MULTI-STAGE CENTRIFUGAL PUMP CONTENTS

Page : 1 of 66

1. MULTI-STAGE CENTRIFUGAL PUMP DESCRIPTION

y Pump Casing y Axial And Radial Split Casing y Interstage Construction

2 AXIAL THRUST IN MULTISTAGE PUMP

y Use Of Double Suction Impeller y Arrangement With Single Suction Impeller y Series Connected Multistage Pump

Page 1 of 72

3 HYDRAULIC BALANCING DEVICES

y Balancing Drums y Balancing Disc y Combination Balancing Disc And Drum

4 MULTISTAGE HORIZONTAL CENTRIFUGAL PUMPS IN PLANT

y Description Of Pump y Double Thrust Bearing y General Dismantling And Reassembling Instructiony Checking Concentricity Of Rotating Elementsy Adjustment Of Pump Rotory Axial Clearances

5. MULTI-STAGE PUMP CONSTRUCTION

y Impellers,Casing Type,Multistage Layout,No.of stage6 MULTI-STAGE PUMP ARRANGEMENT

y Pump arrangement & Hydraulic Balancing Devices7 AXIAL THRUST IN SINGLE STAGE AND MULTI-STAGE PUMP

y Axial thrust8 MULTI-STAGE SHAFT & SHAFT SLEEVES

y Shaft ,Shaft Sleeves9 STUFFING BOXES

y Packing y Mechanical Seal

10 BEARINGS

y Ball bearing,y Roller bearing,y Sleeve bearing y Thrust bearing

11 COUPLINGS

12 PUMP MOUNTING

13 IMPELLER DESIGN

y Single Suction,Double Suction & Multi-stage DesignPage 2 of 72

14 MULTI-STAGE & CENTRIFUGAL PUMP TYPES SELECTION CRITERIA

y Vertical in line single stage pump,Horizontal multi-stage pump,Donut type pump,y Vertical multi-stage pump & Axial thrust in multi-stage pump

15 HYDRAULIC BALANCING DEVICES

y Balancing drum,y Balancing disks y Combination balancing disk & drum.

16 SHAFTS AND SHAFT SLEEVES

17 CRITICAL SPEED

18 RIGID AND FLEXIBLE SHAFT DESIGNS

19 DRAWINGS

Page 3 of 72

MULTISTAGE CENTRIFUGAL PUMP DESCRIPTION

This handout will not deal with the basic fundamentals of centrifugal pump which had been

covered in the basic pumps courses. However those description which are more peculiar to the

multistage horizontal centrifugal pump will be explained.

1.1 Pump Casing

Although most single stage pumps have volute casings, both volute and diffuser casing are used

in multistage pumps. In very high pressure (exceeding 100 bars) applications the diffuser casing

is preferred due to its symmetry and other advantages. It has no complicated castings, permits

more uniform expansion under very high temperature operation. Volute casing develops radial

thrust that sharply increased stage pressure, fig.1. Multistage volute casing design may

incorporate staggered volutes arrangement so that the individual radial thrusts balance out, fig.2.

Fig.1 Radial Thrust In Volute Casing

Fig.2 Staggered Volutes Arrangement

Page 4 of 72

Fig. 2.2 Zero radial reaction in single-volute Fig. 2.3 Radial reaction in a single-

casing volute casing

Fig. 2.4 Magnitude o radial reaction in single- Fig. 2.5 Radial reaction in double-

volute casing. volute pump

F decreases from shut-off to design capacity and the increases with over-capacity. With over-capacity, the reaction is roughly in the opposite direction from that with partial capacity.

Page 5 of 72

1.2 Axial And Radial Split Casing

Both axial and radial split casings are used for multistage centrifugal pumps. The axially split

casing can overcome the radially split casing limitations on accessibility for inspection and

repairs. The casing suction and discharge connection are located in the lower half. The upper

half can be remove for inspection without disturbing the flanges and foundations connection. A

basic weakness of axial split casings, see fig.3. is the difficulty in maintaining a tight horizontal

flange when pressures increase. As most multistage centrifugal pumps have to operate at high

pressure, the trend is towards radial split.

Fig.

3 Axial Split Casing

Radially split casing pumps are used for high working pressure. In earlier design individual stage

sections and separate suction and discharge heads were held together with large through bolts,

fig.4. These pumps had serious dismantling and reassembly problems because suction and

discharge connections had to be separated each time the pump in open.

Page 6 of 72

Fig. 4

Radial Split Casing

The double casing pump retained the advantages of the radially split casing design and solved

the dismantling problems. The basic principle consists in enclosing the working parts of a

multistage centrifugal pump in an inner casing and in building a second casing around this inner

casing. The space between the two casings is maintained at the discharge pressure of the last

stage,

fig.5

and 6.

The

inner

casing

design

follows

either

axial

splitting

fig.5 or

radial splitting fig.6.

Page 7 of 72

Fig. 5

Double

casing

With

Axially

Split

Inner

Casing

The second casing is a solid barrel of cast or forged steel. As the outside of the inner casing is

subject to the pressure greater than the average internal pressure the inner casing is under

compression and the axial or radial joints will remain tight. Further the suction and discharge are

an internal part of the outer casing, thus simplifying maintenance.

Page 8 of 72

Fig. 6 Double

Casing With

Radially Split

Inner Casing

Radially split

inner casing

have better

symmetry and

the interstage joints are of the ring type which is most easily held tight under high pressure as

compared to the axially split interstage leakage.

1.3 Interstage Construction

A multistage pump has adjacent chambers at different pressure. These chambers must be isolated

from one another so that leakage from high to low pressure will occur only at the clearance joints

between stationary and rotating pump parts and this remain minimal.

The isolating wall used to separate two adjacent chambers of a multistage pump is called a stage

piece, a diaphragm, or an interstage diaphragm. The stage pieces, which are usually solid, are

assembled on the rotor along with impellers, sleeves, bearings and similar components. To

prevent the stage pieces from rotating, a locked tongue-and-groove joint is provided in the lower

half of the casing, fig. 7.

Page 9 of 72

Fig. 7 Locked

Tongue-And-

Groove Joint

Stage-piece is

arranged so that

pumping

differential will tend to seat it tightly against the casing fig. 8. Some designs have an elastic seal

ring lodged in the stage-piece face in contact with the casing to insure tightness still further, fig.

9.

Fig. 8 Stage-Piece

Arrangement

Page 10 of 72

Fig. 9 Stage-Piece

Elastic Seal Ring

2. AXIAL THRUST IN MULTISTAGE PUMP

As have described in pump fundamentals there is always an axial thrust set up in a centrifugal

pumps. Several means are provided to compensate this thrust e.g. providing balancing holes, use

of double suction impellers. In multistage pumps where operating pressure are high this thrust

could be tremendous, thus various means are applied to reduce it.

2.1 Use Of Double Suction Impeller

It may seem that the advantages of balance axial thrust and greater suction area of a double

suction impeller (fig. 10) is best for multistage pumps. However it is not so, most multistage

pumps are low capacity pumps, it is seldom necessary to use double suction impellers just to

reduce the NPSH required for a given capacity. Further the use of double suction impeller does

not necessary eliminated the use of a thrust bearing, a certain amount of axial thrust is always

present.

Page 11 of 72

Fig. 10 Thrust On Double Suction Impeller

Most important, use of double suction impeller and needless length to the pump shaft span, thus

increasing casting difficulties and shaft deflection problems. Double suction impellers are

arrange in ascending order, (fig. 11) or practical reasons. These add to sealing problem at the last

stage, which must seal the high final stage pressure. The results, most multistage pumps use

single suction impellers.

Fig. 11

Four

Stages

Pump with

Double Suction Impellers

2.2 Arrangement With Single Suction Impellers

Page 12 of 72

The single suction impellers of the multistage pump shown in fig.12 are discharging the liquid

directly into the next impeller. It is obvious that the shaft of this type of pump is not in axial

balance, so that a special arrangement is required. To obtain the maximum axial balance of the

shaft there are various designs of multistage pumps, although their basic principles are the same.

Fig. 12 Unbalance Axial Thrust Of Multistage Pump

Fig. 13 shows a multistage centrifugal pump where the liquid discharged from the first impeller

flows to the other end of the pump into the suction of the second impeller.

By guiding the liquid in such a way the shaft of the pump remains in axial balance, and no

special balancing arrangement might be installed.

Page 13 of 72

Fig.

13 Arrangement For Axial Balance

Fig. 14 shows a mixed design in which the final stage impeller is mounted in an opposite

direction to the others. This design reduces the axial imbalance but, nevertheless, incorporates a

sp

eci

al

bal

an

ce

arr

an

ge

me

nt.

Page 14 of 72

Fig. 14 Mixed

Design For

Axial Balance

Fig. 15 a) shows the principle of a multistage double suction type of pump. The shaft and the

impellers are completely balanced. The principle of this type of pump is a combination of double

suction and single suction impellers to form a multistage pump.

Fig. 15 a) Combination Of Single And Double Suction Impellers

2.3 Series Connected Multistage Pump

It may be necessary in some cases to have a high head with a low capacity and in other cases a

low discharge head with a high capacity.

The six stages pump in fig.15 b) can be connected in series by means of a switch valve in the

suction line. The pumps act as a multistage single suction pump and the discharge head is six

Page 15 of 72

times the discharge head of the first impeller. The capacity is the capacity of the first impeller.

The shaft is in balance.

Fig. 15b) Series

Connection Of

Impellers

In fig. 16 the same pump

is connected in parallel. Now the pump acts as a multistage double suction pump. The discharge

head is three times the discharge head of the first impeller and the capacity is twice the capacity

of the first impeller. The shaft remains in balance.

As can be seen from the situation shown in fig. 15 the total discharge head is half of that of the

situation shown in fig. 16 and the capacity is twice of the situation shown in fig. 15.

Fig. 16 Parallel

Connection Of

Impellers

Page 16 of 72

3. HYDRAULIC BALANCING DEVICES

A single suction impeller is subjected to axial hydraulic thrust caused by the pressure differential

between its two faces. If all the single suction impellers of a multistage pump face in the same

direction, the total theoretical hydraulic axial thrust acting toward the suction end of the pump

will be the sum of the individual impeller thrusts.

Some form of hydraulic balancing device must then be used to balance the axial thrust and to

reduce the pressure on the stuffing box adjacent to the last stage impeller. This hydraulic

balancing device may be a balancing drum, a balancing disc, or a combination of the two.

3.1 Balancing Drums

The balancing drum is illustrated in fig. 17. The balancing chamber at the back of the last stage

impeller is separated from the pump interior by a drum that is keyed or screwed to and rotates

with the shaft. The drum is separated by a small radial clearance from the stationary portion of

the balancing device, called the balancing drum head, which is fixed to the pump casing.

Page 17 of 72

Fig. 17 Balancing Drum

Fig. 18

Simple

Balancing

Disc

Page 18 of 72

The back of the balancing disc is subject to the balancing chamber back pressure whereas the

disc ace experiences a range of pressure. These vary from discharge pressure at its smallest

diameter to back pressure at its periphery. The inner and outer disc diameters are chosen so that

the difference between the total force acting on the disc face and that acting on its back will

balance the impeller axial thrust.

If the axial thrust of the impellers should exceed the thrust acting on the disc during operation,

the latter is moved toward the disc head, reducing the axial clearance between the disc and the

disc head. The amount of leakage through the clearance is reduced so that the friction losses in

the leakage return line are also reduced, lowering the back pressure in the balancing chamber.

This lowering of pressure automatically increases the pressure difference acting on the disc and

moves it away from the disc head, increasing the clearance. Now the pressure builds up in the

balancing chamber, and the disc is again moved toward the disc head until an equilibrium is

reduced.

To assure proper balancing disc operation, the change in back pressure in the balancing chamber

must be of an appreciable magnitude. Thus with the balancing disc wide open with respect to the

disc head, the back pressure must be substantially higher than the suction pressure to give a

resultant force that restores the normal disc position. This can be accomplish by introducing a

restricting orifice in the leakage return line that increase back pressure when leakage past the

disc increase beyond normal. The disadvantage of this arrangement is that the pressure on the

stuffing box packing is variable a condition that is injurious to the life of the packing and

therefore to be avoided. The higher pressure that can occur at the packing is also undesirable.

The balancing chamber is connected either to the pump suction or to the vessel from which the

pump takes its suction. Thus the back pressure in the balancing chamber is only slightly higher

than the suction pressure, the difference between the two being equal to the friction losses

between this chamber and the point of return. The leakage between the drum and the drum head

is, of course, a function of the differential pressure across the drum and the clearance area.

The forces acting on the balancing drum in fig.17 is:-

a) towards the discharge end the discharge pressure multiplied

b) the front balancing area (area B) of the drum.

c) towards the suction end the back pressure in the balancing chamber multiplied by the back

balancing area (area C) of the drum.

Page 19 of 72

The first force is greater than the second thereby counterbalancing the axial thrust exerted upon

the single suction impeller.

Since varying head and capacity conditions change the pressure distribution, and as the areas of

the balancing drum is fixed, in practice 100% balance is unattainable. The balancing drum is

often design to balance only 90 to 95% of total impeller thrust. The balancing drum satisfactorily

balances the axial thrust of a single suction impellers and reduces pressure on the discharge side

of the stuffing box. It lacks however, the virtue of automatic compensation for any changes in

axial thrust caused by varying impeller reaction characteristic. In effect, if axial thrust and

balancing drum forces become equal, the rotating element will tend to move in the direction of

the greater force.

The thrust bearing must then prevent excessive movement of the rotating element. The balancing

drum performs no restoring function until such time as the drum force again equals the axial

thrust. This automatic compensation is the major feature that differentiates the balancing disc

from the balancing drum.

3.2 Balancing Disc

The operation of a simple balancing disc is illustrated in fig. 18. The disc is fixed to and rotates

with the shaft. It is separated from the balancing disc head which is fixed to the casing by a small

axial clearance. The leakage through this clearance flows into the balancing chamber and from

there either to the pump suction or to the vessel from which the pump takes its suction.

3.3 Combination Balancing Disc And Drum

For reasons just described, the simple balancing disc is seldom used. The combination balancing

disc and drum (fig. 19) was developed to obviate the short comings of the disc while retaining

the advantage of automatic compensation for axial thrust changes.

The rotating portion of this balancing device consists of a long cylindrical body that turns within

a drum portion of the disc head. This rotating part incorporates a disc similar to the one

previously described. In this design, radial clearance remains constant regardless of disc

position, whereas the axial clearance varies with the pump rotor position. The following forces

act on this device:-

Page 20 of 72

a) toward the discharge end:- the sum of the discharge pressure multiplied by area A, plus the

average intermediate pressure multiplied by area B.

b) toward the suction end:- the back pressure multiplied by area C.

Fig.

19

Comb

ination Balancing Disc And Drum

Whereas the position-restoring feature of the simple balancing disc required an undesirably wide

variation of the back pressure, it is now possible to depend upon a variation of the intermediate

pressure to achieve the same effect. Here is how it works When the pump rotor moves toward

the suction end (to the left fig. 19) because of increased axial thrust, the axial clearance is

reduced, and pressure builds up in the intermediate relief chamber, increasing the average value

of the intermediate pressure acting on area B. In other words, with reduced leakage, the pressure

drop across the radial clearance decreases, increasing the pressure drop across the axial

clearance. The increase in intermediate pressure forces the balancing disc toward the discharge

end until equilibrium is reached. Movement of the pump rotor toward the discharge end would

Page 21 of 72

have the opposite effect of increasing the axial clearance and the leakage and decreasing the

intermediate pressure acting on area B.In other words,with reduce,with reduced leakage,the

pressure drop across the radial clearance decreases,increasing the pressure across the axial

clearance.The increase in intermediate pressure forces the balancing dosc toward the discharge

end until equilibrium is reached.movement of the pump rotor towards the discharge end would

have the opposite effect of increasing the axial clearance and the leakage and decreasing the

intermediate pressure acting on area B.

4. MULTISTAGE HORIZONTAL CENTRIFUGAL PUMPS IN PLANT

Two models of large multistage horizontal centrifugal pumps used in most plant i.e sulphinol

pump,boiler feed water pump etc. Their driver are motor driven,turbine driven. The description

that follow apply to both these models unless stated.

4.1 Description Of Pump

Both models of pumps are of the barrel design and consists of three main elements viz. the

barrel, the internals (interstage pieces with rotor) and the cover, see fig.20 and 21. The internals

slide into the barrel as one unit, after which the cover is mounted. A spring element against the

cover permits free expansion at sudden temperature fluctuation. At varying temperatures o the

liquid to be handled misalignment of pump and driver is prevented by centerline mounting of the

pump and pedestals. To permit expansion in longitudinal direction heavy pin fixes the barrel

under the inlet of one side, while on the other side a key guides the barrel.

a) Shaft and sleeves

The shaft is of the stepped design permitting simple dismantling and assembling of impellers.

The shaft sleeves protect the shaft against wear and / or erosion by the packing.

The sleeves are lined with copper on the inside, to obtain a layer which precluded scoffing

during assembly and removal of the sleeve.

b) Impellers

Page 22 of 72

The impellers can move in axial direction on the pump shaft, so that in the event of sudden

temperature changes the impeller hubs or the pump shaft are not exposed to additional stresses.

The shaft power is transmitted to the impellers by means of keys.

c) Interstage pieces with diffusers

Together with the complete rotor they form the internals. The outlet pressure compresses the

interstage pieces which have metal to metal joints.

d) Axial rotor balancing

Axial balancing is effected by a drum and a double axial bearing. The diameter of the drum is

such, that the axial thrusts of the rotor are balances at the best efficiency point. The double axial

bearing takes the positive or negative residual thrusts at other capacities.

The drum reduces the pressure of the leakage liquid to the inlet pressure, so that both shafts

sealings are under the same inlet pressure.

e) Shaft sealing

The flushing of the mechanical seal is effected by a closed system incorporating a cooler. The

circulation of the flushing liquid is effected by a pumping ring which is a part of the rotating

element of the seal.

f) Bearings

The pump shaft is supported by two forced feed lubricated sleeve bearings. Fig.22 shows the

flow diagram. The bearing housings are fixed by means of a heavy flange to the bearing

pedestals of the pump. After a correct height adjustment of the pump shaft by means of adjusting

bolts in the flanges of the bearing pedestals, the position of the bearing housing is fixed by pins.

The bearing housings are axially split; inspection of the bearings is possible without it being

necessary to disassemble further parts. Oil baffles are fitted to prevent oil leakage. The bearing

housing on discharge side is provided with a double axial sleeve bearing to take residual thrusts

caused by axial unbalance of the rotor. This double axial sleeve bearing is also forced feed

lubricated.

Page 23 of 72

4.2 Double Thrust Bearing

The pump is provided with a double thrust bearing, so that axial forces can be absorbed on both

sides.

The clearance A=0.2mm on other side. The double thrust bearing, fig.23 consists of:

001 2 spacers rings

002 2 plain rings

003 16 thrust pads (8 on each side)

004 1 thrust collar

005 1 assembling ring

1 2 3 4 5 6 7 8

Page 24 of 72

1 2 3 4 5 6 7 8

Fig 22 Flow Diagram LUBE OIL & WATER

Page 25 of 72

Fi

Fig. 23 Double Thrust Bearing

Page 26 of 72

If the thrust bearing has been damages, normally only the thrust pads 003 will have to be

replaced. If the thrust collar 004 has also to be replaced, in which case also the assembling ring

005 has to be replaced and made to suit such that clearance A is available again.

Also when new inner casings are mounted the assembling ring 005 has to be replaced and made

to suit. As spare the assembling ring supplied with an over width of 3mm.

4.3 General Dismantling And Reassembling Instructions

Before reassembling the pump, all the parts must be checked. The shaft must be straight, the

deviation may not exceed 0.05mm. All the burrs on the aces of the impellers, wearing rings,

bushings, shaft sleeves, etc. must be removed.

Extreme attention must be paid to the locking of wearing rings, shaft sleeve, nuts etc. and all

check screws theirselves are to be fastened by means of a center punch.

Molykote type z to be used when mounting sleeves etc. on the shaft and for installing stationary

rings and bushing.

Before mounting the bearings have to be heated in an oil bath till 100 Mount the bearings on the

shaft while hot. Half couplings have to be mounted on the shaft ends by means of a hammer or

otherwise.

Before mounting into the pump case the concentricity of the rotor has to be checked.

After reassembling the pump has to be aligned securely with respect to the driver. When all the

pipe have been connected to the pump the alignment has to be checked again. At the same time

must be checked if it is possible to turn the rotor of the complete unit easily by hand. At last the

coupling spacer and the coupling bolts have to be mounted.

4.4 Checking Concentricity Of Rotating Elements

i) Mount all impellers including the balancing drum and shaft sleeves on the shaft. Locked the

parts in running position.

ii) Mount the rotating assembly in centers. Revolves the shaft slowly by hand and with the

indicator dial set at 0, take readings near each end at the centre of each shaft sleeve and on the

circumference of each impeller wearing ring.

If the readings at any point indicated do not very more than 0.04mm the assembly is accurate and

may be fitted into the pump. If the readings vary more than 0.04mm the assembly is not accurate

Page 27 of 72

and should be fitted into the pump until the cause of the inaccuracy has been found and the

trouble has been eliminated.

4.5 Adjustment Of Pump Rotor

To prevent the impellers from wearing the wearing ring and bushing respectively in the first

stage cover and interstage pieces, the bearings have been adjusted in the bearing housing 602A

and 602C (P1101 602B).

After adjustment in the factory the housing is fixed by means of dowels. The radial clearances

between the impellers and bushings and between balancing sleeve 1210 and balancing drum 276

is shown in fig.24 and 25.

Page 28 of 72

Fig. 24 Pump Rotor Clearances

Page 29 of 72

Fig. 25 Pump Rotor Clearances

Page 30 of 72

4.6 Axial clearances

The pump rotor has a total clerance of 9mm (8.25mm) in the axial direction, see fig. 26B

when the balancing drum and thrust bearing has not yet been fitted. Between the shaft collar on

inlet side and the balancing drum 276 the impellers have a total clerance of 2mm, see fig. 26A.

This clearance is essential for free expansion of the impellers and the shaft when these are

heating up. When the pump is running, it is possible that the impellers move on the shaft in the

radial direction, as the axial thrust, which always acts in only one direction (toward the inlet) is

rather high. When the pump has been dismantled, the axial clerances have to be checked by

fitting an assembling bush 200 instead of the balancing drum (seefig. 26B). With the aid of the

assembling bush 200 and nut 256B all the impeller hubs are pushed rigidly together as well as

against the shaft collar at the inlet side of the pump.

The correct position of the rotor is that at which the clerances of 9mm ( 8.25mm) is distributed in

such a way that the clerance to the inlet side is 4.5mm ( 4.75mm), and to the outlet side 4.5mm /

3.5mm), see fig. 27. This must be checked before the stuffing boxes and the bearings are fitted

on the pump.

Here after the assembling bush is removed and the balancing drum is fitted on the shaft. When

the stuffing boxes and the bearings have been fitted, the rotor is adjusted such that the clearance

between the impeller and the interstage piece is 4.5mm (P1101 - 3.5mm) towards the outlet side.

Page 31 of 72

Fi

g.

2

6A Balancing Device

Page 32 of 72

Fig. 26 (B) Balancing Device

Fig. 27 First Stage Impeller

5. MULTISTAGE PUMP

A single-stage centrifugal pump has one impeller while multistage are those with two or more

impellers. Normally these impellers are arranged in series, that is the first impeller discharges

into the suction of the second impeller.

The head at the discharge of the second impeller is greater than the head at the discharge of the

first, thus the greater the number of impellers, the higher the final discharge head is.

So, a single multistage pump can produce higher heads than a single-stage pump.

Since liquid are nearly incompressible, all the impellers in the pump are designed for about the

same capacity and the total capacity is only the capacity of the first impeller.

In rare cases are the impellers connected in parallel, that is exactly the same as two entirely

separate pumps, each taking suction from the same source and discharging to the same place.

Page 33 of 72

In this case the total capacity is the sum of the individual pump or impeller capacities, but the

head is not increased above that for one pump or impeller.

Fig. shows a multistage boiler feed pump. The impellers can either be a single suction or double

suction.

The casing of multistage centrifugal pump assembly may be of the:

a) Split casing type

b) Barrel type or

c) Ring type

The three types of casing are shown in Fig. below.

Page 34 of 72

Fig

.

T

y

p

es

of

s

pl

it

c

asing two stage pump

Page 35 of 72

Fig. Six stage pump, split casing type

Fig. Multistage Pump

Page 36 of 72

Fig. Multistage Pump

The following information is available:

1. 6 stage centrifugal pump

2. Suction pressure 4 kg/cm2

3. Discharge pressure 54 kg/cm2

4. All impeller wear ring diameters 120mm

5. All sleeves and pressure reducing sleeve diameters 60mm

6. Inside diameter disc plate 120mm

7. Outside diameter balancing disc 180mm

8. Discharge pressure reduced over reducing sleeve to 34 kg/cm2

9. Liquid pressure reduced over clearance disc to 4 kg/cm2

What is the axial unbalance force and in which direction does it act?

A four stages centrifugal pump has the following datas:

i) wear ring diameter : 120mm

Page 37 of 72

ii) sleeve diameter : 50mm

iii) suction pressure respectively 10, 25, 40 and 55 kg/cm2

iv) discharge pressure respectively 24, 40, 55 and 70 kg/cm2

A) What is the axial unbalance force and in which direction does it acts (show working).

B) If a balancing drum of 120mm diameter is installed what will be the axial unbalance

force (show working).

6. MULTISTAGE PUMPS:

6.1) Most multistage pumps are built with single suction impellers. To balance the

axial thrust of these impellers, two arrangements are used:

1) The impellers all face the same direction and are mounted in he ascending

order of the stages. The axial thrust is balanced by a hydraulic balancing device

(Fig. 17 and Fig. 18).

2) An even number of single suction impellers is used, one half of these facing in

a direction opposite to the second half as illustrated below (Fig.16).

Fig

. 16

Four Stage Pump With Opposed Impellers

This mounting of single suction impellers back to back is frequently called Opposed Impellers.

Page 38 of 72

Fig. 17

Multistage Pump With Single Suction Impellers Facing In One Direction And With Hydraulic

Balancing.

Fig. 18

Double Casing Multistage Pump With Radially Split Inner Casing

6.2) Hydraulic balancing devices may take the form of:

1) a balancing drum

Page 39 of 72

2) a balancing disc

3) a combination of the two.

Fig.

19 Balancing Drum

a) The balancing chamber at the back of the last stage impeller is separated from

the pump interior by a drum mounted on the shaft.

b) The drum is separated by a small radial clearance from the stationary portion of

the balancing device, called the balancing drum head, which is fixed to the

pump casing.

c) The balancing chamber is connected either to the pump suction or to the vessel

from which the pump takes its suction. The force acting on the balancing drum

are:

1) towards the discharge end the discharge pressure multiplied by the

front balancing area (area B) of the drum;

2) toward the suction end the back pressure in the balancing chamber

multiplied by the back balancing area (area C) of the drum.

d) The first force is greater than the second, thereby counterbalancing the axial

thrust exerted upon the single suction impellers. The drum diameter can be

Page 40 of 72

selected to balance the axial thrust completely or to balance 90 to 95 percent of

this thrust, depending on whether a slight thrust load in a specific direction on

the thrust bearing is desirable.

e) The operation of a simple balancing disc as shown below:

Fig. 19

Simple Balancing Disc

f) The rotating disc is separated from the balancing disc head by a small axial clearance.

The leakage through this clearance flows into the balancing chamber and from there

either to the pump suction or to the suction vessel.

g) The back of the balancing disc is subjected to the balancing chamber back pressure,

whereas the disc face experiences a range of pressures.

These vary from discharge pressure at its smallest diameter to back pressure

periphery.

h) The inner and outer disc diameter are chosen so that the diffrence between the total

force acting on the disc face and that acting on its back will balance the impeller axial

thrust.

i) if the axial thrust of the impellers should exceed the thrust acting on the disc during

operation, the latter is moved toward the disc head, reducing the axial clearance.

Page 41 of 72

j) the amount of leakage through this clearance is reduced so that the friction losses in the

leakage return line are also reduced, lowering the back pressure in the balancing

chamber.

This automatically increases the pressure difference acting on the disc and moves it

away from the disc head, increasing the clearance.

k) now the pressure will builds up in the balancing chamber, and the disc is again moved

toward the disc head until an equilibrium is reached.

l) To ensure proper balancing-disc operation, the cahnge in back pressure must be of an

appreciable magnitude.

This is accomplished by introducing a restricting orifice in the leakage return line.

m) The combination disc and drum (Fig.18) is the most commonly used hydraulic

balancing device.

It incorporates portions rotating within radial clearances of stationary portions and a

disc face rotating within an axial clearance of another portion of the stationary part.

n) The radial clearance remains constant regardless of any displacement of the rotor

within the casing. Such displacement, however changes the axial within the balancing

device. These changes cause changes in the leakage, which in turn change the pressure

drop across the radial clearance and thus increase the average value of the pressure

acting on the disc face.

o) These changes in the intermediate pressure on the disc face act to move the balancing

device in whichever direction is required to restore equilibrium and axial balance.

7.0) Axial Thrust In Single Stage And Multistage Pumps.

7.1) Axial Thrust:

Axial hydraulic thrust is the summation of unbalances impeller force acting in the

axial direction.

a) Theoretically,a double suction impeller is in hydraulic balance, with the pressure on

one side equal to and counter balancing the pressures on the other.

b) In practice, some slight unbalance may exist, and even double suction pumps are

provided with thrust bearings.

Page 42 of 72

The single suction radial flow impeller is subjected to axial thrust because a portion

of the front wall is exposed to suction pressure, with a greater back-wall surface

subject to discharge pressure. In addition, an overhung single suction impeller with

single stuffing box is subjected to an axial force equivalent to the product of the shaft

area through the stuffing box and the difference between suction and atmospheric

pressure.

c) This force acts toward the impeller suction when the suction pressure is less than the

atmospheric and in the opposite direction when it is higher than the atmospheric.

d) To eliminate the axial thrust of single suction impeller,a pump can be provided with

both front and back wear rings.

Pressure approximately equal to the suction pressure is maintained in a chamber

located on the inner side of the back wearing ring by providing a so called balancing

holes through the impeller. Leakage pass the back wear ring is returned into the

suction area through these holes. In large pumps, a pipe connection usually replaces

the balancing holes.

8.0) Shafts and Shaft Sleeves:

8.1) Shaft:

Pump shaft diameters are usually larger than actually needed to transmit the torque

because their size is dictated by the maximum permissible or desirable deflection.

This defection is itself chosen to prevent possible contact at the wearing surfaces

while maintaining reasonable clearances that will not affect pump efficiency too

unfavorably.

The first critical speed of a shaft is related to its deflection.

It follows that a shaft design permitting a deflection of, for instance;

...0.005 to 0.006 inch (0.13 to 0.15mm) will have a first critical speed of 2,400 to

2,650 rpm.

This is the reason for using rigid shafts (operating below their first critical speed) for

pumps that operate at 1750 rpm or lower.

Multistage pumps operating at 3600 rpm or higher use shaft of equal stiffness (for

the same purpose of avoiding wearing contact).

Page 43 of 72

However, their corresponding critical speed is about 25 to 40 percent less than their

operating speed.

This margin is sufficient to avoid any danger to the operation caused by critical

speed effect.

8.2) Shaft Sleeves:

Pump shafts are usually protected from erosion, corrosion, and wear at the stufing

boxes and leakage joints and in the waterways by renewable sleeves.

The most common shaft sleeve function is that of protecting the shaft from wear at a

stuffing box.

Shaft sleeves serving other functions are given speciic names to indicate their

purpose.

For example:

A shaft sleeve used beteen two multistage impellers in conjunction with an interstage

bushing to form an interstage leakage joint is called an interstage or distance sleeve.

9.0) Stuffing Boxes : (Packings and Mechanical Seals)

Stuffing boxes have a primary function of protecting the pump against leakage at the

point where the shaft passes out through the pump casing.

If the pump handles a suction lift and the pressure at the interior stufing box end is

below atmospheric, the stuffing box function is to prevent air leakage into the pump.

9.1) Packings:

The stuffing box takes the form of a cylindrical recess taht accommodates a number

of rings of packing around the shaft or shaft sleeve.

A) if the sealing the box is desired, a latern ring or seal cage is used to separate

the rings of packing into approxiamtely equal sections.

B) the packing is compressed to give the desired it on the shaft or sleeve by a

gland that can be adjusted in an axial direction.

C) water or some other sealing fluid can be introduced under pressure into the

space provided by the seal cage, causing flow of sealing fluid in both axial

Page 44 of 72

directions.

Water or some other sealing fluid is useful for pumps handling flammable or

chemically active and dangerous liquids since it prevents outflow of the

pumped liquid.

D) when a pump handles with a clean, cool water, stuffing box seals are usually

connected to the pump discharge or, in multistage pumps, to an intermediate

Stage.

E) an independent supply of sealing water should be provided if any of the

following conditions exist:

1) a suction lift in excess of 15 ft. (4.5m)

2) a discharge pressure under 10 lbs./ln2 (0.7 kg/cm2)

3) hot water handled adequate cooling (except for boiler feed pumps, in which

seal cage are not used)

4) muddy, sandy, or gritty water handled

5) for all hot well pumps

6) no leakage to atmosphere permitted of the liquid handled

F) when sealing water is taken from the pump discharge, an external connection

is generally made to the seal cage through small diameter piping, or an

internal passage connection is made within the pump itself.

G) high temperatures and pressures complicate the problem of maintaining

stuffing box packing. Pumps in more difficult services are usually provided by

jacketed, water cooled stuffing boxes.

H) if the pressure ahead of the stuffing box makes it impractical to pack the

stuffing box satisfactorily, a pressure reducing breakdown or labyrinth may be

located ahead of the box, with the leakage past the pressure reducing

breakdown being returned to some point of lower pressure in the pumping

cycle.

Basically, stuffing box packing is a pressure breakdown device that is

sufficiently plastic to be adjusted for proper operation.

I) the most common types are asbestos packing and metallic packing, the latter

being composed of flexible metallic strands or foil with graphite or oil

Page 45 of 72

lubricant and with either an asbestos or plastic core.

Other type of packing used may be hemp, cord, braided, duck fabric, chevron,

etc...

Packing is supplied either in continous coils of square cross sections or in

preformed, die moulded rings.

9.2) Mechanical Seals:

Mechanical seals are used in centrifugal pumps when it becomes impractical to use

conventional packing with radial sealing surfaces.

A) the sealing suraces of mechanical seal are located in a plane perpendicular to

the pump shaft and consist of two highly polished surfaces running adjacently,

one surface being connected to the shaft and the other to the stationary portion

of the pump.

B) the srace are held essentially in contact by a spring, the axial clearance

between the surfaces being provided by a thin film of liquid.

The flow of liquid may be only a drop every few minutes or even a haze of

escaping vapor.

C) there are two basic seal arrangements:

1) internal assembly

2) external assembly

Two mechanical seals may be mounted inside the stuffing box to make a double seal

assembly.

Such an arrangement is used or pumps handling toxic or highly inflammable liquids.

A clear, filtered and generally inert sealing liquid is injected between two seals at a

pressure slightly in excess of the pressure in the pump ahead of the seal.

10) Bearings:

All types of bearings are used in centrifugal pimps. Even the basic design of pump is

often made with two or more different bearings, required by varying service conditions.

A) Two external bearings, required are usually used for the double suction single

Page 46 of 72

stage general service pumps, one on either side of the casing.

B) in horizontal pumps with bearings on each end, the inboard bearing is the one

between the casing and the coupling and the outboard bearing is located at the

opposite end.

C) pumps with overhung impellers have both bearing on the same side of the

casing, the bearing nearest the impeller is the inboard bearing, and the one

furthest away the outboard bearing.

10.1) Ball Bearings:

Ball bearing are most common anti-friction bearings used on centrifugal pump.

Ball bearings used in centrifugal pumps are usually grease lubricated, although

some services use oil lubrication.

10.2) Roller Bearings:

Roller bearings are used less often, although the spherical roller bearing is used

frequently for large shaft sizes.

10.3) Sleeve Bearings:

Sleeve bearing are used for large, heavy duty pumps with shaft diameters of such

proportion that necessary anti riction are not commonly available.

A) another application is for high pressure multistage pumps operating at speed of

3,600 to 9,000 rpm.

Still another application is in vertical submerged pumps, such as vertical

turbine pumps in which the bearings are subjected to a water contact.

Most sleeve bearings are oil lubricated.

10.4) Thrust Bearings:

Thrust bearings used in combination with sleeve bearings are gererally

Kingsbury-type bearings.

11) COUPLINGS

Page 47 of 72

Centrifugal pumps are connected to their drivers through couplings of one sort or

another, except or close coupled units in which the impeller is mounted on an extension

of the shaft of the driver.

Couplings used with centrifugal pumps can be either rigid (o the clamp or compressor

type) or flexible (pin-and-buffer, gear, grid, or flexible-desk type)

12) PUMP MOUNTING

It is desirable taht pumps and their drivers be removable from their mountings.

Consequently, they are usually bolted and doweled to machined surfaces that, in turn,

are firmly connected to the foundations.

These machined suraces are usually part of a common bed plate on which the pump and

its driver have been pre-aligned.

Bedplates are made of either cast iron or structural steel. Cast iron or steel soleplates are

customarily used for vertical dry-pit pumps and also for some of the larger horizontal

units.

13) IMPELLER DESIGN

Single suction:

A) liquid enters the suction eye on one side (Figure 37)

B) normally used in multistage pumps

C) preferred for pumps handling suspended matter

D) can result in high axial unbalanced loads due to the differential pressure

Page 48 of 72

between suction and discharge

E) pressure unbalance and resulting axial loads are controlled by:

1) thrust bearings

2) balancing holes in the impeller

3) pump-out vanes on the back side of the impeller shroud

Double suction:

A) similar to having two single suction impellers placed back-to-back in a single

casing (Figure 38).

B) liquid enters the impeller symmetrically rom both sides (Figure 39)

C) increases pump capacity

D) avored because both sides of the impeller are at suction so axial thrust loads

are balanced.

E) greater suction area allows the pump to operate with less net absolute suction

head

F) difficult to manufacture

G) commonly used in conjunction with axially (horizontally) split casings with

bearings on each side of the housing.

Page 49 of 72

Figure 37: Single Suction Impeller

Figure 38: Double Suction Impeller

Page 50 of 72

Multistage design

A) Two or more impellers are attached to the same shaft and are housed in the

same unit (Figure 40).

B) Discharge of each impeller feeds the suction of the next impeller in series.

C) Stage piece and diaphragm isolates adjoining chambers from one another thus:

1) minimizing leakage between high and low pressure sides.

2) allows leakage only at the clearance joints formed between the

stationary and rotating elements of the pump.

D) Commonly used for:

1) high head (pressure) applications such as injection

2) deep wells

3) other high pressure applications

Page 51 of 72

Figur

e 39: Components of a Double Suction Impeller

Figure 40: Multistage Centrifugal Pump

Page 52 of 72

14) MULTI-STAGE & CENTRIFUGAL PUMP TYPES SELECTION CRITERIA

Vertical in line single stage pump

A) Internal

maintenance requires motor

removal

B) Capable of

producing heads up to 700

feet, capacities ranging from

20 to 1200 gpm

and can withstand

temperatures up to 500 F.

Horizontal multistage pump (Figure 77)

A) can be horizontally or radially split (donut) design.

B) in the horizontal split type, the lower half connects to the piping and foundations and

the upper half permits easy access to the internal rotating elements.

C) impellers are mounted back-to-back

D) maximum working pressure in horizontally split designs are limited by the strength of

the joint connecting the top and bottom casings.

E) radial split and barrel type casings are used for higher pressures.

F) both types are limited to the number of stages due to the difficulty in limiting

deflection over the long span between bearings.

G) capable of producing capacities from 20 to 11,000 gpm, producing heads up to 5500

feet (depending on the number of impellers), and can operate at temperatures of 500F.

Page 53 of 72

Fig

ure

77:

Ho

riz

ont

al

Multistage Pump

CENTRIFUGAL PUMP TYPES AND SELECTION CRITERIA

Donut-type pump (Figure 78)

Page 54 of 72

A) consist of a series of impellers and casings that form individually shaped units, called

donuts.

B) held together by long bolts and are connected by a long shaft.

C) capable of producing heads to 1300 feet, capacities up to 1100 gpm, and pumping

temperatures to 300 F.

D) even though they are designed to have optimum speciic speeds, an optimum number

of valves, and an optimum discharge angle, their head capacity curve droops.

The TDH peaks at some capacity greater than zero and falls to shut off, thus the pump

may operate at two different flow rates for the same head.

E) when running two of these pumps in parallel, he lower volume pump will stop

pumping while the higher volume pump will increase its flw rate, both in response to

the same signal of a desired increased flow rate.

F) the shutdown pump cannot be started up against a second pump that may be operating

above the first pumps shut off head.

Page 55 of 72

Figure 78: Donut Pump

Figure 78: Donut

Pump Layout

CENTRIFUGAL PUMP TYPES AND SELECTION CRITERIA

Donut-type pump Page 56 of 72

A) although individual back pressure valves and larger recycle lines can be installed, the

extra cost does not outweigh the small amount of increased effeciency common to

donut pumps.

B) mechanically, these pumps are not ideal due to their many parts, sealing surfaces and

increased chance of leaks and vibration.

Vertical multistage pump (Figure 79).

A) may consist of up to 24 or more, mixed flow impellers located below grade, thus they

have low NPSHs (from 1 to 7 feet).

B) require a large number of close running clearances, thus they are very sensitive to

damage caused by solids or by running dry.

C) maintenance requires taht a great many parts be disassambled, repaired or replaced,

thus these pumps experience very high maintenance costs.

D) capable of producing heads up to 6000 feet, capacities from 5 to 400 gpm, and can

operate at temperature to 400 F.

Page 57 of 72

DEEP WELL CAN

Figure 79: Vertical Multistage Pumps

CENTRIFUGAL PUMP TYPES AND SELECTION CRITERIA

Vertical multistage pump

A) subjected to a variety of thrust loads which are caused by

1) A downward unbalanced discharge pressure across the eye area of the impeller

2) Aan upward force due to the change in direction of the liquid passing through

the impeller, and

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3) A downward force due to the weight of the rotating assembly less the weight of

the displaced fluid

B) If thrust force are too high, the impellers can be hydraulically balanced by holes in the

impeller between the discharge and suction sides

1) This increased internal recirculation and lowers pump efficiency

C)The magnitude and direction of the thrust will vary with flow rate; however it is

possible in most cases to have the maximum values calculated and supplied to the

motor manufacturer for proper motor bearing sizing.

Axial Thrust In Multistage Pumps

Most multistage pumps are build with single suction impellers in order to simplify the

design of the interstage connections. Two obvious arrangements are possible for the single

suction impellers:

1) Several single suction impellers may be mounted on one shaft, each having its suction inlet

facing in the same direction and its stages following one another in ascending order of pressure

The axial thrust is then balanced by a hydraulic balancing device.

2) An even number of single suction impellers may be used, one-half facing in one direction and

the other half facing in the opposite direction. With this arrangement, axial thrust on the first half

is compensated by the thrust in the opposite direction on the other half (Fig. 60). This mounting

of single suction impellers back to back is frequently called opposed impellers.

Page 59 of 72

.

Fig.60: Four stage pump with opposed impellers.

An even number of single suction impellers may be used with this arrangement, provided

the correct shaft and interstage bushing diameters are used to give the effect of a

hydraulic balancing device that will compensate for the hydraulic thrust on one of the

stages.

It is important to note taht the opposed impeller arrangement completely balances axial

thrust only under the following conditions:

1) the pump must be provided with two stuffing boxes

2) the shaft must have a constant diameter

3) the impeller hubs must not extend through the interstage portion of the casing

separating adjacent stages.

Except for some special pumps that have an internal and enclosed bearing at one end, and

therefore only one stuffing box, most multistage pumps fulfill the first condition. Because of

structural requirements, however, the last two conditions are not practical. A slight residual

thrust is usually present in multistage opposed impeller pumps and is carried on the thrust

bearing.

Page 60 of 72

15) HYDRAULIC BALANCING DEVICES

If all the single suction impellers of a multistage pump face in the same direction, the

total theoretical hydraulic axial thrust acting toward the suction end of the pump will be the sum

of the individual impeller thrusts. The thrust magnitude (in pounds) will be approximately equal

to the product of the net pump pressure ( in pounds per square inch) and the annular unbalanced

area (in square inches). Actually the axial thrust turns out to be about 70 to 80% of this

theoretical value.

Some form of hydraulic balancing device must be used to balance this axial thrust and to

reduce the pressure on the stuffing box adjacent to the last stage impeller. This hydraulic

balancing device may be a balancing drum, a balancing disk, or a combination of the two.

Balancing Drums

The balancing drum is illustrated in Fig. 61. The balancing chamber at the back of the

last stage impeller is separated from the pump interior by a drum that is either keyed or screwed

to the shaft and rotates with it. The drum is separated by a small radial clearance from the

stationary portion of the balancing device, called the balancing-drum head, which is fixed to the

pump casing.

The balancing chamber is connected either to the pump suction or to the vessel from

which the pump takes its suction. Thus the back pressure in the balancing chamber is only

slightly higher than the suction pressure, the difference between the two being equal to the

friction losses between this chamber and the point of return. The leakage between the drum and

the drum head is, of course, a fuction of the diffrential pressure across the drum and of the

clearance area.

The forces acting on the balancing drum in Fig. 61 are the following:

1) toward the discharge end: the discharge pressure multiplied by the front balancing area (area

B) of the drum.

2) toward the suction end: the back pressure in the balancing chamber multiplied by the back

balancing area (area C) of the drum.

Page 61 of 72

Fig.61 Balancing Drum

The first force is greater than the second, thereby counterbalancing the axial thrust exerted upon

the single suction impellers. The drum diameter can be selected to balance axial thrust

completely or within 90 to 95%, depending on the desirability of carrying any thrust-bearing

loads.

It has been assumed in the preceding simplified description that the pressure acting on the

impeller walls is constant over their entire surface and that the axial thrust is equal to the product

of the total net pressure generated and the unbalanced area. Actually this pressure varies

somewhat in the radial direction because of the centrifugal force exerted upon the water by the

outer impeller shroud (Fig. 54). Futhermore, the pressures at two corresponding points on the

opposite impeller faces (D and E, Fig. 61) may not be equal because of variation in clearance

between the impeller wall and the casing section separating successive stages. Finally, pressure

distribution over the impeller wall surface may vary with head and capacity operating conditions.

Page 62 of 72

This pressure distribution and design data can be determined by test quite accurately for

any one fixed operating condition, and an effective balancing drum could be design on the basis

of the forces resulting from this pressure distribution. Unfortunately, varying head and capacity

conditions change the pressure distribution,and as the area of the balancing drum is necessarily

fixed, the equilibrium of the axial forces can be destroyed.

The objection to this is not primarily the amount of the thrust, but rather that the direction

of the thrust cannot be predetermined because of the uncertainty about internal pressures. Still it

is advisable to predetermined normal thrust direction, as this can inlfuence external mechanical

thrust bearing design. Because 100% balance is unattainable in practice and because the slight

but predictable unbalance only 90 to 95% of total impeller thrust.

The balancing drum satisfactorily balances the axial thrust of single suction impellers and

reduces pressure on the discharge side stuffing box. It lacks, however, the virtue of automatic

compensation for any changes in axial thrust caused by varying impeller reaction characteristics.

In effect, if the axial thrust and balancing drum forces become unequal, the rotating element will

tend to move in the direction of the greater force. The thrust bearing must then prevent excessive

movement of the rotating element. The balancing drum performs no restoring fuction until such

time as the drum force again equals the axial thrust. This automatic compensation is the major

feature that differentiates the balancing disk from the balancing drum.

Balancing Disks

The operation of the simple balancing disk is illustrated in Fig.62. The disk is fixed to and

rotates with the shaft. It is separated by a small axial clearance from the balancing disk head,

which is fixed to the casing. The leakage through this clearance flows into the balancing

chamber and from there either to the pump suction or to the vessel from which the pump takes its

suction. The back of the balancing disk is subject to the balancing chamber back pressure,

whereas the disk face experiences a range of pressures. These vary from discharge pressure at its

smallest diameter to back pressure as its periphery. The inner and outer disk diameters are

chosen so that the difference between the total force acting on the disk face and that acting on its

back will balance the impeller axial thrust.

If the axial thrust of the impellers should exceed the thrust acting on the disk during

operation, the latter is moved toward the disk head, reducing the axial clearance between the disk

Page 63 of 72

and the disk head. The amount of leakage through the clearance is reduced so that the friction

losses in the leakage return line are also reduced, lowering the back pressure in the balancing

chamber. This lowering of pressure automatically increases the pressure difference acting on the

disk and moves it away from the disk head, increasing the clearance. Now the pressure builds up

in the balancing chamber, and the disk is again moved toward the disk head until an

equilibriumis reached.

To assure proper balancing disk operation, the change in back pressure in the balancing

chamber must be o an appreciable magnitude. Thus, with the balancing disk wide open with

respect to the disk head, the back pressure must be substantially higher than the suction pressure

to give a resultant force that restores the normal disk position. This can be accomplished by

introducing a restricting orifice in the leakage return line that increases back pressure when

leakage past the disk increases beyond normal. The disadvantage of this arrangement is that the

pressure on the stuffing box packing is variable - a condition taht is injurious to the life of the

packing and therefore to be avoided. The higher pressure that can occur at the packing is also

undesirable.

Combination Balancing Disk and Drum

For the reasons just described, the simple balancing disk is seldom used. The combination

balancing disk and drum (Fig. 63) was developed to obviate the shortcomings of the disk while

retaining the advantage of automatic compensation for axial thrust changes.

Fig. 62

Simple

Balancing

Disk

Page 64 of 72

CENTRIFUGAL PUMPS

Fig.

63 Combination Balancing Disk and Drum

The rotating portion of this balancing device consists of a long cylindrical body that turns

within a drum portion of the disk head. This rotating part incorporates a disk similar to the one

previously described. In this design, radial clearance remains constant regardless of disk

position, whereas the axial clearance varies with the pump rotor position. The following forces

act on this device:

1) toward the discharge end: the sum of the discharge pressure multiplied by area A, plus

the average intermediate pressure multiplied by area B.

2) toward the suction end: the back pressure multiplied by area C.

Whereas the position-restoring feature of the simple balancing disk required an undesirably wide

variation of the back pressure, it is now possible to depend upon a variation of the intermediate

pressure to achieve the same effect. Here is how it works: when the pump rotor moves toward

the suction end (to the left in Fig. 63) becaused of increased axial thrust, the axial clearance is

reduced and pressure builds up in the intermediate relief chamber, increasing the average value

Page 65 of 72

of the intermediate pressure acting on area B. In other words, with reduced leakage, the pressure

drop across the radial clearance decreases, increasing the pressure drop across the axial

clearance. The increase in intermediate pressure forces the balancing disk toward the discharge

end until equilibrium is reached. Movement of the pump rotor toward the discharge end would

have the opposite effect, increasing the axial clearance and the leakage and decreasing the

intermediate pressure acting on area B.

There are now in use numerous hydraulic balancing device modifications. One typical

design separates the drum portion of a combination device into two halves, one preceding and

the second following the disk (Fig. 64). The virtue of this arrangement is a definite cushioning

effect at the intermediate relief chamber, thus avoiding too positive a restoring action, which

might result in the contacting and scoring of the disk faces.

16) SHAFTS AND SHAFT SLEEVES

The basic function of a centrifugal pump shaft is to transmit the torques encountered in starting

and during operation while supporting the impeller and other rotating parts. It must do this job

with a deflection less than the minimum clearance between rotating and stationary parts. The

loads involves are (1) the torques, (2) the weight of the parts, and (3) both radial and axial

hydraulic forces. In designing a shaft, the maximum allowable deflection, the span or overhang,

Page 66 of 72

and the location of the loads all have to be considered, as does the critical speed of the resulting

design.

The shaft are usually proprotioned to withstand the stress set up when the pump is started

quickly, for example, when the driving motor is thrown directly across the line. If the pump

handles hot liquids, the shaft is designed to withstand the stress set up when the unit started cold

without any preliminary warmup.

17) Critical Speeds

Any object made of an elastic material has a natural period of vibration. When a pump rotor or

shaft rotates at any speed corresponding to its natural frequency, minor unbalances will be

magnified. These speeds are called the critical speeds.

In conventional pump designs, the rotating assembly is theoretically uniform around the

shaft axis and the center of mass should coincide with the axis of rotation. This theory does not

hold for two reasons. First,there are always minor machining or casting irregularities; second,

there are variations in metal density of each part. Thus, even in vertical-shaft machines having

no radial delection caused by the weight of the parts, this eccentricity of the center of mass

produces centrifugal force and therefore a deflection when the assembly rotates. At the speed at

which the centrifugal force exceeds the elastic restoring force, the rotor will vibrate as though it

were seriously unbalanced. If it is run at the speed without restraining forces, the deflection will

increase until the shaft fails.

18) Rigid and Flexible Shaft Designs

The lowest critical speed is called the first critical speed; the next highest is called the second,

and so forth. In centrifugal pump nomenclature, a rigid shaft means one with an operating speed

lower than its first critical speed; a flexible shaft is one with an operating speed higher tahn its

first critical speed. Once an operating speed has been selected, relative shaft dimensions must

still be determined. In othet words, it must be decided whether the pump will operate above or

below the first critical speed.

DRAWINGS

Page 67 of 72

BARREL-TYPE MULTI-STAGES PUMP

PUMP CONSTRUCTION/LAYOUT DRAWING

Page 68 of 72

MEASURING RING CLEARANCE IN MULTISTAGE PUMP

WITH DIAL INDICATOR

Page 69 of 72

MULTIS

TAGE

PUMP SHOWING WATER FLOW AND BALANCE VALVE

Page 70 of 72

Page 71 of 72

MULTISTAGE PUMP COUPLING HUBS REMOVAL & REINSTALL

Page 72 of 72