Numerical and Experimental Investigation of High-efficiency Axial-flow Pump

SHI Weidong, ZHANG Desheng, GUAN Xingfan, and LENG Hongfei

Technology and Research Center of Fluid Machinery Engineering, Jiangsu University, Zhenjiang 212013, China

1 Introduction

The flow in axial-flow pump is greatly influenced by the effects of turbulence and viscosity. Much of the physical phenomena and laws involved in this complex flow field can’t be fully determined. FARRELL[1]discussed the end-wall vortex cavitation in high Reynolds axial-flow pump. ALPAN, et al[2], analyzed the suction reverse flow in an axial flow pump by experiments. ZIERKE, et al[3–4],introduced the experimental technology and discussed the flow characteristics of high Reynolds pump and tip clearance flow. DUPONT, et al[5], investigated the unsteady effects associated with rotor stator interactions in a vaned-diffuser radial-flow pump. SITARAM, et al[6], used the probe to measure the relative flow in a turbo machinery rotor blade passage. LI, et al[7]and WANG, et al[8],simulated the axial-flow pump with inducer, and discussed unsteady turbulence numerical simulation in axial-flow pump. HUANG, et al[9–10], compared numerical streamlines to the particle image velocimetry(PIV) measurement results.Axial-flow pump research group of Jiangsu University,China, has investigated axial-flow pump models with high-efficiency in the past decade[11–14]. All the research work above focuses on the normal performance and flow field of axial flow pump. The efficiency of axial-flow pump is becoming one of the most important performance indicators in the large pumping stations, and the flow structure in axial-flow pump is associated to its efficiency.

In this paper, in order to analyze the flow structure in the adjustable axial-flow pump model, a high efficiency model with specific speed 800 is simulated by FLUENT code at various conditions, and a five-hole probe was used to measure the flow field of inlet, blade exit and guide-vane outlet. The results of simulation and experiment enable us to improve our understanding of the flow characteristics in high efficiency axial-flow pump.

2 Numerical Simulation of Inner Flow Field

2.1 Axial-flow pump model

Fig. 1 shows 3D geometric of axial-flow pump model with specific speed ns=800 and impeller diameter 300,which is used to simulate and measure its entire flow field.

Fig. 1. 3D geometric of axial-flow pump model

2.2 Governing equations

The three-dimensional Reynolds-average Navier-Stokes equations are solved by FLUENT code in strong conservation form. The transport equations are discredited by using a conservative finite volume method. These simulations were achieved with a second order accurate skew upwind differencing and scheme and physical advection correction. The turbulence effects were modeled by the RNG k–ε turbulence model. The RNG model provides a way to account for the effects of swirl or rotation by modifying the turbulent viscosity appropriately.SIMPLEC arithmetic is applied to solve the pressure and velocity coupling, and standard wall functions are used to model the viscous sublayer.

2.3 Computational domain and boundary conditions

The whole hydraulic passage of the axial-flow pump is taken as the computational domain, and assumed to be steady. The inlet of the solution domain is located approximately at the upstream of the axial-flow blade leading edge with a four impeller diameter distance. A constant axial velocity based on the mass-flow rate is specified at the inlet for each computation. All physical surfaces of the pump are set to be no-slip wall. At the outlet,which is roughly five impeller diameter downstream of the vane trailing edge, the gradients of the velocity components are assumed to be zero. The flow zone is divided into 3 sub-domains. The first is the region defined from the pump inlet to the inlet of rotor. The equations for this region are solved in a stationary framework. The second is the impeller region. This region is attached to the rotating frame, and is solved in a rotating framework by the multiple reference frame(MRF) method. The third is the other region, and is also a stationary zone. The interfaces between the rotating region and stationary regions are shown in Fig. 2.

Fig. 2. Computational grids and interfaces

A hybrid meshing scheme is used, with hexahedral cells in the near blade region and tetrahedral and pyramid cells in the rest of the domain. Relatively fine grids are used near the hub, shroud, and blade surfaces, as well as near the leading, trailing edges and tip clearance.

2.4 Simulation results and discussion

2.4.1 Performance prediction

The performance of the pump at blade angle 0°,predicted from the computational fluid dynamics(CFD)model, was compared with the experimental results. As shown in Fig. 3 and Fig. 4. Here Q is flow rate; Qoptis flow rate at design condition.

Fig. 3. Experimental and predicted results of head

Fig. 4. Experimental and predicted results of efficiency

The prediction data (CFD) shows agreement with the experimental head and efficiency (EXP). The difference between the simulation and experiment may be due to our neglect of the unsteady feature of flow between impeller and guide-vane. Furthermore, the current rid size,turbulence model and wall functions used in this study is not the best selection and further careful investigations are needed in the future. The simulation method in the paper could be used in hydraulic optimization and flow filed analysis of the axial-flow pump.

2.4.2 Static pressure distribution

In Fig. 5, Lxis string length at x direction, Ltis total airfoil string length, R is radius, and Lx/Ltis string span.The results show that the static pressure on pressure side of blades increases slightly at circumferential direction, and keeps almost constant at the same radial, while it increases gradually from inlet to exit along the flow direction on suction side. At leading edges of the blades, local impact results in an increase of pressure.

Fig. 5. Static pressure distributions on different profiles

The static pressure distributions at inlet, impeller outlet and vane outlet at design conditions were simulated. We define location factor r=(Rm–rh)/(rt–rh), where Rmis the radius of measurement point, rtis the radius of tip, rhis the radius of hub. The static pressure at the inlet and vane outlets at radial locations is constant, as shown in Fig. 6.Furthermore, it decreases near hub leakage, and shows agreement with the measurement data.

Fig. 6. Static pressure distributions in pump

2.4.3 Velocity distribution

In the design cycle, the blades of the impeller were designed by using the uniform pattern of meridional velocity. However, this distribution has not been found in the numerical results. Fig. 7 shows the meridional velocity distribution at impeller outlet plane. It can be seen that the meridional velocity changes periodically along the circumferential direction and the change frequency is identical with the vane blade number, which is also influenced by the impeller numbers. In general, axial velocity increases from hub to tip, and it is affected by the tip clearance and hub leakage which lead to decrease of meridional velocity and efficiency.

Fig. 7. Meridional velocity distribution (m/s)

3 Experimental Investigation

3.1 Experimental apparatus

The experiment was performed in fluid machinery laboratory of Jiangsu University, China. Fig. 8 shows the main experimental apparatus.

Fig. 8. Experimental apparatus

Three components of velocity at circumferential direction of the inlet, impeller outlet and guide-vane outlet at different conditions were measured by a calibrated five-hole pressure probe. Semi-measurement method was used to calculate the velocity vector in the experiments[15],and the five-hole probe structure is shown in Fig. 9.

Fig. 9. Five-hole probe structure

The static pressure and total pressure at the same locations were also measured at the same time. The flow field at blade angle Φ= 0° was measured. Fig. 10 shows the measurement locations.

Fig. 10. Probe measurement locations

DUPONT, et al[5], listed sources of errors in conventional probe measurements of flow in turbine machinery. They also estimated the magnitude of these errors. They agreed that wall proximity effects were negligible if the distance between the measurement position and the wall was two times longer than the probe diameter.The first and the last measurement locations shown here fulfilled this condition. The interval between measurement points except the first measurement and last one is 10 mm.The Reynolds number is the same for calibration in air and experiments in water, thus the effects of Reynolds number are almost non-existent. The presence of the probe may perturb the flow but the size of the probe is 10 times smaller than the blade passage, so the effects of the probe blockage are negligible.

3.2 Experimental results

3.2.1 Performance characteristics

The pump was manufactured according to the 3D model,and tested over the operational range at five blade angles(–4°, –2°, 0°, +2°, +4°). The overall performance curves of the model pump with rotor diameter D2=300 mm and rotation speed n=1 450 r/min are shown in Fig. 11, and optimal efficiency points at five blade angles can be seen in Table. The efficiency at blade angle Φ=+4° reaches 86.05%,which is the best efficiency in public in China, so we define it as a high-efficiency axial-flow pump.

Fig. 11. Overall performance curves of the model pump

Table. Optimum parameters at five blade angles

3.2.2 Inlet flow

Due to the size of the probe and the vicinity of the hub,the first measurement is located at r=0.07 and the last one at r=0.92 as shown in Fig. 12. The data were collected using a field point measurement method. In this procedure,the measurement volume remains stationary, and the flow in the pump is assumed to be steady. A five-hole probe resolves the three components of velocity at blade inlet.Fig. 12(a) shows that the absolute flow angle α (α=arctan(vu1/vm1)) is almost equal to 0 at various flow rates, as expected, which means the inlet flow is almost axial and the prerotation at blade inlet is very small. The meridional velocity vm1(Fig. 12(b)) at impeller inlet all increases slightly from tip to hub at 0.8Qopt, 1.0Qoptand 1.2Qoptconditions which may be relevant to the sharp of front hub.The distribution of the static pressure at inlet is almost consistent at design condition, which is the expression of steady flow field. At off-design conditions, the unsteady flow in rotor induces the static pressure change at different radial locations, but the change range is small, which can be observed in Fig. 12(c). Fig. 12(d) shows that the total pressure increases linearly from tip to hub at various flow rates associates in correspondence with the axial velocity changes.

3.2.3 Impeller outlet flow

The hub leakage of adjustable blades, which influences the flow in the vicinity of hub, is the most important factor that influences the efficiency of axial-flow pump. We sealed it up with wax and tested the performance. With its flow field at 0.8Qopt, 1.0Qoptand 1.2Qoptconditions it was measured to compare with the blades with hub leakage. The performance test results show that hub leakage leads to the decrease of efficiency, and the maximum efficiency can reach 86.85% without hub leakage, so our special attention should be paid to examine the flow field near hub with leakage, which is equally important with tip clearance. The hydrodynamic flow field at blade exit is shown in Fig. 13.In Fig. 13, 1 (hollow symbols) means the results of flow field without hub leakage, and 2 (solid symbols) means the results of flow field with hub leakage. At the design point,the distribution of meridional velocity component vm2is almost uniform, which means that the flow at blade outlet is steady without radial flow. When the flow rate reduces,significant changes occur in the flow field. The new feature is that a reverse flow developed in this region, where vm2increases strongly from hub to mid-radius, and then decreases, but increases again near the tip. The significant difference between the blades with and without hub leakage,which can be seen in Fig. 13(a), is the decrease of vm2near the hub owing to the hub leakage. This phenomenon can be explained as follows: the reverse flow from pressure side to suction side reduces vm2, which is associated with the drop of static pressure p, totol pressure p0, and vu2near the hub leakage.

Fig. 12. Hydrodynamic flow field at inlet at various locations at different flow rates

Fig. 13. Hydrodynamic flow field at blade outlet at various locations at different flow rates

Fig. 13(b) shows the distributions of circumferential velocity component vu2. It is noticeable that vu2decreased gradually from hub to tip at 1.0Qoptand 1.2Qoptconditions in accordance with the consistent Г distributions. At small flow rate conditions, the distribution regularity of vu2is similar to vm2at small flow rate conditions, which is also influenced by the reverse flow.

The decrease of the total pressure p0and the static pressure p in the tip clearance and hub leakage in Figs.13(c), 13(d) indicates large energy losses in these regions,which should be examined carefully during the axial-flow design process. vm2also decreases slightly near tip.Nevertheless, the decrease is small because of the small tip clearance of 0.15 mm.

The single profile exit circulation is shown in Fig. 13(e).The distribution of circulation Г at downstream of the impeller blades is the most important factor for efficiency.Compared with two kinds of blades, at design condition and large flow rate conditions, the circulation Г is almost constant but a small decrease near hub, caused by the hub leakage. This phenomenon may be further relevant with the circulation distribution at hub and tip regions in design to make the circulation consistent along the radial locations in practice. The test results (see Fig. 11) show that the maximum efficiency reached 85.29% at Φ=0o, so it could prove that constant circulation flow pattern, which is corrected with the non-linear circulation distributions,could obtain high efficiency. At small flow rate conditions,Г decreases sharply from tip to hub. This is recirculation near hub. Although many experts have studied the flow at small flow rate, the distribution of Г is still hard to control.At small flow rate conditions, the efficiency of pump decreases rapidly.

3.2.4 Guide-vane outlet flow

Guide-vane is an important component of the axial-flow pump. The high-speed flow velocity from the blades must be reduced to turn dynamic energy into pressure energy before it enters to the outlet pipe, which means the vane must be carefully designed to avoid flow separation and instability. The flow quality at outlet of vane has a direct impact on the efficiency. Fig. 14(a) displays that circulation can not be eliminated absolutely at downstream of guide vane. vuis almost constant at 1.0Qoptand 1.2Qopt,conditions, while it increases linearly from hub to tip at small flow rate conditions. vmdistribution is also almost constant at different radial locations, but its distribution is non-uniform when the flow is reduced due to the unsteady flow at the blades exit (Fig. 14(b)). The static pressure and total pressure increases quickly at guide-vane outlet as expected, shown in Figs. 14(c), 14(d). Furthermore, the static pressure remains nearly constant even at small flow conditions. With the effects of unsteady velocity, the total pressure fluctuation appears in the hub region.

Fig. 14. Hydrodynamic flow field at vane outlet at various locations at different flow rates

4 Conclusions

(1) The numerical results of the simulation and measurement of inner flow of the high efficiency axial pump show excellent agreement with the experimental results.

(2) It is found that the static pressure on pressure side of blades increases slightly at circumferential direction, and keeps almost constant at the same radial. On suction side, it increases gradually from inlet to exit along the flow direction.

(3) Experimental results show that inlet flow is almost axial and the prerotation is very small at various conditions.The meridional velocity and circulation distributions are almost identical at impeller outlet at design conditions.

(4) Hub leakage in adjustable blades leads to the decrease of the meridional velocity and circulation at impeller outlet near hub leakage region. The comparison test shows that the efficiency could be improved by 1.56%at blade angle Φ=0° without hub leakage.

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