axial thrust in centrifugal pumps - experimental

15th International Conference on Experimental Mechanics

ICEM15 1

PAPER REF: 2977

AXIAL THRUST IN CENTRIFUGAL PUMPS - EXPERIMENTAL

ANALYSIS

Vasant Godbole1(*)

, Rajashri Patil1, S.S. Gavade

2

1Kirloskar Brothers Ltd., India 2RIT, Rajaramnagar, India (*)Email: [email protected]

ABSTRACT

Reliability of rotating machineries, especially centrifugal pumps is defined on account of wear

rate of components and bearing durability. One of the important performance parameter for

any centrifugal pump is its bearing life, which is dependent on the hydraulic thrusts, Radial[7,9]

and Axial Thrusts[7,9]

. The radial thrust prediction can be done accurately and easily with the

help of prevailing methods available in literature [7,9]

. Whereas to predict the axial thrust, the

prevailing methods are not found very much reliable to use as the number of parameters

influencing axial thrusts are more than those for radial thrust. Also to contribute to it are the

variations in the impeller geometries, which are governed by the hydraulic design and keeps

on changing based on the hydraulic design of the impeller. Therefore to predict the bearing

life and select the suitable bearing for getting optimum bearing life, it is really necessary to

formulate a basis to arrive at the reliable axial thrust values and formulation of such basis,

there is no better tool than experimental analysis.

This paper talks about the experimental analysis of the axial thrust of end suction long

coupled centrifugal process pump and the findings of the analysis.

1.0 INTRODUCTION

One of the important performance parameter for any centrifugal pump is its bearing life. To

have the minimum maintenance and repairs cost, the bearing life for any centrifugal pump

must be as prolonged as far as possible. Particularly in case of process pumps, which are used

at high temperatures, high pressure and more hazardous applications with maximum

reliability, bearing life is of more importance. Bearing life [12,13]

. of centrifugal pump depends

upon two hydraulic forces acting on the impeller i. e. radial thrust and axial thrust. Amongst

the two forces, radial thrust is dependant solely on pressure generated by pump. There are

proven and reliable methods available to predict the radial thrust generated. Radial thrust can

be satisfactorily reduced to harmless extents by using double volute casings or diffuser type

casings.

Whereas axial thrust is dependent on the many aspects viz. shroud and casing clearances,

peripheral shroud speeds, head developed by the pump, impeller geometry etc. due to this one

cannot arrive at the conclusion that the theoretically calculated thrust value and the thrust

experienced by the pump in practice are same. This is mainly due to the manufacturing

limitations and assembly variations. The existing methods cannot be employed in practice to

estimate the correct value of axial thrust or even the permissible variation. The available

literature on measurements of axial thrust in pumps indicates the lack of adequate and

accurate experimental results data. Axial thrust data and improved analytical methods are

Porto/Portugal, 22-27 July 2012

2

critical to the proper selection of thrust bearings due to the consequences of rapid wear,

frequent maintenance and possible pump failure due to large axial forces, it is better to predict

and minimize axial thrust for a range of pump types in a variety of applications.

2.0 NEED FOR OPTIMIZATION OF AXIAL THRUST

The life and size of the bearings in centrifugal process pumps is greatly influenced by the

axial thrust developed.

High axial thrust loads can cause rapid thrust bearing wear and either subsequent pump failure

or frequent overhauls. The axial thrust is balanced by various methods like balancing chamber

and holes, pump out vanes, balancing ring and disc and by using balancing drum. By reducing

the pressure in the space between the back shroud and the casing wall, the axial thrust is also

balanced.

The axial thrust is dependent on the many aspects viz. shroud and casing clearances,

peripheral shroud speeds, head developed by the pump, impeller geometry etc. due to this one

cannot ensure that the theoretically calculated thrust value and the thrust experienced by the

pump in practice are same. This is mainly due to the manufacturing limitations and assembly

variations. None of the existing method can currently be employed in practice to estimate the

correct axial thrust value or even the permissible variance.

The available literature on measurements of axial thrust in pumps indicates the lack of

adequate and accurate experimental results data. Axial thrust data and improved analytical

methods are critical to the proper selection of thrust bearings due to the consequences of rapid

wear, frequent maintenance and possible pump failure due to large axial forces, it is better to

predict and minimize axial thrust for a range of pump types in a variety of applications.

3.0 HYDRAULIC THRUSTS IN CENTRIFUGAL PUMPS

During operation and working of centrifugal pumps, as mentioned earlier, the kinetic energy

of flowing liquid is converted into pressure energy. This high pressure liquid is continuously

flowing all over the circumference of the impeller and also gets entrapped inside the

clearances between impeller and casing / casing cover. This high pressure liquid exerts

pressure on the outlet passages and shrouds of the impeller resulting in generation of two

forces, one in lateral and another in longitudinal direction with respect o shaft axis.

The force generated in lateral direction is due to dissimilar pressure generation in

volute and called as Radial thrust[fig1]

, while another one generated in longitudinal direction is

on account of different areas of impeller exposed to trapped pressurized liquid called as axial

thrust.

3.1 RADIAL THRUST

The hydraulic radial load is due to the unequal velocity of the fluid flowing through the

casing. The unequal fluid velocity results in a non-uniform distribution of pressure acting on

the circumference of the impeller. The radial load is most influenced by the design of pump

casing. The pump casing is designed to direct the fluid flow from the impeller into the

discharge piping. In a theoretical situation at Best Efficiency point (BEP), the volute casing

has a uniform distribution of velocity and pressure around the impeller periphery. There are

proven formulas to predict the hydraulic radial thrust with fair amount of accuracy. Radial


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thrust can be minimized by making double volute casing [fig1]

or by providing the diffuser type

casing.

Fig. 1 – Single volute and Double volute.

3.2 AXIAL THRUST

Axial thrust arises in the centrifugal pumps due to their asymmetry. The clearances between

casing cover and impeller back shroud, casing and impeller front shroud are filled with fluid

at delivery pressure. This pressure acts on the impeller shrouds. As the back shroud is having

larger surface area than front shroud, a net thrust acts on the impeller in the direction opposite

to that of the incoming flow [9]

. The force contributed by the change in the momentum of

incoming flow is also to be considered.

Figure 2 – Axial thrust components

As shown in Fig. 2 above the resultant unbalanced axial thrust is vector summation of the

following forces [1]

–

Force acting on front shroud due to liquid of delivery pressure entrapped between pump

casing and front shroud. (F1)

Force acting on back shroud due to liquid of delivery pressure entrapped between casing

cover and back shroud. (F2)

F1 F2

UNBALANCED THRUST

Fm


4

Force acting in the direction of the liquid flow due to its momentum change. (Fm)

The pressures generated by a centrifugal pump exert forces on both its stationary and rotating

parts. The design of these parts balances some of these forces, but separate means may be

required to counter-balance others. Axial hydraulic thrust is the summation of unbalanced

impeller forces acting in the axial direction. As reliable large-capacity thrust bearings are not

readily available, axial thrust in single-stage pumps remains a problem only in larger units.

Theoretically, a double-suction impeller is in hydraulic axial balance with the pressures on

one side equal to, and counter-balancing the pressures on, the other Fig 6. In practice, this

balance may not be achieved for the following reasons:

The suction passages to the two suction eyes may not provide equal or uniform flows to the

two sides.

1. External conditions such as an elbow being too close to the pump suction nozzle may

cause unequal flows to the suction eyes.

DISCHARGE

PRESSURE

SUCTIO

N

PRESSURE

DISCHARGE

PRESSURE

SUCTIO

N

PRESSURE

DISCHARGE PRESSURE

DISCHARGE PRESSURE

SUCTIO

N PRESSURE

Figure 3 – Origin of pressure acting on impeller shrouds to produce axial thrust.

2. The two sides of the discharge casing may not be symmetrical, or the impeller may be

located off-centre. These conditions will alter the flow characteristics between the impeller

shrouds and casing, causing unequal pressures on the shrouds.

3. Unequal leakage through the two leakage joints will tend to upset the balance.

Combined, these factors create definite axial unbalance. To compensate for this, all

centrifugal pumps, even those with double-suction impellers, incorporate thrust bearings.

The ordinary single-suction radial-flow impeller with the shaft passing through the impeller

eye Fig 3 is subject to axial thrust because a portion of the front wall is exposed to suction

pressure, thus exposing relatively more back wall surface to discharge pressure. If the

discharge chamber pressure were uniform over the entire impeller surface, the axial force

acting towards the suction would be equal to the product of the net pressure generated by the

impeller and the unbalanced annular area.


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4.0 Axial thrust balancing

The thrust magnitude will be approximately equal to the product of the net pump pressure

and the annular unbalanced area. Actually, the axial thrust turns out to be about 70% to 80%

of this theoretical value. The axial thrust can be safely carried out by a thrust bearing; this is

the most efficient way to reduce the thrust on the bearing. This can be done only at the

expense of the pump efficiency. Following are the different methods used to eliminate the

axial thrust of a single-suction impeller,

4.1 Providing wear rings and balancing holes: [1], [7], [9]

One of the method which is usually employed to reduce or eliminate axial thrust in single

stage pumps in a chamber on the back of the impeller is provided with a closely fitted set of

wearing rings and suction pressure is admitted to this chamber either by drilling holes through

the impeller back shroud into the eye or by providing a special channel connecting the

balancing chamber to the suction nozzle [fig 4]

.

The use of drilled holes through the impeller shroud to the balancing chamber is

inferior to the arrangement using a special channel to connect the balancing chamber with the

suction nozzle because leakage through the holes is directed against the flow in the impeller

eye, causing disturbances. The balance by this method is never complete. From 10 to 25 per

cent of the axial thrust always remains depending on the size of the holes. For a complete

balance the diameter of the wearing rings of the balancing chamber should be greater than that

at the impeller eye.

REDUCTION IN

AXIAL

THRUST DUE TO

BALANCE HOLES

IMPELLER

BALANCE HOLE

Figure 4 – Axial thrust reduction with balance holes

4.2 Balancing with radial back vanes: [1], [7], [9]

In this method, radial ribs are used on the back shroud to reduce the pressure in the space

between the impeller, and the pump casing. Figure 5 shows the schematic of an impeller

housed inside pump casing the radial back vanes provided at the back shroud of the impellers

acts as auxiliary impeller which restricts the entry of liquid into the clearances between

impeller back shroud and casing cover.

It is evident that the balancing with wearing and holes method introduces the leakage loss of

the pump which, in turn, increases as wearing rings are worn. Balancing with ribs method

requires some additional power which, however does not change with time. In addition it is


6

cheaper and more effective than the first method. In this method the power consumed is much

less than the power loss due to leakage through balancing holes under normal conditions.

BACK VANES

IMPELLER

REDUCTION IN

AXIAL

THRUST DUE TO

BACK VANES

Figure 5 – Axial thrust reduction with back vanes

There are some other thrust balancing methods like Balancing Disc and ring, Balancing Drum

used specifically to reduce the axial thrust in multistage centrifugal pumps.

5.0 EXPERIMENTAL ANALYSIS [14]

. –

The actual thrust induced on the impeller with back vanes will be measured by virtue of an

experimental test rig. This will include direct measurement of the thrust exerted on the

impeller by employing suitable transducer. The methods of measuring the unbalanced

hydraulic axial thrust experimentally are –

• Providing tapping in pump casing and casing cover for obtaining the pressure values

between the clearances of impeller and casing.

• By making a provision for load cell at the end of the shaft [14]

, this will directly give

the axial end thrust in Kg.

Finally the results obtained experimentally and those obtained by theoretical analysis were

compared to draw a conclusion.

5.1 Experimental analysis procedure[14]

.

The method of orthogonal array was used to conduct this experiment instead of making

simple iterations (conventional method) of predefined parameters. For conducting this

experiment conventionally it would have required 81 different trials based on three variations

of four parameters each, but by utilizing the orthogonal array method, the same results can be

obtained in three trials. With orthogonal array method, it is possible to analyze the effect of

influencing parameters on the axial thrust development quickly and with fewer trials.

In this technique it is proposed that an experiment organizes consecutive small series of trials

in each of which all the factors are simultaneously varied according to a definite rules. The

series are such arranged that after mathematical processing of a proceeding, it will be possible


ICEM15 7

to analyze the effect of each influencing parameters on the response value. For this project

following four parameters are chosen –

1. The number of back vanes,

2. The back vane radius,

3. The back vane height,

4. Clearance between casing cover and back vanes.

All the above factors have three variations each as given in Table-1.

Experiment

no

No of back

vanes

(Z)

Back vane

diameter

(D) mm

Back vane

thickness

(T) mm

Clearance between back

vane and casing

(S-T) mm

1 8 107.00 6.00 5.00

2 8 87.00 3.50 3.00

3 8 67.00 1.00 1.00

4 6 107.00 3.50 1.00

5 6 87.00 1.00 5.00

6 6 67.00 6.00 3.00

7 4 107.00 1.00 3.00

8 4 87.00 6.00 1.00

9 4 67.00 3.50 5.00

Table-1- Orthogonal array experiment design

5.2 Experimental setup (Measurement technique) [14]

.

These tests are been carried out on an end suction pump of 66 specific speed (metric) with 214

mm diameter. The test rig (fig6)

used consisted mostly of the test end suction pump long

coupled to electric motor running 2960 rpm, a modified side bearing arrangement consisting

of three strain gauge type load cells located at 1200 apart over the circumference and on the

both side of cartridge. The transfer mechanism for transmittal of axial thrust from rotating

members to the stationary load cells consists of two thrust ball bearings, which are mounted

on the impeller shaft. The shaft is supported by two radial bearings, with their outer races free

to slide inside the bearing housing. The load cells are pressed slightly against the modified

bearing bracket which comes in contact with the rotating assembly through the thrust ball

bearings. The load cells were used on either side of thrust bearings so as to respond to change

in the axial thrust direction.

The total axial thrust will be the algebraic sum of the forces experienced by all the three load

cells. To avoid any error, all the load cells were preloaded slightly and at the preloaded

condition, the null reading was set.

5.3 Calibration[14]

The test rig was statically calibrated before the start of experiment by pulling and pushing the

impeller shaft against known calibrated load values from UTM (Universal Testing Machine).

The load cell response (Eb mV) to the calibration null load was set based on the known static


8

load application, the graph-1 shows the response of the load cells output voltage with respect

to the known load applied.

Graph 1 – Load cell output voltage (Eb) Vs Load (kg)

LOAD CELL

(front)

LOAD CELL

(Back)

PUMP SIDEDRIVER SIDE

possible axial thrust direction

Fig-6 Experimental setup

5.4 Equations set used for theoretical analysis

The basic theory used to evaluate the axial thrust is taken from K J Zanker(1)

paper

specifically for the impellers with back vanes. The theoretical analysis parameters were also

measured during the tests performed by mean of tapings on pump casing at various locations.

The theory presented is as given below -

Force on front shroud (kg)

FR = π * ρ {g * Ho*(R2-r

2)-W

2/4(R

2-r1

2/2)

2}

Force on back shroud (with back vanes) (kg)

FB= π * ρ *{g HoR2

–w2/4(R

2-b

2/2)

2-gHoS1-W

2/8(R

2-b

2)*(b

2-S1

2)-B

2 (b

2-S1

2/2)

2}

The momentum thrust due to radial out thrust

Tm = (ρ*Q2)/A1. (kg)

Net Axial Thrust


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T = (Force on back shroud considering back vanes –pressure force on front shroud –

momentum thrust)

T = FB - FR - Tm (kg)

Where,

• W =Angular velocity of impeller (Rad/s)

• B = Angular velocity of fluid (Rad/s)

• ρ = Density (Kg/m3)

• H = Pump head (meters)

• Ho = Static head at the impeller outlet (meters)

• R = Impeller radius (meter)

• r1 = Impeller wear ring radius (meter)

• b = Radius of back vanes (meter)

• h = Hydraulic efficiency of pump

• Cd = Coefficient of contraction of jet flow at the impeller back vane

• u2 = Peripheral velocity at the impeller outlet (Rad/s)

• S = Axial gap between casing and impeller back shroud (mm)

• T = Vane width (mm)

• (s-t) =Vane clearance (mm)

• Z = Number of back vanes

• Re =Reynold’s number

• γ = Kinematic viscosity

(m2/sec)

• S1 =Radius of shaft (meter)

• Q = Discharge (m3/hr)

• ηh = Hydraulic efficiency of pump.

• N = speed in (rpm)

• A1 = Eye opening area (m2)

6.0 Observations [14]

1) The thrust data was obtained throughout the performance range (H-Q curve) of the

pump. It is observed that the parameter controlling axial thrust are as follows in order

of importance:

• Back vane radius

• Back vane height

• No. of back vane

• Clearance between casing cover and Impeller back Vane.


10

2) From graph it is observed that as number of back vanes increases the back vane force

developed also increases since the increase in number of back vanes increases the

angular velocity of the fluid between the back shroud of impeller and pump casing.

But after some limit, if no. of back vanes increased that will not be served. In these

experiments the optimum number of back vanes was found to be 8 [graph 2]

.

Graph 2- Variation of back vane force with number of back vanes

3) From graph-3. it is observed that back vane force increases as the radius increases. The

increase of radius results in increasing the rotation of the fluid. (Over a radial length).

A small increase in back vane radius affect the back vane force for a great extent. If

full diameter back vanes are provided it will rotate the total water trapped in the axial

gap. There by increasing the angular velocity of the fluid over the whole region of

impeller. If the back vanes provided are less than the impeller diameter, the fluid

trapped in the region of back vanes and casing will have the angular velocity higher

than the angular velocity of fluid trapped in the remaining region of the impeller

diameter. In this case the back vane force developed will be due to combined effect of

the two angular velocities which will result in the lower back vanes force.

Graph 3 – Back vane force Vs Back vane radius

4) If the axial gap is very small, the angular velocity of the fluid in the gap will be high

and hence the back vane force as well. Back vane height also helps to increase the

rotation of the fluid in the axial gap. The optimum value has been found to be 5 mm

from these experiments.

(Kg

)


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Graph 4 – variation of back vane force with back vane height

5) Graph 5 shows that the back vane force developed is higher for low clearance values

as well a with very high clearance values, when the clearance is very low the angular

velocity of the fluid is higher in this small gap. This will result in decrease in pressure

on the impeller shroud. As the clearance increases a very large value the eddy of

recirculatory flow will start through the axial gap. This will result in equalization of

pressure behind the impeller at the outer diameter and at the inner most diameters

naturally increasing the total back vane force.

Graph 5 – Axial thrust Vs Clearance between casing cover and Back vane face

6) The axial thrust variation with respect to flow for all 9 tests is as given below in graph

6 -

(Kg

)


12

Axial thrust Vs Flow

0

100

200

300

400

500

600

700

800

0 20 40 60 80 100

Flow (lps)

Ax

ial

Th

rus

t (k

g)

Expt 4

Expt 1

Expt 3

Expt 8

Expt 2

Expt 7

Expt 9

Expt 6

Expt 5

Graph 6 – Axial thrust Vs Flow

7.0 Conclusion

1. Maximum influencing parameter

From the experimental data, it is observed that the back vane radius contributes to the

maximum extent to the variation of axial thrust.

2. The order of importance of other back vane parameters is as follows:

Back vane height


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Number of vanes

Clearance between back vanes and casing cover

3. Close control of back vane parameters is essential to maintain optimum thrust values

for mass produced pumps. The variation in thrust values due to change in the

dimensions of back vane height and clearance has been tabulated[14]

Experiment no. (Refer

table 1)

Axial Thrust at BEP condition (kg)

Axial Thrust at 20% Left of BEP condition (kg)

Pump efficiency at BEP condition

Without

back vane

With back

vane

Without

back vane

With back

vane

Without

back vane

With back

vane

Test 1

Test 2

Test 1

Test 2

Test 1

Test 2

Test 1

Test 2

Test 1

Test 2

Test 1

Test 2

1 640 642.5 100 101.5 815 812 295 298 82.3 82 78 78.5

2 630 628.5 378 377 795 794 594 592 82.3 82 78 78.5

3 590 588 553 552 673 674 707 580 82.3 82 78 78.5

4 590 588 282 283.5 673 674 497 501 82.3 82 78 78.5

5 640 642.5 556 551 815 812 749 748 82.3 82 78 78.5

6 630 628.5 510 513 795 794 690 691.5 82.3 82 78 78.5

7 630 628.5 545 547.5 795 794 730 733 82.3 82 78 78.5

8 590 588 388.5 385 673 674 547 546 82.3 82 78 78.5

9 640 642.5 482.5 480 815 812 688 688 82.3 82 78 78.5

References

1) K J Zanker, “Experiments with back vanes used for balancing axial thrust on centrifugal

pump impellers”, British Hydromechanics Research Association, Vol. 3, 5th

Annual meet.

(Apr. 1962), pp.7-14. 2) George Schaefer and Eric Olson, ”Experimental Evaluation of Axial Thrust”, Journal of

World Pumps, Issue 393, June 1999, pp. 34-37.

3) Majan Gantar, Dr. Dussn Florjancic, Dr. Brane Sirok, “Hydraulic Axial Thrust in

Multistage Pumps - Origins and Solutions”, proceedings of ASME FEDSM’01, May 29 -

June1.2001, pp. 1-8.

4) Marc P. Mignolet Byeong-Keun Choi, “Robust Optimal Positioning of Strain Gages on

Blades”, journal of turbo machinery- transactions of ASME Vol. 125, October 2002. pp.

155-164.

5) J. Szwedowicz, S. M. Senn and R. S. Abhari, “Optimum Strain Gage Application to

Bladed Assemblies”, journal of turbo machinery- transactions of ASME Vol. 125, January

2003. pp. 606-613.

6) Igor J Karassik & Ray Cartor, “Centrifugal pumps”, Tata McGraw Hill Book company-

2nd

Edition, 1960. pp. 3.129-3.339.

7) A.J. Stepanoff, “Centrifugal and Axial flow pumps - Theory, Design & Application” ,

John Wiley and Sons – 2nd

edition – 1953. pp. 182-223.


14

8) Stephen Lazarkiewicz & Adam T. Troskolanski, “Impeller pumps”, Pergramon Press- 1st

edition 1965. pp. 346-365.

9) Val S. Lobanoff & Robert R. Ross, “Centrifugal Pumps design and application”, Jaico

publishing house – 2nd

edition 1995. pp. 333-353.

10) “Centrifugal pumps for petroleum, heavy duty chemical and gas industry services”,

American Petroleum institute – 610 10th

edition-October 2004

11) “General catalogue” SKF bearings.

12) “Bearings in Centrifugal Pumps – Application handbook – Part I &II”, SKF100955_1,

SKF bearings.

13) “ISO 281: Rolling Bearings - Dynamic Load Ratings and Rating Life.”, Second edition

1990-03-25.

14) “Experimentation on suction pump for axial thrust evaluation”, R&D, Kirloskar Brothers

limited, Kirloskarvadi.

axial thrust in centrifugal pumps - experimental

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