INVESTIGATION ON CENTRIFUGAL SLURRY PUMP PERFORMANCE WITH VARIATION OF OPERATING SPEED. sq sp

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1 International Journal of Mechanical and Materials Engineering (IJMME), Vol. 8 (2013), No. 1, Pages: INVESTIGATION ON CENTRIFUGAL SLURRY PUMP PERFORMANCE WITH VARIATION OF OPERATING SPEED S. Kumar 1*, S.K.Mohapatra 1 and B.K.Gandhi 2 1 Mechanical Engineering Department, Thapar University, Patiala, India 2 Mechanical and Industrial Engineering Department, IIT Roorkee, India * Corresponding author s satish.kumar@thapar.edu Received 19 October 2012, Accepted 29 March 2013 ABSTRACT In this paper three-dimensional fluid flow behavior of centrifugal slurry pump has been studied using commercial Computational Fluid Dynamics (CFD) code FLUENT at design and off-design conditions. Steady state simulation with Moving Reference Frame (MRF) model is used to consider impeller-volute interaction. Different turbulence models namely, standard k-, RSM, k- and RNG k- are applied for simulation of flow through the pump, which show reasonably close prediction of head flow characteristics of pump by k- model. Performance characteristics of the pump is numerically predicted at four different operating speeds namely 1000 rpm, 1150 rpm, 1300 rpm and 1450 rpm with water. The numerical results are compared with the experimental measurements. The comparison indicates that the specific head, specific power and efficiency characteristics prediction are within an error band of 5 %. Simulation results showed that standard affinity relations are applicable to the slurry pump also. Key words: Centrifugal slurry pump, Performance characteristics, Numerical simulation, Affinity laws. Notation D Impeller diameter, mm F Body forces, N g Gravitational acceleration, m/ s 2 H Total head, m I Unit vector N Impeller speed, rpm P out Output power, kw P in Input power of pump, kw P d g P s g P sd g p ss g Q sh Experimental static pressure at outlet passage, m Experimental static pressure passage, m Static pressure simulation, m at inlet at outlet passage through Static pressure at inlet passage through simulation, m Flow rate, m 3/ sec Specific head sq sp T V d Vs Z d Z s Specific discharge Specific power Torque of pump shaft, N-m Velocity of fluid at outlet passage,m/s Velocity of fluid at inlet passage,m/s Relative velocity,m/s Height of the delivery pressure transducer from ground,m Height of the suction pressure transducer from ground,m ρ Mass density of water, kg /m 3 μ Molecular viscosity, N s/m 2 Angular velocity vector of rotating frame, rad/s Efficiency of pump 1. INTRODUCTION Centrifugal slurry pumps are used to transport solid particulate materials for a wide range of applications such as dredging, dewatering of open mines, transportation of solids in mineral plants, disposal of ash in thermal power stations etc. They are popular due to consistent flow rate, low maintenance cost and excellent stability. The design of a centrifugal pump for slurry handling system need special consideration to ensure that the flow passages do not offer any restriction to the passage of solids. The efficiency of centrifugal slurry pump is lower as compared to that of a conventional centrifugal pump because of robust impeller and nearly concentric casing design, large running clearances and relatively wide throat impeller clearance. It may not be feasible to evaluate the performance characteristics of every pump at all the rotational speeds whereas the pump need to be operated at different value of rotational speeds in the field. The head developed with water at any operating speed is estimated using affinity relationship as under Specific head gh sh (1) 2 N 2 D Specific discharge Q (2) sq 3 N D Specific power Pin sp (3) 5 N 3 D 40

2 Equations (1) to (3) are dimensionless parameters applicable to conventional centrifugal pump handling water Stepanoff (1957) and Wilson et al. (1992). Attempts have been made by researchers (Sellgren, 1979; Gandhi et al., 1998; Gandhi et al., 2001 and Gandhi et al., 2002) to investigate experimentally the applicability of the affinity relations [equation 1-3] for centrifugal slurry pumps, because these pumps are geometrically different from the conventional pumps. During the last few years, design and performance analysis of turbo machinery have experienced great progress due to availability of less expensive high performance computers and user friendly computational fluid dynamics software (CFD). The numerical simulation can provide information on the fluid flow behavior inside the pump, and thus helps the engineer to obtain thorough flow analysis of a particular pump. Also the cost and time of the trial-and-error process by fabricating and testing of the prototypes pumps reduces the profit of the pump manufacturers. For this reason, CFD analysis is currently being used in hydrodynamic design for many different pump types Hornsby (2002) and Cao et al. (2005). Effort has been made by many researchers in the past few years to simulate the flow field of a conventional centrifugal pump to establish its performance characteristics with water. Many researchers Gayo et al. (2002), Gonz alez et al. (2002), Byskov et al. (2003), Asuaje et al. (2005), Gonzalez and Santolaria. (2006) have studied the performance analysis of conventional centrifugal pumps using CFD code. These researchers predicted the pressure and velocity distribution, reverse flow inside the impeller and casing at design and off-design conditions. Byskov et al. (2003) also validated the design and off-design simulation results with the measurements by particle image velocimetry (PIV) and laser Doppler velocimetry (LDV) systems. Asuaje et al. (2005) carried out the numerical simulation of centrifugal pump using CFX code with different turbulent models and found that standard k-ε model shows reseanable agreement with the experimental results. Some of the investigators Zhou et al. (2003), Cui et al. (2006), Liu.et al. (2010) and Jafarzadeh at al. (2011) have studied the influence of performance characteristics of the centrifugal pump by changing the number of impeller blades of the impeller. It is found that the optimum number of impeller blades produces maximum head and efficiency of the pump.it is also observed that the position of blades with respect to the tongue of volute has play a role in development of the separation. The review of literature shows that the performance of a conventional pump has been predicted successfully using numerical technique and so far such technique has not applied for centrifugal slurry pump. The present work has been carried out to investigate the flow characteristics of a centrifugal slurry pump using a commercial code FLUENT. Different turbulence models k-ε, k-ώ, RSM and RNG k- have been used and the results evaluated by each of turbulence models are compared. The steady state simulations have been carried out using moving reference frame (MRF) with rotorstator interaction. The performance characteristics of the centrifugal slurry pump with water have been evaluated at four different speeds namely 1000, 1150, 1300 and 1450 rpm. It has been found that predicted result by standard k- models are in reasonable agreement with the experiment results compared to the other turbulence models. Simulation results are also used to evaluate the applicability of affinity relation for specific discharge, specific head and specific input power of the centrifugal slurry pump handling water. 2. EXPERIMENTAL PROGRAMME Figure1 shows the photographic view of the experimental setup used for the present work to evaluate the performance of the pump at different speeds. Magnetic Flow Meter Sampler Delivery Valve Separator Centrifugal Slurry Pump Stirrer Mixing Tank Figure 1 Photographic view of experimental set-up. Deflector Measuring Tank The water or slurry is drawn from a hopper shaped mixing tank by the pump and returned back to either the mixing tank or a measuring tank (used for discharge measurement). The slurry for the tests is prepared in the mixing tank having a suitable stirring arrangement to keep the slurry continuously in mixed state during experiments. For continuous measuring of the flow, the 41

3 electromagnetic flow meters were installed in the vertical pipe section of the test loop. Measuring tank is placed near the mixing tank. Both tanks are provided with drain plug. Flow deflector is provided in-between mixing and measuring tank to divert the fluid during the calibration of flow meters. The pump delivers the water or slurry back to the mixing tank for recirculation. Separators are installed in suction and delivery sides of centrifugal slurry pump. The purpose of a separator is to eliminate entrance of solid particles in the tube connecting to pressure transmitter. Pressure transducers are attached to measure the suction and delivery pressures of the pump. Pump is driven by 7.5kW, 440V AC variable speed induction motor. For variation of the motor speed, a frequency modulator is used. The pump input power is determined indirectly with the measurement of torque and speed at pump shaft by using a torque transducer. The nominal range of the torque transducer is 0.1 Nm to 200 Nm with the accuracy of ±0.2% and that of speed is ±1 rpm. Pressure transducers installed at suction and delivery sides have the accuracy of 0.75%.The accuracy of the electromagnetic flow meter is ±0.25% of the full scale.total head developed by the pump is calculated from the measured pressures at inlet and outlet of the pump, velocity head and elevation of pressure transmitters. Equation (4) to (7) is used to calculate the performance characteristics of the pump. 2 2 Pd ps Vd Vs H ( Z d ) ( Z s ) g g 2g 2g (4) (The pressure transducers are attached at the same level so that Z d =Z s ) The geometry of impeller, volute casing and inlet passage are modelled by using pre-processor Gambit A schematic diagram of the pump impeller and casing assembly is shown in Figure 2 and some of the geometrical details of the pump are given in Table 1. The fluid volume of the internal fluid passages has been described using Gambit and grid quality was checked. Since the geometry of the pump is very complex, combined unstructured hexahedral and tetrahedral mesh was used for the grid generation Equi angle skew coefficient of the grid was kept less than 0.82, and aspect ratio was less than Meshed pump assembly is shown in Figure 3 and detailed of the pump component with different mesh size is given in Table 2. Figure 2 Centrifugal slurry pump assembly. P in =T ώ =2π NT/60 (6) (5) Efficiency of pump (7) 3. NUMERICAL MODEL The geometry is obtained from measurement of an available 50M WILFLEY centrifugal slurry pump at IIT Roorkee on which experiments have been performed. The pump impeller consists of five identical blades enclosed in a volute casing. Table 1 Geometrical details of centrifugal slurry pump Inlet passage radius 50 mm Outlet Passage radius mm Inlet Impeller width mm Outlet Impeller width mm Blade number 5 Blade thickness mm Base volute radius mm Volute width 85 mm Impeller inside radius 55 mm Impeller outside radius mm Figure 3 Meshed pump assembly. 4. GOVERNING EQUATIONS AND BOUNDARY CONDITIONS The continuity equation and momentum equations for steady state, incompressible fluid flows are given below (Versteeg and Malalasekera., 1995). ( ) 0.( ) p.( ) g F (8) (9) 42

4 Mesh type Component Table 2 Mesh Quality of the centrifugal slurry pump component. Mesh Size (mm) Tetrahedral element Hexahedral element Total element Eque Size Skewness Aspect ratio A Inlet Passage Casing Impeller B Inlet Passage Casing Impeller C Inlet Passage Casing Impeller The stress tensor is given by, T 2 [( )]. 3 I The absolute velocity and relative velocity are related as (10) Left hand side of momentum equations appears as follows for an inertial reference frame: (11) For a rotating reference frame, left hand side written in terms of absolute velocity becomes (12) In terms of relative velocity, left-hand side is given as ( ) Where, coriolis force is (13) The calculation procedure of the total head using simulation results is given below. 2 2 Psd pss Vd Vs H ( ) z g g 2g 2g (14) Developed head, input power and efficiency of the pump are calculated using Equations (5) to (7). Water is taken as working fluid. Multiple reference frame (MRF) model is used to simulate the flow through the rotating impeller and stationary volute casing. The relative velocity formulation is used to consider the Coriolis Effect. A SIMPLE scheme is used for pressure velocity coupling. To minimize the computational time, under-relaxation factor used are 0.2 for pressure, 0.7 for momentum equation, 0.8 for turbulence kinetic energy and 0.8 for turbulence dissipation rate. The mass flow rate is applied as inlet boundary conditions and the static pressure is specified at the outlet. 5. GRID INDEPENDENCY TEST In the present work three types of mesh namely type-a, type-b and type-c have been taken for checking of the grid independence test. The details of mesh quality inside the pump components for all the three types of mesh are given in Table 2. These simulations results are compared at best efficiency point of the pump with those obtained by the experimental measurement at 1450 rpm. The total pressure head developed by the pump with simulation of all the three types of the mesh is listed in Table 3. Description Experimental data at B.E.P Table 3 Grid independency test. Flow rate (kg/s 13.2 Head (mwc) % Deviation Mesh type-a Mesh type-b Mesh type-c Table 4 Simulation with different turbulence models. Turbulence model Total head (mwc) Experimental data at B.E.P % deviation Standard k Standard k-ώ RSM RNG, k The deviation between predicted head and experimental head with type-a mesh is 6.9%, type-b mesh is 5.25%, type-c mesh is 3.1%. The mesh type-c shows the closer 43

5 Specific head result with the experimental data and, used for present work. The different standard turbulence models k-ε, k-ώ RSM and RNG k- have been used and the results evaluated by each of turbulence models at best efficiency point of 1450 rpm speeds are compared with experimental result, shown in Table 4. From the Table -4 it is observed that the deviation in head with standard k- model, k-ώ model, RSM model and RNG, k- model 3.09%, 7.27%, 8.41% and 6.13%, respectively. The result shows that the lowest deviation in head is observed with standard k- turbulence model. In case of k -ε model, two additional transport equations are solved, but in the RSM model, seven additional transport equations need to be solved in 3D (Asuaje et al., 2005).Since k ε has lower number of transport equations to solve, it needs lower memory compared to RSM and other models. Aside from the time per iteration, the choices of turbulence model also affect the ability of FLUENT to obtain a converged solution. CFD simulation is carried out with all the turbulent models on the same computer systems. Simulation time for k-ώ model, RSM model and RNG k- is around 1.7, 2.8, and 2.5 times of that to the standard k- model. Thus the standard k- model requires lowest computational time compared to the other models and also results in better predictions. Many researchers Miner (2000), Tamm et al. (2001), and Zhou et al. (2003) have also used the k- ε turbulent model for CFD analysis of conventional centrifugal water pump. 6. RESULTS AND DISCUSSION The performance characteristic of the centrifugal slurry pump evaluated at different operating speeds 1000, 1150, 1300 and 1450 rpm for handling water is carried out using CFD simulation with standard k- turbulence model scheme. The performance characteristic results obtained in terms of head, power and efficiency which was used to calculate the specific discharge, specific head and specific power, and compared with the experimental results. The results at all the operating speeds as shown in Figures 4-9. Figure 4 and 5 shows the numerical and experimental results of the specific head characteristics of the pump at all the operating speeds. The trend of the specific head curve for both numerical and experimental results are decreasing with the increasing the specific mass flow rate at all the operating speeds.the maximum deviation of the specific head developed experimentally and numerically is 2.99% at 1450 rpm,2.38% at 1300 rpm,3.49% at 1150 rpm and 4.15% at 1000rpm. Figure 6 and 7 shows that the variation of specific input power of the pump with specific flow rate. It is observed that the numerical results follow the same trends as experimental results and input power increases linearly with the specific flow rate at all the operating speeds. The maximum deviation of the specific input power is 4.48% at 1450 rpm, 4.9% at 1300 rpm, 4.7% at 1150 rpm and 4.79% at 1000rpm. Figure 8 and 9 shows the variation of efficiency of the pump with flow rate.it is observed that numerical results follows the same trends as that of experimental results at all the operating speeds. Specific head Specific head at 1450 rpm Specific head at 1300 rpm Specific head at 1150 rpm Specific head at 1000rpm Specific flow rate Figure 4 Simulation specific head curve. Specific head at 1450 rpm Specific head at 1300 rpm Specific head at 1150 rpm Specific head at 1000rpm Specific flow rate Figure 5 Experimental specific head curve. It is seen that the maximum pump efficiency is46.70% numerically and 45.02% experimentally at 1450 rpm, 44.80% numerically and % experimentally at 1300 rpm, 44.6% experimentally and % numerically at 1150 rpm, 41.8% experimentally and 42.6% numerically at 1000 rpm. The above results show that pump efficiency calculated numerically is closer to the experimental value. The maximum deviation of the numerical and experimental evolution of the pump in the specific head, specific power and efficiency are within 5%. These simulation data shows reasonable agreement with the experimental data at the same speeds. Numerical and experimental studies are carried out to study the effect of speed on the performance characteristics of the centrifugal slurry pump with water. 44

6 Efficiency Specific power Efficiency Specific power Specific speed at 1450 rpm Specific speed at 1300 rpm Specific speed at 1150 rpm Specific speed at 1000 rpm Specific flow rate Figure 6 Simulation specific power curve. Specific power at 1450 rpm Specific power at 1300 rpm Specific power at 1150 rpm Specific power at 1000 rpm Specific flow rate Figure 7 Experimental specific power curve. From the Figure 4 and 5, it is seen that the specific head decreases with increase in the specific flow rate at all speeds both numerically and experimentally. It is also seen that there is no significant effect on these characteristics with pump speed. The maximum deviation in the specific head due to change in the pump speed is within 4.3% numerically and 4.6% experimentally for the entire operating range. Hence, it can be concluded that the parameters defined for head and capacity of the conventional pumps are also applicable to the slurry pumps with water despite the constructional differences. Similar variation has also been reported by Gandhi et al. (2002). The variation in specific power with specific flow rate is shown in Figure 6 and 7 and shows that almost linear increases the specific power with the specific rate at all speed. Similar variation has also been reported by experimentally Gandhi et al. (2002). However, the specific power at any specific flow rate decreases with increase in the pump speeds, the decrease being almost linear with the speed. This phenomenon can only be explained on the basis of losses taking place in the pump. The overall losses appear to reduce with increase in the speed. The magnitude of the maximum specific power decreases by 15.83% numerically, 16.65% experimentally and for a 45% increase in the pump speed from 1000 to 1450 rpm. The above observation shows that the use of an affinity relation for estimating the input power with water will result in a significant error for large variation in the pump speed Efficiency at 1450 rpm Efficiency at 1300 rpm Efficiency at 1150 rpm Efficiency at 1000rpm Flow rate (littre/sec) Figure 8 Efficiency -curve (Simulation). Efficiency at 1450 rpm Efficiency at 1300 rpm Efficiency 1150 rpm Efficiency at 1000rpm Flow rate(littre/sec) Figure 9 Efficiency -curve (Experimental). 6.1 Static Pressure and Velocity Distribution The CFD simulation shows a virtual image of the internal flow in the system allowing the analysis of more complex phenomena. Pressure distribution of the flow field in the pump at best efficiency point (Q/Q bep =1) and (Q/Q bep =0.2) at 1450 speed are presented in figure 10 and velocity distribution shown in Figure 11. From the figure 10, it is observed that static pressure gradually increases from impeller inlet to outlet at both mass flow rate conditions. At the impeller-volute interaction section non uniform pressure distribution is observed. The static pressure on diffusion section of volute outlet increases at small flow rates (Q/Q bep =0.2) while the static pressure on the same place decreases clearly at larger flow rate. From the figure11 it is observed that the good guidance of the flow in the volute section but recirculation zone appears at the diverging section of the volute casing at 20% of the best efficiency point flow rate. Inside the impeller flow velocities are 45

7 relatively uniform for all blade passages but backflow occurred near the pressure surface of the pump impeller at 20% of the best efficiency point flow rate. (a) (Q/Q bep =0.25) (b) (Q/Q bep = 1) Figure 10 Pressure distribution of pump at 1450 rpm. (a) (Q/Q bep =0.25) (b) (Q/Q bep = 1) Figure 11 Velocity distribution of pump at 1450 rpm. 7. CONCLUSION A large amount of numerical simulations were carried out in order to evaluate the performance characteristics of the centrifugal slurry pump. The following conclusions can be drawn based on the present investigation: Standard k- shows reasonable agreement with the experimental results as compared to the other turbulence models. The deviation of predicted and experimental specific head, specific power and total efficiency are less than 5%.Thus CFD results show satisfactory agreement with experimental data in a complete operating range of the pump. From the simulation studies of the pump characteristics at different pump speeds, it is concluded that the affinity relations applicable to conventional pumps for head and capacity can be applied to the slurry pumps handling water. The present study has demonstrated that the numerical method in this paper produces a good prediction of the centrifugal pump performance and can be applied in practice. ACKNOWLEDGEMENT The authors are grateful to Mr. Nitin Kumar, M.E student, Mechanical Engineering Department, Thapar University, Patiala for providing much useful information with regards to the design and performance of centrifugal slurry pump. REFERENCES Asuaje, M., Bakir, F., Kouidri, S., Kenyery, F. and Rey, R Numerical modelization of the flow in centrifugal pump: volute influence in velocity and pressure fields, Int. J. Rotating Machinery 3: Byskov, R.K., Jacobsen, C.B. and Padersen, N Flow in a centrifugal pump impeller at design an offdesign conditions-part Ⅱ: large eddy simulations, Journal of Fluids Engineering 125: Cao, S. and Peng, G Hydrodynamic design of rotodynamic pump impeller for multiphase pumping by combined approach of inverse design and CFD analysis, ASME Trans. J. Fluids Engineering 127: Cui, B., Zuchao, Z., Jianci, Z. and Ying, C The flow simulation and experimental study of lowspecific-speed high-speed complex centrifugal impellers, Chinese Journal of Chemical Engineering 14 (4): Gandhi, B.K., Singh, S.N. and Seshadri, V Prediction of performance characteristics of a centrifugal slurry pump handling clear liquid, Indian Journal of Engineering & Material Science 5: Gandhi, B.K., Singh, S.N. and Seshadri, V Performance characteristics of centrifugal slurry pumps, Transactions of ASME Journal of Fluid Engineering 123: Gandhi, B.K., Singh, S.N. and Seshadri, V Effect of speed on the performance characteristics of a 46

8 centrifugal slurry pump, Transactions of ASCE Journal of Hydraulic Engineering 128: Gayo, J.L., Gonz alez, J. and Fern andez-francos, J The effect of the operating point on the pressure fluctuations at the blade passage frequency in the volute of a centrifugal pump, Transactions of the ASME Journal of Fluids Engineering 124 (3): Gonz alez, J., Fern andez-francos, J., Blanco, E. and Santolaria, M.C Numerical simulation of the dynamic effects due to impeller-volute interaction in a centrifugal pump, Transactions of the ASME, Journal of Fluids Engineering 124 (2): Gonzalez, J. and Santolaria, M.C Unsteady flow structure and global variables in a centrifugal pump, Journal of Fluids Engineering 128: Hornsby, C CFD driving pump design forward, World Pumps: Jafarzadeh, B., Hajari, A., Alishahi, M.M. and Akhbari, M.H The flow simulation of a low-specificspeed high-speed centrifugal pump, Journal of applied Mathematical Modeling 35: Liu, H., Yong, W., Shouqi,Y., Minggao,T. and Kai, W Effects of Blade Number on Characteristics of Centrifugal Pumps, Chinese Journal of Mechanical Engineering 23: 1-5. Miner, S.M Evaluation of Blade Passage Analysis Using Coarse Grids, ASME Journal of Fluids Engineering 122: Sellgren, A Performance of a centrifugal pump when pumping ores and industrial minerals, Hydro transport 6, BHRA Fluid Engineering, Canterbury, UK, Paper G1. Stepanoff, A.J Centrifugal and axial flow pumps, Theory, design and application, John Wiley& Sons, New York. Tamm, A., Ludwig, G. and Stoffel, B Numerical, Experimental and Theoretical analysis of the Individual Efficiencies of a Centrifugal Pump, Proceedings of ASME FEDSM- 01, New Orleans, Louisiana, May 29- June 1: Versteeg, H.K. and Malalasekera, W An Introduction to Computational Fluid Dynamics, Prentice Hall, Loughborough. Wilson, K.C., Addie, G.R. and Clift, R Slurry transport using centrifugal pumps, Elsevier Science, New York. Zhou, W., Zhao, Z., Lee, T.S. and Winoto, S.H Investigation of flow through centrifugal pump impeller using computational fluid dynamics, International Journal of Rotating Machinery 9: