Effect of Suction Pressure on Hydraulic Parameters of Booster Pumping Station

Transcription

1 Effect of Suction Pressure on Hydraulic Parameters of Booster Pumping Station Ibrahim R.Teaima, Mechanical and Electrical Research Institute, National Water Research Center, Egypt Abstract This paper seeks to study experimentally and numerically the effect of suction pressure in the suction side of Ahmed Badway Booster Pumping Station on the hydraulic performance of a single stage split case centrifugal pump. Various hydraulic parameters including flow, velocity, head, power and efficiency were measured in the field. These values were measured at various suction pressure ratios (H/H des ) ranging from 0.1 to 0.6. The numerical simulation of impeller and volute interaction was performed using FLUENT under ANSYS 13.0 commercial code with solving 3D URANS. The flow was assumed to be three dimensional, viscous and incompressible. A k-ε RNG turbulence model, with non-equilibrium wall functions was used. In this paper, the pump is driven by 340 hp electric motor connected directly with the pump. The operating head and flow rate of this pump were 97 m and 500 m 3 /hr, respectively, and the motor speed was 1490 rpm. The specific speed value was The number of impeller blades was seven. Numerical simulations are performed to investigate the unsteady flow in booster pump with different inlet suction pressure ratios based on the field measurements. The research indicated that the pressure increased gradually along stream-wise direction in the impeller passages. Operation under suction pressure ratios more than 0.3 produce unsteady flow in the impeller passage and the volute casing leading pressure fluctuation and unforeseen hydrodynamic force reducing component life time. Keywords: Suction pressure; Hydraulic parameters; Booster pump; Numerical simulation. Nomenclature b = Impeller width, (mm) Cp = Pressure fluctuation coefficient CFD = Computational fluid dynamics D1 = Impeller suction diameter, (mm) D2 = Impeller diameter, (mm) dn = Hub diameter, (mm) Δt = Time step, (s) g = Acceleration due to gravity, (m/s2) H = Delivery head, (m) k = Turbulence kinetic energy, (m2/s2) n = Rotating speed, (rpm) L = Volute length, (mm) ns = Specific speed, (-) Q = Flow rate, (m3/s) R, r = Radius, (mm) RNG = Re-normalization group Tu = Turbulence intensity U = Circumferential velocity, (m/s) W = Relative velocity, (m/s) z = Blade number Greek symbols β = Baled angle, (degree) ρ = Density of water, (kg/ m3) μ = Viscosity (Pa s) μt = Eddy viscosity (Pa s) Ω = Rotational speed, (rpm) θ = Volute cone angle, (degree) θ * = Normalized circumferential coordinate Subscripts 1 = Inlet impeller 2 = Outlet impeller 3 = Outlet volute outlet = Pump outlet Open Access Journals Blue Ocean Research Journals 9

2 PS SS des = Pressure side = Suction side = Pump operating point Introduction In recent years, concerns about organization and management of the world s water resources have been continuously increasing. With a better feasibility Study and understanding of the water application problems, an ever increasing demand for water production and the advances in technologies, there has been a change in the pumping qualifications needs. The trend is nowadays towards superior high capacity and medium head units pumping solutions. The design demands a detailed understanding of the internal flow during design and off-design operating conditions. Complex flow pattern inside a centrifugal pump is strong three-dimensional with recirculation flows at inlet and exit where flow separation and cavitation. The suction pressure is an improved parameter for the pump performance. If the suction pressure is insufficient, it probably leads to cavitation. To improve design of centrifugal machines, a better understanding of the flow behavior of such machines is required. This paper deals with an experimental and theoretical study of the hydraulic parameters of low specific speed booster centrifugal pump. Ahmed Badawi Booster Pumping Station is located at Ahmed Orabi association of agricultural cooperative [1]. The main pumping station is located at 14 km from the Ahmed Orabi intake pumping station. The station consists of 5 units split case centrifugal pump. The pump operating point is 500 m 3 /hr discharge and 97 m total head. The specific speed value is (ns= 17.5). The suction and discharge pipelines are 860 mm diameters. After three months of operation, the volute of pump unit number three was found to be blocked and cracked. The researchers have treated the problem of the interaction of the impeller and its surroundings numerically and experimentally A detailed analysis of the results at design flow rate, Q design and off-design conditions, Q = 0.43 Q design and Q = 1.45 Q design [4], presented the application of particle image displacement velocimetry to the measurement of fluid velocities in a centrifugal pump diffuser. Measurements were taken at different operating points and allowed defining the variation of radial and tangential velocity components along a pitch. Results were also compared with laser Doppler measurements taken in the same facility. A computational fluid dynamic (CFD) have successfully contributed and is more efficient to the prediction of the flow through pumps and the enhancement of their design. These advances have made it possible for pump designers to carry out analysis and simulation of various flow and impeller behavior occurring inside pumps. To improve design of these pumps, a better understanding of the flow of such machines is required. The numerical simulations seem to predict reasonably the total performance and the global characteristics of the laboratory pump. An unsteady simulation of a centrifugal pump was done by [2] based on FLUENT to predict the dynamic characteristics of flow between impeller and volute as well as the pump performance. The research was validated by the experimental data. [3], made an investigation to simulate the complex internal flow in a centrifugal pump impeller with six twisted blades by using a three-dimensional Navier-Stokes code with a standard k- ε two-equation turbulence model. In the present study, This work aims at analyzing the flow behavior through a pump cascade with the help of Computational Fluid Dynamics using the FLUENT software s, computational flow analysis is done using sliding mesh method for six values of suction pressure ratios, (H/Hdes) ranging from 0.1 to 0.6 with step of 0.1 for spiral volute casing centrifugal pump. The performance characteristics are obtained from field measurements. The suction pressure effects on the velocity and pressure distributions will be analyzed for impeller and volute by using Fluent under Ansys Ver.13. Governing Equations The incompressible flow through the rotating impeller is solved in a moving frame of reference with constant rotational speed equal to the rotational speed of the impeller. The flow through the stationary parts of the pump is solved in an inertial reference frame. Three-dimensional, unsteady Reynolds-averaged Navier-Stokes equations are solved. For 3-dimensional, unsteady, incompressible flow [11], the continuity equation (1) and conservation of momentum (2) can be written as follows: ρ u r = 0 (1) ρu + 2ρΩ u + ρω Ω r = p + µ r r where is the density of the fluid, p is the static pressure, u r is the vector fluid velocity in the rotating system, Ω is the rotational speed and μ eff is the dynamic effective viscosity which is a linear combination of laminar and turbulent viscosity derived from k-ε model of turbulence. The last two terms in the left hand side of equation (2) are the effects of the Coriolis and centrifugal forces due to the rotating frame of reference. For the stationary parts of the centrifugal pump, the governing equations are formulated eff (2) 2 u r Open Access Journals Blue Ocean Research Journals 10

3 in the stationary reference frame. The continuity equation remains the same, but the momentum equation reduces to ρu p + µ eff 2 u Where, u is the vector fluid velocity in the stationary frame of reference. The re-normalization group (RNG) k-ε two-equation model is used in the present calculation. The k and ε equations can be written as: ρuk µ t = µ + + σ K G k µ t ρuε = µ ε + + C σ ε µ t = ρ C m 2 k ε k 1ε (3) ρε (4) k G ε k C where U is the local velocity vector, k is the turbulent kinetic energy, ε is the dissipation rate, μ is the laminar viscosity, μ t is the turbulent viscosity, G k represents the generation of turbulent kinetic energy due to the mean velocity gradients, σ k and σ ε are the turbulent Prandtl numbers and C 1ε =1.39, C 2ε =1.39 and C m =0.09 are the constants of the model. (5) (6) 2ε Table 1. Geometrical parameters and operating conditions The discretization in space is of second order accuracy. The sliding mesh method is used to model the rotor zone motion in order to simulate the impeller-volute casing interaction. The governing equations are solved using the segregated solver and a centered SIMPLE algorithm is used for the pressure-velocity coupling. The time step Δt is set to s. A hybrid mesh is created using FLUENT s preprocessor GAMBIT. Velocity inlet and pressure outlet boundary conditions are applied at the inlet and at the outlet, respectively. Model Geometry And Grid Generation The test pump stage consists of an impeller and volute. The impeller is shrouded with seven two-dimensional and strongly backswept blades, with an exit angle of 11.7 deg relative to the tangential direction. The pump (impeller spiral volute) parameters are presented in Table (1) in detail. εthe structured grid for computational domains in a 2 ρ meridional k plane and at the midspan of the impeller and volute is shown in Fig. 1. The grid is created in two parts by using the commercial software Cfturbo [12] with multiblock templates for impeller and volute, respectively. A hybrid mesh is created using FLUENT s preprocessor GAMBIT 2.2 [13]. The grid maximum aspect ratio is Fig. 2 shows the impeller and volute casing notations. The pump impeller consists of seven cross-sectional planes that cut in according to the various locations in volute casing for later discussion. The impeller passages are labeled from 1 to 7 in anticlockwise direction with passage 1 closest to the volute tongue. Description Design operation point Impeller dimensions Blade properties Casing Radial pump with volute casing Flow rate, Q des 500 m 3 /hr Rotational speed, n des 1490 rpm Delivery head, H des 97 m Specific speed, n s 17.5 (rpm, m 3 /s, m) Number of blades, z 7 Hub diameter, d N 74.7 mm Suction diameter, D mm Impeller diameter, D mm Outlet width, b 2 36 mm Outlet width ratio (b 2 /D 2 ) Number of meridional sections 7 Thickness leading edge, SLE 6.1 mm Thickness of trailing edge, STE 7 mm Angle of leading edge, β 1 13 Angle of trailing edge, β Type Spiral volute type Diameter D mm Width, b 3 66 mm Diffuser direction tangential Open Access Journals Blue Ocean Research Journals 11

4 End cross section circle Exit diameter, D outlet mm Length, L mm Cone angle, θ 5 (a) (b) (c) (d) Fig. 1. (a) Spiral volute casing, (b) Pump impeller and spiral volute, (c) Computational domain and (d) Meshes of the volute and impeller flow domain Doutlet W 2 V 2 β 2 L Plane I (0 ) STE b2 u 2 θ W 1 SLE b3 b1 V 1 Plane II (90 ) ω D1 Plane IV (270 ) D 1 u1 β1 Impeller ω D 2 D2 D3 Volute Plane III (180 ) (a) Field Measurements (b) Fig. 2. (a) Impeller and volute notations and (b) Pump impeller notations. Open Access Journals Blue Ocean Research Journals 12

5 The hydraulic performance evaluation of Ahmed Badwi Booster Pumping Station is done in all pumping units, especially on unit number three. After long period of operation the volute casing is exposed to erosion and crack and the unit out of service as shown in Fig. (3-a), but after changing the new volute all measurements and analysis are done as shown in Fig. (3-b). A transit-time ultrasonic flow-meter type, Siemens-Fup is used to measure the volume flow rate. Calibrated pressure transducers are used to measure pressures at the suction and delivery sides of the pumping units. Three phase energy analyzer, (Fluke 434) is used to measure voltage, ampere, active power, energy, apparent power, frequency and power factor. The pressure head developed by the pump is recorded with time at different operating conditions. erosion (a) (b) Fig. 3. (a) Volute casing of unit number three, (b) New volute after changing. RESULTS AND DISCUSSIONS Field measurements Hydraulic performance tests are done during normal operating conditions around the design point, where the control valve in the delivery pipe is used to control flow rate. Discharge, suction pressure, delivery pressure, power quality values are monitored and total head is computed for unit number three. Fig. 4 shows the manufacture (H-Q) curve and the results measurement of (H-Q) curve and (H- Efficiency) curve at different operation conditions. From the figure, the actual curve at measured speed of 1490 rpm is slightly less than the manufacture pump head in the flow rate range. Furthermore the pump doesn t match the operation point according to ISO It is believed that, the big deviation between the measured performance curves and the standard manufacturing curves is due to hydraulic problems in the suction side based on the variation of suction pressure. However, the flow details, which could not be experimentally obtained inside the pump, are numerically obtained. Open Access Journals Blue Ocean Research Journals 13

6 Numerical results It is difficult to understand the flow phenomena due to impeller volute interaction from only experimental studies because of the complicated flow structures in centrifugal pumps. Moreover the flow in centrifugal pumps is exceedingly complex, involving curvature, system rotation, separation, turbulence, and secondary flows. Different flow rates were specified at inlet boundary to predict the characteristics of the pump. The CFD technique carried out of sliding mesh is mainly a factor typical to 50 more than the use of a mixing plane technique and a factor 30 more than the use of the Multiple Reference Frame method [6]. Although of the further improve the aim of pump operation performance for design and off-design operating conditions, it is extremely hardly effort which may be difficult. The effect on the impeller and diffuser due to other conditions such as the boundary layer separation, vortex dynamics and interactions are difficult to manage because of the presence of the rotating and stationary components. [7, 8] found that jet-wake structure takes place near the exit of the impeller and it is independent of flow rate and locations. Byskov et al [9] based on shroud of six-bladed impeller using the technique of large eddy simulation (LES) at design and off-design conditions. During the design load, it was observed that the flow field inside the impeller was slightly smooth and with no takes into account of separation. At part load design like in case of quarter load, a steady stall phenomenon was observed in the inlet and a relative eddy was developed in the remaining of the passage. [10]. Moreover, the geometry is often asymmetric due to the volute shape. As a result, the relative motion between impeller and volute generates an instability which affects not only the overall pump performance, but also is responsible for pressure fluctuations, hydraulic noises and unforeseen hydrodynamic forces. These fluctuations generate noise and vibration that cause unacceptable levels of stress and reduce component life due to fatigue. Unfavorable characteristics of pump performance are also generated even at or near the design point Static pressure distribution inside impeller and volute. The pressure distribution increases gradually along streamwise direction within rotor impeller passage between blades in which pressure surface has a higher pressure than suction surface for each plane. Also it is observed that the static pressure at volute outlet is higher for high suction pressure inlet and lower for low suction pressure inlet and that is clearly shown in Fig. 5 in which the contours of variation of static pressure distribution in the middle surface at H/H des = 0.1, 0.2, 0.3, 0.4, 0.5 and 0.6 respectively. The uniform pressure increases with radial direction keeping axisymmetric impeller channels for the inlet suction pressure rations less than 0.3. For inlet suction pressure ratios more than 0.3 a region with higher pressures is found just preceding the volute tongue, in the rotating direction, while a region of low pressure distribution is observed just following the volute tongue. At H/Hdes = 0.1, the blade passage of the impeller arriving at the volute tongue undergoes a strong pressure drop, so that a nonuniform pressure zone is created at the outlet section. But for H/Hdes = 0.6, the blade passage leaving the volute tongue experiences the largest pressure gradient. Its influence on the pressure field is stronger in the high inlet pressure than in low inlet pressure to the pump. The unsteady pressure field near the tongue region of the volute of centrifugal pump appears when the suction pressure ratio reaches 0.4. The conversion of dynamic pressure produced by the impeller rotation into static pressure in the volute casing can be seen, thus the maximum of pressure is obtained at the outlet duct, except at high inlet pressure and design operating point. Whatever the flow rate and the pump are, a nonhomogeneous pressure distribution is observed at the zone around the gap between volute tongue and impeller periphery, characterized by a high gradient of pressure. The volute tongue whose role is to drive the flow towards the impeller outlet presents a singularity for the flow. The logarithmic shape of the volute casing creates a geometrical asymmetry. This phenomenon influences the pressure field distribution and fluctuating stresses applied on the impeller blades. The maximum static pressure area appears at volute tongue and outlet regions and the minimum one at the back of blade at impeller inlet region. Open Access Journals Blue Ocean Research Journals 14

7 (a) H/H des = 0.1 (b) H/H des = 0.2 (c) H/H des = 0.3 (d) H/H des = 0.4 (e) H/H des = 0.5 (f) H/H des = 0.6 Fig. 5. Computed static pressure fields at different inlet pressure ratios. Instantaneous static pressure coefficient The dimensionless coefficient of static pressure (C P ) focused on the magnitude of pressure fluctuation in which the standard deviation calculate the instantaneous static pressure normalized by the dynamic pressure based on the impeller tip speed U2. The distribution of pressure fluctuation at midspan of the impeller outlet is plotted in Fig. 6. The high pressure fluctuation was previously observed from the tongue, [14] and decreased gradually with the impeller rotation direction. Consequently the pressure fluctuation increases with increase of the suction pressure ratio. It can be also observed that the pressure has a much stronger fluctuation for suction pressure ratio more than 0.3, but at low pressure ratios less than 0.3 the pressure fluctuation exhibits a good periodicity every two impeller channels. Open Access Journals Blue Ocean Research Journals 15

8 Fig. 6. Static pressure coefficient along impeller circumference. Absolute velocity distribution inside impeller and volute The contour plot of absolute velocity distribution is shown in Fig. 7. As shown in the figure, the velocity increases from impeller inlet to outlet and reaches a peak value of 28 m/s at impeller outlet. After entering the volute, the velocity begins to fall down, reaching the lowest at the outlet region inside the volute. The flow enters the impeller eye; it is diverted into the blade-to-blade passage. Separation of flow can be observed at all passages leading edge. The leading edge velocity contours field with separation and inflow incident angle are non-tangent to the blade leading edge. At six different suction pressure ratios, the leading edge flow separation patterns are slightly different. This could be contributed to the boundary layer thickness that grows from the leading edge at suction side with different suction pressures, which could lead to energy loss in the pump and further influence the flow field in impeller passage in stream-wise direction. The impeller passage flow at low suction pressure is very smooth and well-guided except at the leading edge as discussed above. The flow follows the blade curvature profile from impeller passage entrance till the exit without any separation on blade pressure side. This is matching the potential flow theory for flow passing around turbo-machinery blade. (a) H/H des = 0.1 (b) H/H des = Open Access Journals Blue Ocean Research Journals 16

9 (c) H/H des = 0.3 (d) H/H des = 0.4 (e) H/H des = 0.5 (f) H/H des = 0.6 Fig. 7. Contour of absolute velocity magnitude at different inlet pressure ratios. Relative velocity vectors distribution inside impeller and volute The instantaneous relative velocity vectors in the pump are plotted in Fig. 8 for different inlet pressure ratios. The volute tongue zone presents a strong recirculation of the fluid particles at the gap between the volute tongue and the impeller periphery. The velocity fields show more significant variations when the pump works at high inlet pressure ratios. Consequently, when the suction pressure ratio increases to 0.6, a dead volume zone with low velocity magnitude is observed in quarter of periphery from the volute tongue. The volute tongue is a singularity for the flow that creates a strong recirculation zone at the volute diverging outlet where the fluid particles are slowed down as shown in Fig Open Access Journals Blue Ocean Research Journals 17

10 (a) H/H des = 0.1 (b) H/H des = 0.2 (c) H/H des = 0.3 (d) H/H des = 0.4 (e) H/H des = 0.5 (f) H/H des = 0.6 Fig. 8. Relative velocity vectors distribution at different inlet pressure ratios. Mean relative velocity field Fig. 9 presents the mean (time-averaged) relative velocity contours. A local region near the suction (concave) side with relatively high relative velocity is detected. The relative velocity on the suction side is bigger than that on the pressure (convex) side in the front impeller part; this is in accordance with the fact that high-momentum fluid is displaced toward the suction side near the inlet section according to the potential theory. It also indicates that in the inner part of the passage, the meridional curvature associated with the axial-to-radial entry bend dominates over rotational effects. However, this phenomenon is reversed in the impeller rear part due to the fact that Coriolis force that drives the fluid from the suction side to the pressure side in large radii near outlet section, [3]. This phenomenon is observed by the existence of the jet-wake flow Open Access Journals Blue Ocean Research Journals 18

11 structure near the impeller outlet, which is characterized by low relative velocity on the impeller suction side near the trailing edge and relatively high velocity on the corresponding pressure side. It is also observed from Fig. 9 that the relative velocity contours decrease at volute outlet when suction pressure ratio increases than 0.3. Furthermore in the range of suction pressure ratios varied from 0.1 to 0.3 the relative velocity increases at the volute outlet; thus because when increasing the suction pressure ratios more than 0.3 the outlet relative velocity decreases at the volute outlet. (a) H/H des = 0.1 (b) H/H des = 0.2 (c) H/H des = 0.3 (d) H/H des = 0.4 (e) H/H des = 0.5 (f) H/H des = 0.6 Fig. 9. Phase-averaged relative velocity field W/u 2 at different inlet pressure ratios. Open Access Journals Blue Ocean Research Journals 19

12 Turbulence intensity The turbulence intensity distribution (Tu) is shown in Fig. 10. It is observed that the high-turbulence region on the volute tongue increases with increase of the inlet suction pressure ratio more than 0.3, reaching the volute outlet. Furthermore the high turbulence values are also observed on the impeller suction side, starting near the impeller leading edge corresponding to the impeller pressure side. In the second half of the impeller passage, high turbulence level is indicated in the impeller pressure side corresponding to the impeller suction side especially at the suction pressure ratio of 0.1. (a) H/H des = 0.1 (b) H/H des = 0.2 (c) H/H des = 0.3 (d) H/H des = 0.4 (e) H/H des = 0.5 (f) H/H des = 0.6 Fig. 10. Shows the turbulence intensity (T u ) distributions at different inlet pressure ratios. Open Access Journals Blue Ocean Research Journals 20

13 Verification Of The Model Simulation In this section a comparison between the present CFD work and [15] results. Fig. 11 shows quantitative comparison of the average relative velocities obtained by the present CFD work and Feng et al experimental and CFD work at impeller outlet. The normalized circumferential coordinate, θ*=0, at impeller pressure side (PS) and θ*=1, at the suction side (SS). The present CFD calculations show an excellent agreement with Feng et al [15] results near the blade suction and pressure sides. It is also observed that, near the impeller outlet, the jet-wake flow structure is observed, which is characterized by low relative velocity on the impeller suction side near the trailing edge and relatively high velocity on the corresponding pressure side. Fig. 11. Comparison of present numerical results and Feng et al. [15] average relative velocity at impeller exit. Conclusion And Recommendation Experimental study and numerical simulations are performed to investigate the unsteady flow in booster pump with different inlet suction pressure ratios equal to 0.1, 0.2, 0.3, 0.4, 0.5 and 0.6, respectively. Calculations are performed at pump operating points, for the impeller and volute. The following headings can be concluded: a) For booster pump with low inlet pressure ratios less than 0.3 at design point, the internal flow or velocity vector is very smooth along the curvature of the blades. But at inlet suction pressure ratios more than 0.3; the flow pattern has changed significantly from the well-behaved flow pattern at design load condition. b) The high pressure fluctuation is observed from the tongue and increases gradually with increase of suction pressure ratio. c) Increasing the suction pressure ratios more than 0.3 leads to decrease the outlet relative velocity at the volute outlet. d) The pressure fluctuation on the impeller pressure side is much bigger than on the impeller suction side, and it is more evident at the rear part of the impeller blade. e) The uniform pressure increases with radial direction for the inlet suction pressure ratio less than 0.3. f) For inlet suction pressure ratio more than 0.3, a region with higher pressure is found just preceding the volute tongue. g) Increasing inlet suction pressure ratio more than 0.3, high-turbulence region in the volute tongue is generated. h) From the previous conclusions it can be recommended that, the inlet suction pressure ratios don t exceed than 0.3 to prevent the pressure fluctuation and unforeseen hydrodynamic force which reduce component life time due to fatigue accumulation. Open Access Journals Blue Ocean Research Journals 21

14 References [1] M. A. Younes, "Investigation of Hydraulic Problems in Pumping Station; Case Study", Twelfth International Water Technology Conference, IWTC12, pp , Alexandria, Egypt, [2] J. L. Parrondo, J. González, J. Fernández, and L. Fernández, "An Experimental Study on the Unsteady Pressure Distribution around the Impeller Outlet of a Centrifugal Pump," ASME-FEDSM , [3] K. W. Cheah, T. S. Lee, S. H. Winoto, and Z. M. Zhao, "Numerical Flow Simulation in a Centrifugal Pump at Design and Off-Design Conditions," International Journal of Rotating Machinery, vol. 2007, 8 pages, [4] N. Paone, M. L. Riethmuller and R. A. Van den Braembussche, "Experimental Investigation of the Flow in the Vaneless diffuser of a centrifugal pump by particle image displacement velocimetry", Experiments in Fluids 7, , Springer-Verlag [5] ANSYS, "Fluent User Document, version 13.0", ANSYS Europe, [6] O. E. Abdellatif, M. Abd Elganny, and I. Shahin," Performance and Unsteady Flow Field Prediction of a Centrifugal Pump with CFD Tools", Proceedings of ICFD 10: Tenth International Congress of Fluid Dynamics, Ain Soukhna, Red Sea, Egypt, [7] M. J. Zhang, M. J. Pomfret, and C. M. Wong, "Three Dimensional Viscous Flow Simulation in a Backswept Centrifugal Impeller at the Design Point", Computers and Fluids, vol. 25, no. 5, pp , [8] M. J. Zhang, M. J. Pomfret, and C. M. Wong, "Performance Prediction of a Backswept Centrifugal Impeller at Off-Design Point Conditions", International Journal for Numerical Methods in Fluids, vol. 23, no. 9, pp , [9] R. K. Byskov, C. B. Jacobsen, and N. Pedersen, "Flow in a Centrifugal Pump Impeller at Design and Off-Design Conditions part II: Large Eddy Simulations", ASME Journal of Fluids Engineering, vol. 125, no. 1, pp , [10] F. Gu, A. Engeda, M. Cave, and J. L. Di Liberti, "A Numerical Investigation on the Volute/Diffuser Interaction Due to the Axial Distortion at the Impeller Exit", ASME Journal of Fluids Engineering, vol. 123, no. 3, pp , [11] Johann Friedrich Gülich, "Centrifugal Pumps", First Edition, Springer-Verlag Berlin Heidelberg, [12] CFturbo, "Turbomachinery Design, Version 9.0.7", Cfturbo Software &Engineering GmbH, Dresden- Munich, Germany, December 14, [13] The FLUENT User's Guide, Gambit November 28, [14] P. Adami, S. Della Gatta, F. Martelli, "Multi Stage Centrifugal- Pumps: Assessment of Mixing Plane Method for CFD Analysis", 60 Congresso Nazionale ATI, September, 2005, Roma. [15] J. Feng, F. K. Benra, and H. J. Dohmen, "Numerical Investigation on Pressure Fluctuations for Different Configurations of Vaned Diffuser Pumps", International Journal of Rotating Machinery, vol. 2007, 10 pages, Open Access Journals Blue Ocean Research Journals 22