effects of discharge on the performance of an industrial...

Effects of Discharge on the Performance of an Industrial

Centrifugal Pump

Atiq Ur Rehman1, Subhash Shinde

2, Akshoy Ranjan Paul

3, Anuj Jain

4

1Research Scholar,

2Former M.Tech. Student,

3Assistant Professor,

4Professor

Department of Applied Mechanics, Motilal Nehru National Institute of Technology Allahabad, India

*Email for correspondence: [email protected]

Abstract:Three-dimensional, steady-state computational fluid dynamics (CFD) analysis of an industrial

centrifugal pump is carried out at five different discharge conditions in order to study the flow behavior in the

pump. The pump used for this purpose is single-stage, single-entry centrifugal pump having double-volute

casing. The computational results are compared against the experimental data and are found to be in good

agreement. Multiple reference frame technique (MRF) used to simulate the flow field in rotating impeller and

stationary casing of centrifugal pump in steady state condition.The pump flow mechanisms are studied in

terms of operating characteristics and plotting various contours. The computational results are compared

against the experimental results and found to be in good agreement. The velocity vectors show the re-

circulating flow pattern at a lower discharge.

Keywords: Centrifugal pump, Computational fluid dynamics (CFD), CFD-Analysis, Multiple reference

frame (MRF), Performance prediction.

1. Introduction

In today's era of cut throat competition among the centrifugal pump manufacturers, the major issues of

pump industry are better resource utilization and reduction in production time with superior product quality.

The flow behaviour in the centrifugal pump is indeed complex due to involvement of both stationary and

rotating parts. These facts have forced the researchers to look into the internal flow physics of the centrifugal

pump.

Experimental analysis of centrifugal pumps is an expensive and time consuming technique, as it

involves constructing and testing physical prototypes in a trial-and-error process, thus reducing the

profit margins of the manufacturers. For this reason, computational fluid dynamics (CFD) analysis with

suitable turbulence modelling currently has an edge over other techniques. CFD simulations can provide

quite accurate information on the fluid and flow behaviour in the turbomachines, and thus help the engineers

obtain a thorough performance evaluation of a particular design.

For turbomachinary analysis using CFD, three methods are used, namely- Multiple Reference Frame

method (MRF), the Mixing Plane method (MP) and the Sliding Mesh method (SM). The MRF method and the

MP method are basically steady state flow methods. While in the SM method, unsteady flow equations are

solved. In all three methods, the flow in the rotor is calculated in a rotating reference frame, while the flow in

the stator is calculated in an absolute reference frame. The cost of the unsteady method is, however, typically

30 to 50 times higher than the cost of the steady methods. So for practical applications, it is necessary to

know how far the steady techniques are realistic. Many researchers used the MRF method for performance

analysis of centrifugal pump [1-5].

Since the fluid is rotating around the axis of the pump, the fundamental equations of fluid dynamics must be

organized in two reference frames, stationary and rotating reference frames. To accomplish this, the MRF

ISBN 978-93-84422-40-0

Proceedings of 2015 International Conference on Computing Techniques and Mechanical Engineering

(ICCTME'2015)

Phuket, October 1-3 2015, pp. 24-32

http://dx.doi.org/10.17758/UR.U1015114 24

mailto:[email protected]

model is used. The basic idea of the model is to simplify the flow inside the pump into an instantaneous flow

at one position in order to solve unsteady-state problem with a cost-effective steady-state computational

technique. In this approach, it can be concluded that the flow pattern of a centrifugal pump can be described

quite well with the MRF method [6-8].

Over the past few years, with the rapid development of the computer technology and CFD, the

computational flow analysishas become an important tool to study flow field in pumps and to predict pump

performance. Therefore, in order to predict performance of pumps exactly, flow field in pump must be

obtained correctly. In the present study, the CFD analysis of centrifugal pump is carried out to study the

complex flow behaviour of centrifugal pump at five different discharge conditions.

2. Geometry and Grid Generation

The geometry of the centrifugal pump is modelled using a CAD Software (CATIA) shown in Fig 1.The

flow domain of centrifugal pump is shown in Fig. 2. It consists of three zones namely suction pipe, impeller

and pump casing. The impeller is single-stage, single-entry, radial type and has twisted blades. The casing

modeled here is a double-volute. The major dimensions of the pump are furnished in Table- I.

Fig. 1: CAD geometry of Centrifugal pump Fig. 2: Geometry of centrifugal pump

TABLE I:DIMENSIONS OF CENTRIFUGAL PUMP

Quantity Values Units

Pipe length 400 mm

Pipe diameter at inlet 150 mm

No of blades 3

Impeller inlet diameter 156 mm

Impeller outlet diameter 399 mm

Outlet blade angle 16 degree

Outlet blade width 23 mm

Speed of impeller 2900 rpm

For present simulation is generated on Ansys ICEM-CFD software using tetrahedral method with patch

independent algorithm. Before simulation, grid independency test is performed to ensure the CFD solution is


independent of grid size and is shown in Fig.3. The variation in head is found to be negligible after 4164202

elements. Hence the total number of elements is chosen as 4164202 for further CFD analysis. The generated

grid corresponding to 4.16 million elements is shown in Fig.4.

Poor quality grid causes inaccurate solution or slow convergence. So, creation of good quality mesh is

important in getting the proper and quick results. Though there are lots of quality criterion are defined for

generated mesh, skewness, aspect ratio and orthogonality are the main parameters to check. In the final

meshing, sknewness of only less than 1% element is found 0.85, maximum aspect ratio is 11.18 and

orthogonality of grid is 0.99.

Fig. 3: Grid independency plot Fig. 4: Computational grid in pump

3. Mathematical Modelling

3.1. Governing Equations

For the three dimensional steady state analysis of centrifugal pump, the equations for continuity and

momentum conservation Navier-Stokes equations for rotating coordinate an incompressible fluid are

described as below.

0u j

t x j

(1)

xu ji

ex xu uu j ij i pi

Sit x x xj i i

(2)

3.2. Turbulence Modelling

For the present analysis standard k turbulence model is used. It is mentioned that the convergence is found

to be better with standard k than any other for centrifugal pump application [9]. The equation for turbulence

kinetic energy k and dissipation rate are as follows, on the right-hand side of Eq. (2), there is the effective

Viscosity

e t (3)

212.2

209.6

208.3 207.8

205.3 205.2

204

205

206

207

208

209

210

211

212

213

0 1 2 3 4 5

Hea

d (

m)

Mesh elements (Million)


2

tC k

(4)

It is directly related to k and as shown in equation (4). The transport equation for k and ε are described as

j jtk

j j k j j

ku uk kP

t x x x t x

(7)

(7)

2

1 2

j ji i tt

j j i j j j

u uu uC C

t x k x x x k x x

(8)

The model constant has the following default values

1 20.09, 1.44, 1.92, 1.0, 1.3kC C C

3.3. Boundary Conditions The velocity inlet is considered at inlet boundary condition while mass flow rate is imposed at outlet. The

inlet pipe section and casing sections are considered as stationary reference frames while impeller is

considered as rotating reference frame these frames are coupled through frozen-rotor interface model. Walls

associated with impeller are considered as rotating walls while those associated with casing are considered as

stationary walls. Details of used boundary conditions are tabulated below in table II.

TABLE II: BOUNDARY CONDITIONS

Discharge (m3/h) Speed(rpm) Velocity at inlet (m/s) Mass flow rate (kg/s)

83.63

144.86

2944

2939

1.32

2.28

23.19

40.17

212.05

242.8

2949

2933

3.34

3.82

58.81

67.32

258.74 2933 4.07 71.74

3.4. Solution Methodology For the three dimensional steady state numerical analysis of centrifugal pump using Ansys CFX-14 [10]

domains is pipe, impeller and casing are assigned properly to their locations. Each domain is considered as

fluid domain, pipe and casing are considered as stationary domains while impeller is considered as rotating

domain. The material is selected as water for all domains from material library and it is considered as

continuous fluid. The reference pressure is set equal to one atmospheric pressure. Fluid model for heat transfer

is kept isothermal with 25 C temperatures and turbulence is modelled using standard k model with scalable

wall function. Two interfaces are defined; one interface is in between pipe and impeller while other is between

impeller and casing. Frame change option is selected as frozen rotor pitch change is set as automatic. The

considered interfaces are of fluid-fluid type.

Though the rotational speed of five considered discharge conditions is different, they represent the specific

flow rate condition. Flow rate 144.86 m3/hour represents the low flow rate condition. Flow rate 242.8 m

3/hour

is the flow rate corresponding to best efficiency point, while flow rate 258.74 m3/hour represents the operating

point condition near the best efficiency point.

4. Results and Discussion

The computationally obtained characteristic parameters of centrifugal pump are compared against the

experimentalresults. The main characteristic parameters are total head, total efficiency, input power and output

power of the pump. The validation study is performed for five discharge conditions obtained during the

experimental results and are shown in Figs. 5-8. Error statistics shows that the maximum error between

computational result and experimental results occurred at low flow rate condition(144.86 m3/hour). The head


comparison shows the maximum error of 2.78% while efficiency and input power comparison shows

maximum error of 5.22% and 7.71% respectively. The computational work over-predicted the head value and

it is accepted as the head loss due to various reasons such as leakage, friction etc. is not considered during

calculation of head value. Disc friction losses are considered during calculation of efficiency value and hence

the under-predication. The computational results are therefore considered to be in acceptable for further

analysis.

Fig.5: Head vs discharge Fig. 6: Efficiency vs discharge

Fig.7: Input power variation over discharge Fig. 8: Output power variation over discharge

The contours of static pressure are shown in Fig.9 for five different considered positions. From the static

pressure contours, it is observed that the pressure drop between outlet and inlet is more for low flow rate

condition i.e. at 144.86 m3/hour. The distribution of pressure in the upper section of the volute is found to be

uniform with no pressure gradient for all five discharge conditions. While for low flow rate condition there

existed a pressure gradient in volute section surrounding the impeller. This is found to be due to the reversal of

the flow. This reversal of the flow at low flow rate condition is shown in vector diagram in Fig.10.


144.86 m

3/h 242.8 m

3/h 258.74 m

3/h

Fig. 9: Static pressure contours

The inward direction of arrows in Fig. 10 clearly indicates the reversal of the flow for low flow rate

condition while such a kind of huge reversal of the flow is not detected for flow rates at 242.8 m3/hour and

258.74 m3/hour.

144.86 m

3/h 242.8 m

3/h 258.74 m

3/h

Fig. 10: Reversal of the flow

The huge reversal of the flow at 144.86 m3/hour caused the large amount of flow separation and is shown

through the streamline pattern shown in the fig. 11. The flow separation is observed at both tongue positions

and for all five discharge conditions. The amount of flow separation for 242.8 m3/hour and 258.74 m

3/hour

discharge conditions is less than that for 144.86 m3/hour condition. The flow is found to be attached more to

the suction side of the blades than the pressure side.

The variation of static pressure, total pressure and velocity from pipe inlet to casing outlet is studied by

computing the values of these parameters at different locations such as suction pipe inlet, impeller inlet,

impeller outlet and casing outlet and are designated by numbers 1, 2, 3, 4 respectively. These variations are

shown in Figs. 12-14. Static pressure distribution in Fig. 12 indicated the continuous increase in the pressure

from impeller inlet to casing outlet for all considered positions but the magnitude of the pressure is different at

different considered flow rate conditions. The highest static pressure at impeller periphery is recorded for flow

rate 83.63 m3/hour.


144.86 m3/h 242.8 m

3/h 258.74 m

3/h

Fig. 11: Streamline pattern for different flow rate conditions

Total pressure in Fig. 13 is found maximum at impeller outlet and decreased in casing portion gradually due to

the fact of decrease of velocity. Velocity loss also takes place in the volute due to continuous increase in the

casing area and is shown in Fig. 14.

Fig. 12: Static pressure variation Fig.13: Total pressure variation Fig.14: Velocity variation

4. Conclusions

The three-dimensional numerical analysis of pump is carried out at five different discharge conditions

using steady-state CFD approach. The pump flow mechanisms are studied in terms of operating characteristics

and plotting various contours. The computational results are compared against the experimental results and

found to be in good agreement. The velocity vectors show the re-circulating flow pattern at a lower discharge

of 144.86 m3/hour. The steady state analysis with MRF technique is found sufficient to analyze the flow

behavior at different discharge conditions.

5. Acknowledgement

The authors extend their sincere thanks to Bharat Pumps and Compressors Limited, Naini, Allahabad

(India) for providing the necessary drawing details and experimental data of an industrial centrifugal pump.

6. References

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Journal of the Serbian Society for Computational Mechanics , 8 (1).

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Journal of EEA,Vol. 28.

[3] S.R. Shah, S.V. Jain and V.J. Lakhera, “CFD based flow analysis of centrifugal pump”, 37th National & 4th

International Conference on Fluid Mechanics and Fluid Power, 2010, IIT Madras, Chennai, India.

[4] M.L. Hedia, K. Hatema and Z.Ridhaa (2012). Numerical Analysis of the Flow through in Centrifugal Pumps.

International Journal of Thermal Technologies, International Journal of Thermal Technologies, ISSN 2277 – 4114.


[5] L.G. Das, M.K. Rawat and N. Kuri, "Flow Analysis of a Centrifugal Slurry Pump While Handling Clear Water at

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[6] Shahin, I. AbdEelganny.M, Abdellatif, O. E, Ayad, Samir S. AbdrRabbo, M.F. (2010). “Performance and Unsteady

Flow Field Prediction of a Centrifugal Pump with CFD Tools” Tenth International Congress of Fluid Dynamics

December 16-19, Stella Di Mare Sea Club Hotel, AinSoukhna, Red Sea, Egypt.

[7] L. Tan, B. Zhu, S. Cao, H. Bing and Y. Wang (2014). “Influence of Blade Wrap Angle on Centrifugal Pump

Performance by Numerical and Experimental Study” Chinese journal of mechanical engineering Vol.27, No.1.

[8] M.Djerroud, G.D. Ngom W. Ghie (2011). Numerical Identification of Key Design Parameters Enhancing the

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[9] S. Kumar, S.K. Mohapatra and B.K.Gandhi (2013). Investigation on Centrifugal slurry pump performance with

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[10] Ansys Fluent Theory Guide. Release 14.0, November 2011.


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