORTED LAMINATIONSAND LOOSE IRONAND LOOSE IRON

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

bull Electrical line frequency(FL) = 50Hz = 3000 cpm60HZ = 3600 cpm

bull No of poles (P)

bull Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm

bull Synchronous speed (Ns) = 2xFL)

bull Slip frequency ( FS )= Synchronous speed ndash Rotor rpm

bull Pole pass frequency (FP )= Slip Frequency x No of Poles

bullbull Electrical line frequency(Electrical line frequency(FLFL) = ) = 50Hz = 3000 cpm50Hz = 3000 cpm60HZ = 36060HZ = 3600 cpm0 cpm

bullbull No of polesNo of poles ((PP))

bullbull Rotor Bar Pass Frequency (Rotor Bar Pass Frequency (FbFb) = ) = No of rotor bars x Rotor rpm No of rotor bars x Rotor rpm

bullbull Synchronous speed (Synchronous speed (NsNs)) = = 2xFL2xFL))

bullbull Slip frequency ( Slip frequency ( FFS S )= )= Synchronous speed Synchronous speed ndashndash Rotor rpmRotor rpm

bullbull Pole pass frequency (Pole pass frequency (FFPP )=)= Slip Frequency x No of PolesSlip Frequency x No of Poles

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

FREQUENCIES PRODUCED BY ELECTRICAL MOTORS

PP

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Loose stator coils in synchronous motors generate high amplitude at Coil Pass Frequency

bull The coil pass frequency will be surrounded by 1X RPM sidebands

SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull Phasing problems can cause excessive vibration at 2FL with 13 FL sidebands

bull Levels at 2FL can exceed 25 mmsec if left uncorrectedbull Particular problem if the defective connector is only

occasionally making contact

POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ELECTRICAL PROBLEMSELECTRICAL PROBLEMS

bull 1X 2X 3X RPM with pole pass frequency sidebands indicates rotor bar problems

bull 2X line frequency sidebands on rotor bar pass frequency (RBPF) indicates loose rotor bars

bull Often high levels at 2X amp 3X rotor bar pass frequency and only low level at 1X rotor bar pass frequency

ROTOR PROBLEMSROTOR PROBLEMS

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysis

Gear

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

CALCULATION OF GEAR MESH FREQUENCIES

CALCULATION OF GEAR MESH FREQUENCIES

20 TEETH20 TEETH

51 TEETH51 TEETH

1700 RPM1700 RPM

31 TEETH31 TEETH

HOW MANY TEETH ON THIS GEARHOW MANY TEETH ON THIS GEAR

8959 RPM8959 RPM

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

GEARSNORMAL SPECTRUM

GEARSNORMAL SPECTRUM

bull Normal spectrum shows 1X and 2X and gear mesh frequency GMF

bull GMF commonly will have sidebands of running speedbull All peaks are of low amplitude and no natural frequencies

are present

14 teeth

8 teeth GMF= 21k CPM2625 rpm

1500 rpm

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

bull Gear Mesh Frequencies are often sensitive to loadbull High GMF amplitudes do not necessarily indicate a

problembull Each analysis should be performed with the system at

maximum load

GEARSTOOTH LOADGEARSTOOTH LOAD

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

GEARSTOOTH WEARGEARS

TOOTH WEAR

bull Wear is indicated by excitation of natural frequencies along with sidebands of 1X RPM of the bad gear

bull Sidebands are a better wear indicator than the GMFbull GMF may not change in amplitude when wear occurs

14 teeth1500 rpm

8 teeth2625 rpm

GMF = 21k CPM

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

GEARSGEAR ECCENTRICITY AND BACKLASH

GEARSGEAR ECCENTRICITY AND BACKLASH

bull Fairly high amplitude sidebands around GMF suggest eccentricity backlash or non parallel shafts

bull The problem gear will modulate the sidebandsbull Incorrect backlash normally excites gear natural

frequency

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

GEARSGEAR MISALIGNMENT

GEARSGEAR MISALIGNMENT

bull Gear misalignment almost always excites second order or higher harmonics with sidebands of running speed

bull Small amplitude at 1X GMF but higher levels at 2Xand 3X GMF

bull Important to set Fmax high enough to capture at least2X GMF

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

GEARSCRACKED BROKEN TOOTH

GEARSCRACKED BROKEN TOOTH

bull A cracked or broken tooth will generate a high amplitude at 1X RPM of the gear

bull It will excite the gear natural frequency which will be sidebanded by the running speed fundamental

bull Best detected using the time waveformbull Time interval between impacts will be the reciprocal of the

1X RPM

TIME WAVEFORM

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

GEARSHUNTING TOOTHGEARS

HUNTING TOOTH

bull Vibration is at low frequency and due to this can often be missed

bull Synonymous with a growling soundbull The effect occurs when the faulty pinion and gear teeth

both enter mesh at the same timebull Faults may be due to faulty manufacture or mishandling

fHt = (GMF)Na(TGEAR)(TPINION)

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Vibration analysisBearings

Outer Race(BPFO)

Inner Race(BPFI)

Ball Spin(BSF)

Cage or Train FTF

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

D0

D1DB

Note shaft turning outer race fixedF = frequency in cpmN = number of balls

BPFI = Nb2 (1+(BdPd)cosӨ) RPM

BPFO = Nb2 (1-(BdPd)cosӨ) RPM

BSF = Pd2Bd (1-((BdPd)cosӨ)2) RPM

FTF = frac12 (1-((BdPd)cosӨ)) RPM

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ROLLING ELEMENT BEARINGS STAGE 1 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 1 FAILURE MODE

bull Earliest indications in the ultrasonic rangebull These frequencies evaluated by Spike EnergyTM gSE

HFD(g) and Shock Pulsebull Spike Energy may first appear at about 025 gSE for this

first stage

gSE

ZONE BZONE A ZONE C ZONE D

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ROLLING ELEMENT BEARINGS STAGE 2 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 2 FAILURE MODE

bull Slight defects begin to ring bearing component natural frequencies

bull These frequencies occur in the range of 30k-120k CPMbull At the end of Stage 2 sideband frequencies appear above

and below natural frequencybull Spike Energy grows eg 025-050gSE

ZONE A ZONE B ZONE C ZONE D

gSE

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

ROLLING ELEMENT BEARINGS STAGE 3 FAILURE MODEROLLING ELEMENT

BEARINGS STAGE 3 FAILURE MODE

bull Bearing defect frequencies and harmonics appearbull Many defect frequency harmonics appear with wear the

number of sidebands growbull Wear is now visible and may extend around the periphery of

the bearingbull Spike Energy increases to between 05 -10 gSE

ZONE A ZONE B ZONE C ZONE D

gSE

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Examples

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Singing Propeller

0 50 100 150 200 250 300 350 400

0

006

012

018

024

030

036

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

0 50 100 150 200 250 300 350 400

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Starboard side Port side

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

LFPS 1 024 Route Spectrum 28-JUL-06 215644

OVRALL= 279 V-DG RMS = 276 LOAD = 1000 RPM = 92 RPS = 153

80 100 120 140 160 180 200

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec Dfrq

14228 9324 186 1534

bullbull Sideband activity around the Sideband activity around the troubled frequency (140 Hz)troubled frequency (140 Hz)

bullbull The modulationsideband The modulationsideband activity tells us that the activity tells us that the troubled frequency is working troubled frequency is working along with the rpm of the along with the rpm of the shaftshaft

bullbull Dfrq (Delta frequency) = Dfrq (Delta frequency) = 1534 Hz (60sec)= 92 RPM1534 Hz (60sec)= 92 RPM

bullbull 92 rpm = shaft speed when 92 rpm = shaft speed when measurements were takenmeasurements were taken

Singing PropellerDescribing the frequency spectra

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Singing PropellerConclusionAfter thorough measurementsanalysis our conclusion is that the port side propeller suffers from a phenomenon called a singing propeller The conclusion is justified by

bull A frequency of approximately 140 Hz is causing the noisevibration bull This frequency is independent from rpm within the troubled range of propeller revolution (60-105 rpm)bull The ~140 Hz frequency only appears on the port side propeller shaft This was confirmed by single propeller transit on both starboard and port sidebull The ~140 Hz frequency measured has sideband (modulation) which is directly connected to the speed of the port side shaft This indicates that the troubled frequency is situated somewhere along this shaftbull There is no other ldquorpm independentrdquo component along port side shaft line that can be a source to this frequency The size and weight to the propeller can possibly fit to the ldquosingingrdquo frequency

RecommendationGrinding an anti singing edge on the propellerResult The grinding of the propeller blades were carried out and the singing tone disappeared

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damageSF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL

Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5436 26608 03517

gtFAG 6322 F=BPFI 544

F F F F F F F F F F

SF8000182 645 AKSEL REIMHJUL 1 LAGER RADIELL Route Spectrum 10-MAY-05 120736

OVRALL= 1023 V-DG RMS = 171 LOAD = 1000 RPM = 2937 RPS = 4895

0 20 40 60 80 100

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

3540 17327 01331

gtFAG 6322 E=BPFO 356

E E E E E E E E E E

Observing frequencies that matches ball pass frequencies inner race (fault frequencies BPFI) on bearing FAG 6322

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing FAG 6322

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damageTrend Display of 1 - 20 kHz

-- Baseline -- Value 1143 Date 26-FEB-03

0 200 400 600 800 1000

0

05

10

15

20

25

30

35

40

45

50

Days 10-JAN-03 To 10-MAY-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

10-MAY-05 120740 4281

OFF ROUTE ORP OFF ROUTE MEASUREMENT POINT DATA

Label WF 63 1RER-1

Route Spectrum 10-MAY-05 120949 (Demod-HP 1000 Hz)

OVRALL= 149 A-DG RMS = 150 LOAD = 1000 RPM = 2937 RPS = 4895

0 2 4 6 8 10 12 14 16 18 20 22

0

02

04

06

08

10

Frequency in Order

RM

S A

ccel

erat

ion

in G

-s

Ordr Freq Spec

5433 26594 715

gtFAG 6322 F=BPFI 544

F F F

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Also the demodulated measurement indicates fault frequencies from the bearing inner ring on bearing FAG 6322

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

FAG6322 (outer race)FAG6322 (outer race)

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damageOuter ring

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELL Route Spectrum 01-MAR-05 094729

OVRALL= 1510 V-DG RMS = 414 LOAD = 1000 RPM = 2622 RPS = 4370

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

30017 6869 00788

gtSKF NU2224 E=BPFO 2996

E E E E E E E E E E

SF8000129 716 AKSEL REIMHJUL 2 LAGER RADIELLTrend Display of 1 - 20 kHz

-- Baseline -- Value 986 Date 03-FEB-03

0 100 200 300 400 500 600 700 800

0

1

2

3

4

5

6

7

Days 03-FEB-03 To 01-MAR-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

01-MAR-05 094737 5531

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU2224

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damageOuter ring

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing TMK HH840200 (HH840249210)

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Route Spectrum 06-JUN-05 210414

OVRALL= 2182 V-DG RMS = 658 LOAD =15500 RPM = 1505 RPS = 2509

0 1000 2000 3000 4000

0

03

06

09

12

15

18

21

24

27

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

25502 1017 102

gtTMK HH840210249 E=BPFO 2565

E E E E E E E E

003 - GEAR SN 618860320101G0008 -086 GEARINNGAKS 1LAGER RADIAL

Trend Display of 1 - 20 kHz

-- Baseline -- Value 2937 Date 12-MAR-03

0 200 400 600 800 1000

0

1

2

3

4

5

6

7

8

Days 09-JAN-03 To 06-JUN-05

RM

S A

ccel

erat

ion

in G

-s

Date Time Ampl

06-JUN-05 210415 6656

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damageOuter ring (large transmission)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-JAN-05 140435 551

Observing powerful increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Input shaft motor side Input shaft drive side

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Trend Display of 2 - 20 kHz

-- Baseline -- Value 00000 Date 28-MAY-98

0 200 400 600 800 1000 1200

0

05

10

15

20

25

30

35

40

Days 09-JAN-02 To 03-JAN-05

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

09-JAN-02 110324 340

Bearing damageOuter ring (large transmission)

Points of observed Points of observed damages on same type of damages on same type of bearingbearing

Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particular lar shaft our conclusion is that there is a bearing damage shaft our conclusion is that there is a bearing damage

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damage on inner race motor sideBearing damage on inner race motor side

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damage on inner race drive sideBearing damage on inner race drive side

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Bearing damageOuter ring (thrust bearing)

Observing increasement in the area 1-20 kHz (which represents the are of bearing noise) This supports the assumption of a bearing damage under development

Route Spectrum 03-NOV-3 1437

OVRALL= 1824 V-DG RMS = 230 LOAD = 1000 RPM = 1500 RPS = 2500

0 400 800 1200 1600 2000

0

03

06

09

12

15

18

21

24

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

24750 9900 1047

gtSKF NU1026 E=BPFO

E E E E E E E ETrend Display of 1 - 20 kHz

-- Baseline -- Value 00000 Date 16-JUL-96

0 100 200 300 400 500 600 700

0

2

4

6

8

10

12

Days 22-JAN-2 To 03-NOV-3

RM

S A

ccel

erat

ion

in G

-s

ALERT

FAULT

Date Time Ampl

03-NOV-3 143754 9625

Observing frequencies that matches ball pass frequencies outer race (fault frequencies BPFO) on bearing SKF NU1026

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Gear damageInput crown wheel

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage Possible broken tooth

Observing harmonic rpm frequencies on the input shaft of this gear

Route Spectrum 03-FEB-04 143703

OVRALL= 331 V-DG RMS = 4406 LOAD = 1000 RPM = 278 RPS = 463

0 100 200 300 400 500 600 700

0

005

010

015

020

025

030

035

040

Frequency in Hz

RMS

Acc

eler

atio

n in

G-s

Freq Ordr Spec

2519 5437 02161 Time in mSecs

Acce

lera

tion

in G

-s

0 40 80 120 160 200 240

PlotSpan

-4

4

29-NOV-02 133402

12-JUN-03 120411

12-SEP-03 114913

07-OCT-03 130916

08-JAN-04 122213

03-FEB-04 142602

Time Ampl

3215 -906

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Gear damageIntermediate shaft

Waveform Display 07-OCT-3 1316

RMS = 1089 LOAD = 1000 RPM = 296 RPS = 494

PK(+) = 7243 PK(-) = 8067 CRESTF= 741

0 100 200 300 400 500 600

-10

-08

-06

-04

-02

-00

02

04

06

0810

Time in mSecs

Acc

eler

atio

n in

G-s

Time Ampl Dtim Freq

24057 559 19561 5112

Route Spectrum 07-OCT-3 1320 (Demod- HP 500 Hz)

OVRALL= 0701 A-DG RMS = 0700 LOAD = 1000 RPM = 76 RPS = 127

0 5 10 15 20 25 30 35 40 45 50

0

001

002

003

004

Frequency in Hz

RM

S A

ccel

erat

ion

in G

-s

Freq Ordr Spec

5094 3998 02843

Time-waveform indicates that there is a pulsation on time per revolution This supports the assumption of a gear damage

Demodulated measurement shows that there is a harmonic frequency of 5094 Hz 5094 Hz x 60 Hz = ~300 RPM which is close to the intermediate shaft speed Therefore it is likely to believe that there is a tooth damage on this shaft

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Resonance problemCaseOn two main gears several tie-anchor bolts for the pinion bearings on the first gear step broken just after a couple of hundred hours and therefore Maskindynamikk AS was engaged to identify and analyze the vibration in these two gears It was soon discovered to be abnormally high levels of vibration in a specific speed-load area around these bolts (close to maximum speed) and these vibrations were amplified by the gearmesh frequencies of the input shaft This was the first observation that pointed in the direction of a possible resonance ndashproblem

Additional examination was therefore carried out to identity this resonance-problemAn element analysis was carried out to sort which of the gear components had natural frequencies in this frequency range (resonant area) This was not a easy case as more than one component could be involved in thisThru this investigation it was revealed that the bolts had radial natural frequencies which were amplified (excited) by 1st level gearmesh frequency

The resolution to the problem was therefore divided in two First stage involved redesigning and replacing the bolts with others with lower natural frequencies and thereafter to change the propeller curve so that we achieve a lower maximum

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

rpm and a lower maximum gearmesh In addition to this we also achieved to obtain the power by increasing the pitch curve

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 004828

RMS = 838 LOAD = 730 RPM = 1050 RPS = 1750

0 400 800 1200 1600

0

2

4

6

8

10

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

71690 4097 6594

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 010027

RMS = 2305 LOAD = 800 RPM = 1080 RPS = 1800

0 400 800 1200 1600

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

73577 4088 1337

The two engines is running at 1060 rpm which gives a gearmesh of 718 Hz with a amplitude of 67 mms This is normal

The two engines is running at 1080 rpm which gives a gearmesh of 736 Hz with a amplitude of 134 mms An 25 increasement on the gearmesh frequency doubles the amplitude and this clearly indicates a resonance problem

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 08-SEP-07 014106

RMS = 2924 LOAD = 860 RPM = 1100 RPS = 1833

0 400 800 1200 1600 2000

0

3

6

9

12

15

18

21

24

27

30

33

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

Freq Ordr Spec

75353 4110 2638

BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT

Analyze Spectrum 16-SEP-07 100408

RMS = 259 LOAD = 150 RPM = 600 RPS = 1000

0 200 400 600 800 1000 1200

0

01

02

03

04

05

06

07

Frequency in Hz

RM

S V

eloc

ity in

mm

Sec

722

55

836

46

Freq Ordr Spec

83700 8370 220

The two engines is running at 1100 rpm which gives a gearmesh of 753 Hz with a amplitude of 264 mms An 58 increasement on the gearmesh frequency increases the amplitude four times and this definitely indicates a resonance problem

The two engines is running at low and variable rpm with 1st order gearmesh around 350-400 Hz This gives a 2nd order gearmesh frequency in the are 700-850 Hz Also the 2nd order is strongly amplified something which confirms our assumption This proves that there is a resonance problem in this area (700-800Hz)

The measurement technique which were used her is called ldquorpm sweeping with peak-hold functionrdquo which means that you sweep a frequency area to map possible resonance problems

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Unbalanced flexible coupling

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

bull Initial vibration analysis revealed mechanical unbalance in the coupling

bull Unbalance is indicated by a dominating 1st order frequency amplitude

bull Unbalance can have different reasons ndash Insuficcient dynamic balancing ndash Coupling damages as here where the stress

between the rubber elements and the inner ring (steel) has excedeeded the force limits and the rubber elements were damaged after only a few months

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Generator with a unbalanceddamaged coupling elements

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

1orden

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

The outer steel ring of the coupling was turned 180 degrees vs the rubber elements - wich in this case was the rebalancing trick to reduce

the 1st order vibration levels from 18 to 4 mms

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018

Before vs after dynamic balancing ndash reduced 1st order035 - GENERATOR 2

Gen 2 -P05 GENERATOR DEVERTIKAL Route Spectrum 20-SEP-07 152028

OVRALL= 2212 V-DG RMS = 2032 LOAD = 1000 RPM = 1800 RPS = 3000

0 40 80 120 160 200

0

2

4

6

8

10

12

14

16

18

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 1310

035 - GENERATOR 2Gen 2 -P05 GENERATOR DEVERTIKAL

Route Spectrum 28-SEP-07 105416

OVRALL= 1082 V-DG RMS = 1049 LOAD = 1000 RPM = 1801 RPS = 3001

0 40 80 120 160 200

0

2

4

6

8

10

Frequency in Hz

RM

S Ve

loci

ty in

mm

Sec

Freq Ordr Spec

3000 1000 4018