EPRI SYMPOSIUM ON POWER PLANT FANS: THE STATE OF THE ART

CONTROLLING FAN VIBRATION - CASE HISTORIES

Donald R . Smith Senior P ro jec t Leader

Engineering Dynamics Inc. San Antonio, Texas

and

J. C . Wachel Manager of Engineering

Engineering Dynamics San Antonio, Texas

Sponsored by

ELECTRIC POWER RESEARCH INSTITUTE

October 14-18, 1981 Hyat t Regency I n d i anapol i s

I n d i anapol i s, I n d i ana

Key Words : V i b r a t i o n Foundation Natural Frequency Dynami c Soi l Proper t i es Mu1 t i - p 1 ane Balancing La te ra l C r i t i c a l Speed Disc Resonance In1 e t Vortex Rota t ing S t a l l

CONTROLLING FAN VIBRATION - CASE HISTORIES

ABSTRACT Dinald R. S m i t h and J . C. Wachel

F a n vibration problems have been a serious cause of plant unreliability in large f ossi 1 -fired power plants and have resulted in operational problems, shutdowns, and reduced generation. The basic causes of most problems are dynamic resonances asso- ciated with the system. These have t o be identified before practical and effective recomnendations can be made for corrective action or design modifications. The most effective solutions can best be determined from computer models which match the measured f ?el d data.

T h i s paper discusses several case h i stories and i l l ustrates methods and i nstrumenta- tion for analyzing existing fan problems. Experimental techniques are described for defining problem symptoms, and these are related t o root causes by the use of data analysis and canputer simulation techniques. Computer modeling techniques can then be used to evolve reliable fixes or used in the design stage t o simulate interaction between rotor dynamic response and complete system response t o prevent potential dynamic problems.

Section 1

I NTRODKTI ON

Donald R. Smithy

J . C. Wachel

Mu l t i p l e fan systems are t y p i c a l l y complex and v ib ra t ion problems can be d i f f i c u l t t o def ine due t o dynamic acoustical and mechanical in te rac t ion between fans [I]. When fans are i n s t a l l e d i n mu l t i s to r ied buildings, the problems are fu r ther compl i- cated due t o structure-borne vibrations. Sanetimes it appears t ha t "everything shakes," inc lud ing the structure, f loor , fans, ducts, and motors. Plant personnel are o f ten f rus t ra ted i n attempting t o corre la te the countless number o f var iables i n the system t o determine the exact cause o f increases i n v ib ra t ion levels. Often the v i b ra t i on increases are correlated t o seemingly unrelated var iables such as t ime o f day, r a i n f a l l , o r load changes on other units.

I n the analysis o f severe fan v ib ra t ion problems, it i s o f ten found t ha t the high v ibrat ions are due t o coincidences o f one o r more natural frequencies which amp1 i f y low leve l exci tat ions. A l i s t o f possible exc i ta t ion sources and natural frequen- c ies general ly found i n fan systens i s presented below:

EXCITATION SOURCES

Mechanical Defects: m i sal igrment, improper tie-down, bad grout, bent shaft, shaf t rub, warped t h rus t c o l l ar, 1 oose r i v e t s on wheel, bad shrink f i t between wheel and hub, etc.

Unbalance: dust o r ash bui ldup on blades or ins ide a i r f o i l blades, blade erosion, thermal d i s t o r t i o n (sha f t bow, wheel warpage), inadequate balancing a t low speeds

Pul sation: f l ow exc i t a t i on across obstructions, vortex shedding, i n l e t box vortex, r o t a t i n g s t a l l , blade passage

NATURAL FREQUENCIES

shaft lateral cri t ical speed torsional critical speed

0 pedestal , foundation, fl oor/support structure di sc wobble (wheel rocking) acoustical

SYSTEM ANALYSlS

The f i r s t step in defining and solving vibration problems i s t o determine i f the vibrations are due t o ( 1 ) high level excitations, or ( 2 ) low level excitations which are ampl ified by the coincidence of one or more natural frequencies. I f the excita- tions are high, the solution i s generally t o reduce the energy level. If coinci- dences of resonance exist, the solution i s t o modify the system t o change the nat- ural frequencies away f r m the excitation frequency. The most diff icult vibration problems t o analyze and solve are those which have several resonances at the same frequency which increases the cross coupling and interaction between the resonances.

Detailed f ield tes ts coup1 ed with analytical analyses are generally required to sep- arate and identify the excitation sources and their ampl itudes and the system nat- ural frequencies. The analysis techniques and field instrumentation have been greatly improved in the l as t few years. Equipment which used t o be considered for laboratory use only can now be easily transported to the f ield for on s i t e da t a analysis. A typical field instrumentation setup i s shown in Figure 1-1 which includes a minicomputer, real time analyzer, trim balance analyzer, mu1 ti-channel FM recorder, oscill oscopes, signal conditioning ampl i f i e r s and i ntegrators, and X-Y p l otters. In cunpl ex problems mi ni cmputers are used to gather and process several data channels simultaneously and plot the vibration mode shapes. Mu1 ti-channel telemetry systems can be used to collect data such as strain and temperature on the shaft and blades during on-1 ine conditions.

The objective of this paper i s t o i l lus t ra te techniques to identify and solve fan . vibration problems; however, these techniques are also routinely appl ied t o solve vibration problems on other types of rotating and reciprocating equipment. Sane of the mohe unique fan vibration problems analyzed by SwRI have been selected for th is paper t o i l lus t ra te the diagnostic procedures and solution development.

S e c t i o n 2

CASE HISTORIES.

CASE NO. 1

Dur ing t h e i n i t i a l s t a r t u p o f t h e ID f a n s a t a l a r g e power p l a n t , t h e f a n and moto r b e a r i n g hous ing v i b r a t i o n s were g r e a t e r t h a n t h e s p e c i f i e d maximum a1 1 owable 1 eve1 o f 2 mils peak-peak a t t h e runn ing speed o f 900 rpm,

The f o 1 1 owi ng symptoms were obse rved :

Dur ing t h e f i e l d b a l a n c i n g by t h e f a n m a n u f a c t u r e r , t h e f a n s were s e n s i t i v e t o sma l l ba l a n c e we igh t s .

The v i b r a t i o n s were g r e a t e r on t h e motor t h a n t h e f a n .

When t h e motor was run uncoupled , t h e moto r v i b r a t i o n s were l e s s t h a n 1 mil b u t t h e f a n o u t b o a r d b e a r i n g hous ing v i b r a t i o n s were g r e a t e r t h a n 3 m i l s .

The h o r i z o n t a l v i b r a t i o n s a t t h e t o p o f t h e f o u n d a t i o n were a l s o above t h e d e s i r e d limit.

The f a n v i b r a t i o n s were s e n s i t i v e t o t h e i n l e t g u i d e vane s e t t i n g .

The v i b r a t i o n s o f each f a n were a f f e c t e d by t h e v i b r a t i o n s o f t h e a d j a c e n t f a n s even though t h e y were n o t i n s t a l l e d on a common f o u n d a t i o n mat.

Tempera tu re changes i n t h e f l u e g a s had a s i g n i f i c a n t effect on t h e v i b r a t i o n s .

The v i b r a t i o n s on one o f t h e f a n s i n c r e a s e d a f t e r a heavy r a i n .

n e t a i l ed i n v e s t i g a t i o n s r e v e a l e d t h a t t h e f a n s were o v e r l y s e n s i t i v e t o sma l l c h a n g e s i n unba lance , l o a d , and t e m p e r a t u r e due t o amp1 i f i c a t i o n from o p e r a t i n g n e a r two r e sonances . The first f o u n d a t i o n n a t u r a l f r e q u e n c y and t h e f a n s h a f t i n s t a l l e d r e s o n a n t speed were b o t h n e a r t h e f a n r u n n i n g speed. The n a t u r a l f r e q u e n c i e s c o u l d n o t b e a d e q u a t e l y s e p a r a t e d w i t h o u t m a j o r m o d i f i c a t i o n s ; t h e r e f o r e , t h e v i b r a t i o n 1 eve1 s were reduced by h o t b a l a n c i n g e a c h f a n .

FOUNDATION NATURAL FREQUENCY

V i b r a t i o n l e v e l s measured on t h e f a n b e a r i n g hous ings d u r i n g s t a r t u p s and coas tdowns r e v e a l e d a n a t u r a l f r e q u e n c y j u s t below t h e runn ing speed ( F i g u r e 2- 1) . The v i b r a- t i o n mode s h a p e was o b t a i n e d by measu r ing t h e v i b r a t i o n s on t h e b e a r i n g hous ing , . b e a r i n g p e d e s t a l , and s e v e r a l l o c a t i o n s on t h e s i d e o f t h e c o n c r e t e f o u n d a t i o n . V i b r a t i o n d a t a were p l o t t e d on a s c a l e d drawing ( F i g u r e 2-2). The e n t i r e f o u n d a t i o n

Figure 1-1. Typical F i e l d Equipment Used i n Vibrat ion Analys i s

and bearings moved together as a r i g i d body and rocked about a po in t several f e e t below the foundation which i s c h a r a c t e r i s t i c o f a foundation resonance.

FREQUENCY, Hz

Figure 2-1. Spectral Analysis o f Bearing Housing V ibra t ions During '

Star tup

FOUNDATION ROCKING MODE (8y)

Figure 2-2. Fan Foundation V ib ra t ion Mode Shape

On three o f the fans, the foundation natura l frequency was j u s t below the running speed, bu t on the f o u r t h fan, the resonance was s l i g h t l y above the running speed. To v e r i f y t h a t the measured responses were foundation resonances, a va r iab le speed, un id i rec t iona l , mechanical shaker was attached t o the foundation and run through a

speed range of O t o 1800 rpm. The shaker- exci ted v ib r a t i on response v e r i f i e d t h a t t h e f i r s t foundat ion na tura l f requency ranged f run 12 t o 16 Hz f o r t h e s e fou r "iden- t i c a l " fans . The fan with t h e foundat ion resonance a t 16 Hz was t h e most s e n s i t i v e t o small changes and had t h e g r e a t e s t ampl i f ica t ion a t t h e running speed of 900 rpm (15 Hz).

These foundat ions were designed t o have t h e i r lowest foundat ion na tura l frequency a t approximately 1500 cpm (70% above running speed). Foundation ca lcu l a t i o n s were made i n t h e design s t age using a va lue of Young's Modulus ( E ) o f 500,000 ps i f o r t h e sandstone beneath t h e foundat ions.

To eva lua t e t h e discrepancy between t h e measured and ca l cul a t ed foundat ion na tura l f r equenc i e s , t h e foundat i on- soil dynamic system was modeled on a computer program developed by SwRI. Foundation na tura l f requenc ies were ca l cu l a t ed f o r a range of s o i l modul i and compared wi th t h e shaker data . As shown i n Figure 2-3, an e f f e c t i v e s o i l modulus of 80,000 t o 120,000 ps i was required t o match t h e measured foundat ion na tura l f requencies .

gN FOUNDATION

ELASTIC SOIL MODULUS (El, PSI X lC3

Figure 2-3. Comparison of Measured and Calculated Foundation Natural Frequencies

These effective E values as empirically defined are much lower than the typical values given for sandstone which can range from 500,000 t o 3,000,000 psi. The orig- inal soil analysis report revealed that the soil was highly strat if ied, and the depth t o each layer varied considerably over short distances. The soil modulus beneath the foundations could also have been reduced due to blasting and over- excavation when the foundations were constructed.

This case history i l lustrates that the major unknown in predicting the resonant fre- quency of a fan-foundation system i s the effective soil modulus. In the design of these fans, a minimum value for the soil modulus was obtained f r m soil bore tes ts , b u t even th is m i n i m u m value was approximately five times too large.

The field data indicated that fan vibration levels were also affected by moisture content of the soil because the vibrations would change significantly after a rain. Moisture changes the soil modulus and can shif t the foundation natural frequency closer to the running speed which amp1 i f ies the vibrations. Also, the damping can be reduced due to the cohesion of the soil and separation from the concrete caused by vibration.

Since the foundation dynamic design i s highly dependent upon the E value of the soi l , every effort should be made in the design stage t o obtain effective dynamic soil modulus directly below the foundation. There are several methods which can be used to obtain more accurate dynamic soil data, as given below:

1. A 1 arge o u t p u t , 1 ow frequency shaker can be used to excite Ray1 eigh ground waves a t the foundation elevation. The soi 1 modulus can be calculated f r m th i s data [2]. The effective soil modulus should be obtained after the s i t e has been excavated and the f i l l added.

2. The soil modulus can also be obtained by using cross hole tests. Two holes are bored to the bottom of the foundation [ Z ] . A shear wave i s created by an impact at the bottom of the hole and measured w i t h a transducer i n the other hole which i s a known distance away. This method could a1 so be used to determine the effective modulus a t the bottom of piles.

3. A shaker can also be used t o excite the natural frequencies on a small or partial ly compl eted foundation. The partial ly compl eted foundation can then be modeled on a computer and the effective soil modulus can be determined by comparing the calculated natural fre- quencies w i t h the measured frequencies.

A shaker was used on a partially completed foundation block of an adjacent u n i t (Figure 2-4) to determine the effective soil modulus. The t es t s indicated that the effective soil modulus would cause the foundation natural frequency t o be near the fan running speed. Foundation modifications t o reduce the vibrations were analyzed using the computer program (e.g., increase mat size, piles, added mass, etc) . Most of the foundation modi flcations i nvestf gated were i neffectf ve, were not cost effec- tive or were space limited. Other methods were then investigated t o reduce the vibrations.

I S 10 IS 20 25 30

SHAKER SPEEq He

Figure 2-4. Comparison of Measured and Ca1 cul a ted Foundation Vibrat ion Response

Shaker t e s t s on one foundat ion revea led t h a t t h e v i b r a t i o n s were increased by a f a c t o r of s i x when d i r t was removed from t h e s i d e of t h e foundation. Based upon t h i s t e s t , sand bags were temporar i ly placed aga in s t t h e s i d e of t h e foundat ion t o i nc r ea se t h e damping and l a t e r a l r e s t r a i n t . Vibra t ions were reduced by a f a c t o r o f approximately two t o one.. These t e s t s i nd i ca t ed t h e s i g n i f i c a n t e f f e c t of s o i l damping i n t h e o i e r a l l design of foundat ions. Foundation v i b r a t i o n s can a l s o be reduced i n some cases by shor ten ing t h e he ight of t h e foundat ion above t h e ground o r des ign ing ' the foundation i n t h e shape of a t runca ted pyramid ins tead of a t a l l rec- tangul a r block a s i s usual 1y done.

FAN SHAFT LATERAL RITICAL SPEED 2 The f an s h a f t l a t e r a l na tura l frequency was measured t o be 960 cpm using a v a r i a b l e speed shaker a t tached t o t h e foundation. For t h e s e t e s t s a c c e l e r m e t e r s were a t t a ched t o t h e fan s h a f t and proximity probes were mounted a t t h e bearings t o sepa- r a t e t h e r o t o r response from t h e foundat ion response. The fan i n s t a l l e d resonant speeds were a l s o ca l cu l a t ed a s shown i n t h e c r i t i c a l speed map (F igure 2-5) where t h e s h a f t natural f requenc ies a r e p l o t t e d versus t h e e f f e c t i v e support s t i f f n e s s . The e f f e c t i v e support s t i f f n e s s inc ludes t h e s t i f f n e s s e s of a l l t h e sp r ings from t h e r o t o r t o ground including t h e o i l f i l m , bear ing pedes t a l , foundat ion, and s o i l .

STIFFNESS (LWIN)

Figure 2-5. ID Fan Critical Speed Map

The fan manufacturer had calculated the cri t ical speed t o be 1180 rpm; however, the calculations were based on a "rigid bearing cri t ical" . For many years fan manufac- turers used this calculation which assumes the rotor i s mounted on rigid supports because many of the supporting stiffness values were not known. As a result, there has been some confusion in the fan industry as t o the exact definition of cri t ical speed and resonant speed. To clarify the problem, the Air Movement and Control Association (AMCA) C31 adopted the following definitions:

Critical S eed: A cri t ical speed i s that speed which corresponds t o Ti--- t e natural frequency of the rotating element (impeller and shaft assembly) when mounted on rigid supports. (Note: This i s generally referred to as the rigid-bearing cri t ical speed.).

Design Resonant Speed: Design resonant speed i s that speed which corresponds to the natural frequency of the combined spring-mass system of the rotating element, oil film, bearing housing, and bearing supports b u t excluding the foundation (foundation stiffness i s considered as infini te) .

Installed Resonant Speed: Installed resonant speed i s that speed which corresponds t o the natural frequency of the combined spring- mass system of the rotating element, oil film, bearing housing, bearing supports, and includes the effect of foundation stiffness.

The calculated rigid-bearing cri t ical was 1180 rpm and the calculated instal 1 ed res- onant speed was 960 rpm. Normal ly the instal led resonant speed should be a t least 20% from the running speed t o prevent excessive vibration ampl ification, In th is case the calculated installed resonant speed was only 7% above the running speed; therefore, amp1 i ficat i on could be expected. The rigid-beari ng critical speed i s not applicable because in the real world the rotor i s no t rigidly supported. The actual stiffness of the concrete and pedestal for this systen was approximately 5 x 10

6

Ib/in which i s much less t h a n rigid.

A forced vibration response analysis i s normal ly required to accurately determine the rotor installed resonant speed. The rotor should be modeled using the oil film and the combined spring-mass system of the bearing housing, bearing supports, and foundation. The bearing oil film for journal bearings i s usually represented by horizontal , vertical , and cross coupl ing stiffness and damping terms. The cross coupl i ng terms significantly affect the cal cul ated rotor response, particul arly i f the horizontal stiffness i s much less than the vertical stiffness. in these cases the rotor instal led resonant speeds cannot be accurately calculated if the oil film i s represented as a single stiffness and damping value and the effects of the cross coupl ing terms are omitted.

Therefore, i t i s important that fan users be aware of the cri t ical speed definitions used in the fan industry and refer to the instal led resonant speed when writing design specifications.

For th is systen major modifications such as shortening the bearing span would have been required t o increase the installed resonant speed t o 2045 above the running speed. Since these modifications were no t practical, i t was decided t o reduce the shaft vibrations by improving the fan balance.

.-

SYSTEM BALANCE

I t was desired t o reduce the vibrations on the fan and motor bearing housings by adding balance weights only i n the fan. This means that the vibration levels on a11 four bearing housings must be considered when calculating the balance weights.

To detennirie i f the vibration ampl itudes were stable and repeatable, the vibration amp1 itudes were plotted versus the phase angles as shown in Figure 2-6. The varia- tion in amplitude and phase angle for each bearing created elliptical patterns due to the interaction from adjacent fans. The best balance was obtained by reading the vibration amplitude and phase angle a t the center of each el l ipt ical pattern.

On one fan a particularly strong interaction from an adjacent fan was observed and a stable vibration phase angle could n o t be obtained. When the adjacent fan was

.

tripped, the fan vibrations immediately stabilized (Figure 2-71. To further docu- ment th is interaction, the fan vibrations were plotted with the key phaser installed on the adjacent fan which showed that the fan vibrations were more closely related t o the adjacent fan.

PHASE ANGLE, DEGREES PHASE ANGLE, DEGREES

Figure 2-6. Var ia t ion of Motor and Fan F igure 2-7. Fan Vibra t ion I n t e r a c t i o n Bearing Vibra t ion During wi th Adjacent Fan Hot and Col d Condit ions

Even wi th t h e i n t e r a c t i o n between t h e f a n s and t h e amp l i f i c a t i on due t o t h e founda- t i o n na tura l frequency and i n s t a l l e d resonant speed, t h e v i b r a t i o n s on t h e f a n s and motors were reduced below 2 mi l s co ld using a two-plane balance (F igu re 2-6). In t h e balancing procedure, a l e a s t squares balancing technique was used which inc ludes t h e v i b r a t i o n s a t a l l f o u r bear ing l o c a t i o n s f o r t h e two balance pl anes t h a t were a v a i l a b l e .

HOT BALANCE

When t h e b o i l e r was f i r e d during t h e i n i t i a l steam blows and t h e tempera ture of t h e exhaus t gases i n t h e ID f an reached 250' F, t h e v i b r a t i o n s on t h e bear ing housings immediately increased and t h e phase ang l e s changed (F igure 2-6). The ba lance of a l a r g e f a n normally changes when t h e f a n is heated due t o thermal d i s t o r t i o n ; how- e v e r , i f t h e fan i n s t a l l e d resonant speed and foundat ion na tura l frequency a r e nea r t h e running speed, t h e small changes i n unbalance can r e s u l t i n s i g n i f i c a n t i n c r e a s e s i n v i b r a t i o n amp1 i tudes.

I t was t h e r e f o r e recommended t h a t t h e system be "hot balanced" t o reduce t h e bear ing housing v i b r a t i o n s t o meet t h e requi red v ib ra t i on 1 imi t s . Hot balancing f i n s with t h e p l a n t a t f u l l load i s a d i f f i c u l t , t ime consuming process . One of t h e major problems was t h a t t h e f a n s d id no t have t u rn ing gea r s and t h e r o t o r s bowed due t o thermal s t r a t i f i c a t i o n of t h e gases i n t h e f a n housing. When t h e f a n s were r e s t a r t e d with a bowed s h a f t , t h e v i b r a t i o n s were excess ive and t h e f a n s were imme- d i a t e l y t r i pped , Each s t a r t u p and coastdown reduced t h e bow due t o more uniform hea t ing of t h e r o t o r , Sometimes a f t e r two o r t h r e e s t a r t s and coastdowns, t h e v i b r a t i o n s would be reduced t o l e v e l s where t h e fan could be operated.

The s h a f t thermal bow problem was l a t e r solved by i n s t a l l i n g a temporary t u r n i n g g e a r t o r o t a t e t h e s h a f t when t h e f an was off . I t i s highly recommended t h a t

turning gears be included i n the original design on all hot fans because i t i s dif- f i cul t and expensive t o re t rof i t turni ng gears.

I t took approximately 10-12 hours of running time before the fan thermally stabi- lized and the vibration and phase angle quit changing. Balance data taken before this time could result in inaccurate balance weights.

The fan and motor vibrations were reduced by hot balancing; however, the fans remained very sensitive t o small changes. The balance weights used were much less than the allowable residual balance as specified for fans of th is type [4].

The bearing housing vibrations could not be maintained below the specified level of 2 mils; however, there was no bearing damage because the foundation moved with the shaft. I t was found that the bearing housing vibration levels could be increased to 5 mils peak-peak without causing excessive shaft-to-housing vibrations. The plant operated for several years w i t h these vibration I evels without bearing damage. This i 1 lustrates that a1 1 owable vibration levels must be determined for each individual piece of equipment and published vibration cri teria should be used as a guide or reference.

CASE NO. 2

Two boilers were converted from forced draft systems t o balanced draft systems w i t h the additfon of two large ID fans on each unit. The ID fans were originally instal led with louver controls which were later changed to variable inlet vanes to eliminate high pulsations i n the ducts. The variable inlet vanes apparently reduced the pulsations in the fan discharge; however, the fan and motor bearing vibration increased t o unacceptable levels when the units were run above half load.

The fol 1 owing symptoms were observed:

1. After a few days, the fan bearing housing vibrations would increase i n ampl itude from 2 mils t o over 5 mils and the phase angle would shi f t over 180 degrees.

2. ifot balancing the fans was not effective in reducing the vibrations for a 1 ong time period.

3. The groundborne interaction between the fans was excessive and only one fan could operate a t a time.

I t was determined that the fan wheel disc wobble natural frequency was slightly above the running speed which ampl ified low level excitations and resulted i n exces- sive vibrations. The problem was corrected by raising the disc wobble natural fre- quency further above the running speed by adding a stiffener plate to the center plate.

DISC WOBBLE

The bearing housing vibrations and shaft vibrations were recorded during several startups and shutdowns (Figure 2-8) t o determine i f the fan was operating near a mechanical natural frequency. I t i s known that vibrations due to unbalance alone (discounting resonances) wi 11 increase as a function of the speed squared. However, the vibration at 900 rpm was significantly greater than the level due t o a pure unbalance, which indicated that the fan was operating near a natural frequency. The 1 ack of a 180 degree phase shi f t on the bearing housing vibrations indicated that t h i s resonance was probably not a shaft cri t ical speed.

Figure 2-8. Bearing Housing Vibration Response During Startup

Pulsations measured in the discharge duct were synchronous and phase coherent with shaft speed. The pulsation amplitudes were approximately 0.7 psi a t the fan running speed and produced a force equivalent t o a 10 oz unbalance at the outside radius of the fan. The fans were apparently sensitive to small pulsations or aerodynamic excitations because changing the shaft dust covers had a significant effect on the fan vibration characteristics. I t i s diff icult t o reduce low 1 eve1 pul sations; therefore, i t was fe l t that the long range solution would be t o identify and modify the mechanical natural frequency which was ampl ifying the vibrations.

To identify the resonant frequency, a large variable speed shaker was bolted t o the concrete foundation with the shaking force applied in the horizontal direction and run through a frequency range from well below the fan running speed t o 50% above. Vibrations were measured on the foundation, bearing housings, shaft, and fan wheel to identify the resonance just above the running speed.

The foundation response was found t o be highly damped and offered no significant ampl ification of the vibrations a t the running speed. The major effect of the foun- dation was as a vibration transmission path between adjacent fans.

The fan sha f t na tura l frequency was measured t o be 1200 cpm w i t h t h e shaft r e s t i n g on t h e bearings. Ca lcu la t ions i nd i ca ted t h a t t h e i n s t a l l e d resonant speed when the sha f t was running on t h e o i l f i l m would be lowered t o 1080 cprn which was s t i l l ade- quate ly above t h e running speed.

V ib ra t i ons measured on t h e fan wheel i nd i ca ted a d i sc resonance a t 930 cpm (Figure 2-9). There was a c lass i c 180 degree phase s h i f t i n t h e v ib ra t i ons as the shaker speed increased above the resonant frequency. The v i b r a t i o n s were measured a t sev- e ra l l o c a t i o n s on t h e wheel t o determine t h e v i b r a t i o n mode shape.

I DISC WOBBLE MODE SHAPE

SHAKER SPEED, CPM

Figure 2-9. Disc-Wobble Mode V i b r a t i o n . Response t o Shaker

E x c i t a t i o n

The shaker t e s t i l l u s t r a t e d t h a t t he a x i a l wheel v i b ra t i ons a re e a s i l y coupled i n t o ho r i zon ta l v i b r a t i o n s on the bear ing housing because t h e wheel resonance was exc i ted by a ho r i zon ta l shaking f o r c e on t h e foundation.

It was f e l t t h a t t h e bearing housing v i b r a t i o n amp1 i f i c a t i o n a t running speed could be reduced by r a i s i n g the d i sc resonant frequency. To evaluate t h e e f f e c t o f s t i f f - ening t h e center p la te , wedges were d r i v e n between t h e fan impe l l e r and t h e i n l e t cone t o r e s t r a i n t he wheel. When t h e shaker t e s t was repeated, t h e response a t 930 cpm was e l im inated on t h e fan wheel and t h e bearing housings.

An evaluation was made of a plate stiffener that could be fabricated and bolted t o the center plate in the field to raise the disc resonant mode. Since i t i s d i f f i- cult to mathematical 1y model a complex bolted joint, a small 1/10 scale model was fabricated and tested. The model t e s t confirmed that the center plate stiffness could be adequately increased by bolting a reinforcement plate to the center plate.

After a stiffener plate was installed in one fan, a shaker tes t was performed which showed that the wheel natural frequency increased 25%. When the fan was restarted, the bearing housing vibrations were significantly reduced as shown in Figure 2-8. The other fans were then similarly modified and have been in operation for several years and are no longer sensitive t o low level excitations even a t full load.

CASE NO. 3

During the ini t ial startup of two ID fans a t a power plant, high vibration levels were observed on the fan inlet housing which resulted in structural damage requiring extensive repairs. One of the inlet cones was broken away from the common sheet and several of the pipe braces which extend.between the common sheet and the side sheets were a1 so broken. The unit had been in operation for only 11 days during the ini- t i a l firing and steam blows of the boiler. The second fan had run for only 4 days and i t t oo had suffered extensive damage similar t o the f i r s t fan.

The following symptoms were observed:

1. The vibrations on the fan inlet housing were excessive and increased as a function of the inlet damper position.

2. The vibrations on the fan bearing housing (pedestals) and foundation were low and were not affected by the inlet housing vibrations.

The problem was due t o an inlet vortex which created a rotating s ta l l condition and generated high pulsations a t multiples of 20 Hz. The pulsations a t 40 Hz matched the mechanical natural frequency of the inlet cone and resulted in high vibrations and fatigue failures of the inlet cone and the internal pipe braces. The inlet vortex was eliminated by adding sp l i t t e r plates in each inlet.

Accelerometers, strain gages, and pressure transducers were instal 1 ed a t several cri t ical locations on the fan housing t o identify the source of excitation. I t was found that the pulsation, vibration, and strain levels increased as a function of the damper position and the levels were excessive a t a damper position of approxi- mately 70% open. Strain levels of 350 in/in x peak-peak were measured which were higher than the allowable for th is material and configuration.

The predominant pulsation and strain frequencies were a t multiples of 20 Hz (Figure 2-10). Rotating stal l generally occurs a t multiples of 1/3 shaft speed. This fan ran at 15 Hz and the pulsations at 20 Hz were at 4/3 shaft speed. The s ta l l was due t o marginal flow or improper preswirl conditions a t the fan inlet which caused the a i r to impinge on the fan blade a t a poor angle of attack.

FREQUENCY, HI

Figure 2-10. I n l e t Pu l sa t i on and S t r a i n Spec t ra Showing Mu1 t i pl e s of Rotat ing S t a l l Frequencies

Several modi f ica t ions suggested by t h e manufacturer were made t o t h e duc t s and t u rn ing vanes upstream of t h e f an , but no reduc t ion of pu l s a t i on o r v i b r a t i o n occurred. Test ing was then conducted t o determine t h e e f f e c t of d i scharge ' throt t l e con t ro l . Flow through most f a n s is con t ro l l ed on t h e i n l e t s i d e and o u t l e t dampers a r e used f o r f an i s o l a t i o n . For this test, t h e o u t l e t dampers were temporar i ly s e t up f o r manual con t ro l . The i n l e t dampers were opened t o 45% which r e s u l t e d i n high pu l sa t i ons and v ibra t ions . The i n l e t dampers were l e f t on a u t a n a t i c con t ro l and t h e o u t l e t dampers were p a r t i a l l y c lo sed , keeping t h e flow r a t e constant . When t h e o u t l e t dampers closed t o 5 5 X , t h e inTet opened t o 51% and t h e v i b r a t i o n s and pulsa- t i o n s immediately reduced. Th i s t e s t suggested t h a t a f i x e d o r i f i c e i n t h e d i s- charge duc t could reduce t h e p u l s a t i o n s and v ib ra t i ons , but i t s e f f e c t s a t f u l l load were not known. A proper ly designed damper t o cont ro l d i scharge fl ow was impossi bl e t o ob t a in and i n s t a l l i n a reasonable time period; t h e r e f o r e , t e s t i n g was continued t o o b t a i n another so lu t i on .

I t was f e l t t h a t an i n l e t box vor tex could be c r ea t ed by t h e i n l e t dampers and t h e i n l e t duc t con f igu ra t i on C51. The i n l e t vor tex could cause improper i n l e t flow con- d i t i o n s and r e s u l t i n a r o t a t i n g s t a l l condi t ion . In o r d e r t o de s t roy any vortex forming tendenc ies , s p l i t t e r p l a t e s were i n s t a l l ed i n each i n l e t box (F igure 2-1 1 ) d i r e c t l y oppos i t e t h e i n l e t dampers. The v i b r a t i o n s , pu l s a t i ons , s t r a i n , and no i se l e v e l s were g r e a t l y reduced. In add i t i on , t h e a i r flow a t comparable vane s e t t i n g s was increased by a f a c t o r of 1.7.

SPLITTER PLATE TO BREAKUP

VORTEX

Figure 2-11. Inlet Flow Splitter Used to Reduce Inlet Vortex

Section 3

CONCLUSIONS

The case hi stories discussed i l lus t ra te that fan vibration problems are usually system related and caused by mechanical or acoustical interaction between the fan, motor, foundation, and ducts. Systematic diagnostic fie1 d t e s t procedures and com- puter analyses of individual components are usually needed to define exact causes and optimum solutions. Several problems common to fans were identified and guide- lines for analyzing and solving the problems are summarized below.

FOUNDATIONS

Foundation natural frequencies should be at least 20% away from the fan running speed because they can amp1 ify fan and motor vibrations.

Accurate dynamic soil properties must be known to properly simulate foundat ion-soil systems.

The dynamic soil modul us should be experimentally measured at the s i t e of each foundation after the excavation has been completed and the backfill has been added.

Most foundation vibration problems are due t o a rocking mode about a point below the foundation. Foundations are often designed as t a l l rectangular blocks with the height greater than the width. By increasing the width of the foundation block and mat, the lateral s tabil i ty of the entire system could be improved.

The outboard concrete pedestal i s typically underdesigned for the dynamic radial and axial forces from the fan. The dynamic stiffness should be canparabl e with the inboard pedestal.

SHAFT LATERAL CRITICAL SPEED

1. The fan shaft "installed resonant speed" as defined by the AMCA should be used in the design or specification of a fan system.

2. The installed resonant speed should be a t least 20% removed from the running speed.

3. To accurately calculate the rotor installed resonant speed, a forced vibration response analysis should be performed modeling the bearing oil film stiffness and damping characteristics with the eight coeffi- ci ent representation.

B ALANC I NE

1 . When bal anci ng, extraneous i nf1 uences such as excessive mi sal ignment, fan interaction, etc., need t o be minimized to obtain the proper bal- ance data.

2. For sensitive systems alignment specifications may have t o be tightened.

3. Hot balancing often requires waiting for 10-12 hours for a large fan t o stabilize and q u i t shifting. The fan i s generally stabilized when the rotor stops growing i n the axial direction. Balancing w i t h data taken before thermal stabilization may result i n ineffective correction.

4. The 1 east-squares mu1 t i pl ane bal anci ng method a1 lows mi nimi zi ng vibration a t any number of points. This method has advantages for systems operating near resonance and can minimize the number of bal- anci ng runs which reduces the required downtime.

5. A11 hot fans should have t u r n i n g gears.

WHEEL RESONANCES

The disc wobble resonance should be a t 1 east 20% removed from the running speed. Fans w i t h the disc resonance close t o running speed can be overly sensitive t o low level excitations such as pulsation and unbalances.

The disc axial resonant vibrations can cause amp1 ification of hori- zontal vibrations on the bearing housings.

Disc resonances can be experimentally determined by impact tes ts or shaker tests. All centrifugal fans should be checked during manufac- turi ng to determine di sc resonant frequencies.

The disc wobble resonant frequency in the example was increased by stiffening the center plate. A stf ffening plate could provide a long-term fix for existing fans with th is problem.

ACOUSTIC EXCITATION

The inlet ducts should be designed to prevent forced inlet vortices (spi n-swi rl ) . The i nlet vortex causes improper fl ow conditions and results i n reduced fan performance and possibly rotating s ta l l on the fan blades.

Pulsations generated by rotating s ta l l usually occur a t multiples of 1/3 shaft speed and, when sufficiently 1 arge, can cause structural damage.

Spli t ter plates can be installed i n the fan inlet t o break up the i n l e t vortex.

REFERENCES

1. D. R. Smith and H. R. Simmons. "Unique Fan V ibra t ion Problems: Their Causes and Solutions." Col lege Stat ion, Texas: Proceedings o f the 9 th Turbomachinery Sym- posium. Gas Turbine Laboratories, Texas A M Un ivers i ty , pp. 33-43.

2. S. C. Arya, M. W. O 'Ne i l l , and G. Pincus. Design o f Structures and Foundations f o r V ibra t ing Machines. Houston: Gu1 f Pub1 i shing Company, 1979, pp. 62-64.

3. Power P lant Fans Spec i f i ca t i on Guide1 ines. A i r Movement and Control Association, Pub l ica t ion 801, 1977.

4. "Balance Q u a l i t y o f Rota t ing R ig id Bodies." Acoustfcal Society o f America, ASA sm 2-1975.

5. Fan Appl i c a t i o n Manual, Pa r t 1. A i r Moving and Condit ioning Association, Pub- l i c a t i o n 201, 1976.