Numerical Simulation and Model Test on Pressure Fluctuation and Structural Characteristics of Lightweight Axial Flow Pump

Abstract

In order to explore the relationship between the hydraulic performance, pressure pulsation, and structural characteristics of lightweight axial flow pumps, this paper takes the axial flow pump as the research object and analyzes pressure pulsation and structural characteristics by changing the blade length and thickness to realize the lightweight design of the axial flow pump impeller. Compared with the initial scheme (Scheme 1), the mass of the impeller is reduced by 17.2% (Scheme 2) and 29.9% (Scheme 3), respectively. The lightweight axial flow pump (Scheme 2 and Scheme 3) has a lower pressure pulsation amplitude at the impeller outlet monitoring point under large flow conditions. After the lightweight design of the axial flow pump impeller, the blade becomes shorter, the mass becomes lighter, and the cavitation performance will become worse. The maximum equivalent stress of Scheme 2 and Scheme 3 is increased by 2.8% and 31.9%, respectively, compared to Scheme 1. The maximum deformation of Scheme 3 increased by 27% compared to Scheme 1. This research can provide guidance for the lightweight optimization design of the axial flow pump.

1. Introduction

Axial flow pumps have the advantages of small footprint, light weight, and high specific speed, and are widely used in irrigation, water supply and drainage, and water transfer engineering [1,2,3]. Recently, many scholars have conducted extensive research on lightweight design. Aiming at minimizing the mass, Yao et al. [4] determined the specific location of the material removal area of the fuselage structure and realized the lightweight design of the fuselage structure. Liu et al. [5] studied the lightweight design of power battery boxes, and the results showed that introducing safety indicators during the dimension optimization process reduced the deformation of battery modules by 43.08%, and the weight of battery boxes by 20.81%, pursuing the goal of being lightweight while also safeguarding the device’s strength rigidity and security features. Dai et al. [6] explored the feasibility and comprehensiveness of applying BIM lightweight technology in water conservancy engineering. BIM lightweight means to compress the size of the BIM model as much as possible to improve data transmission and loading efficiency on the premise of ensuring the data quality and security of the BIM model. The research results showed that the use of a BIM lightweight collaborative design platform in the design stage can significantly improve the efficiency of 3D model design, while effectively improving the quality of construction management and ensuring the timely and high-quality completion of projects. In summary, research on lightweight design has primarily focused on the press machine, battery boxes, and BIM hydraulic engineering, and all have made some progress, but there is relatively little research on the lightweight design of axial flow pumps.

During the operation of the axial flow pump, the impeller is subjected to the action of periodic water pressure, which produces complex unstable flow and violent dynamic and static interference [7,8]. The interior of the axial flow pump contains three-dimensional unsteady turbulence, and numerous scholars have carried out significant research on the pressure pulsation and structure of axial flow pumps. Stosiak et al. [9] identified and analyzed the source of the pressure pulsation in detail, gave a simulation method of the pulsation flow, and gave the experimental results and proposed a method of reducing the pressure pulsation. Under the action of periodic water flow, the impeller blade and shafting will deform to a certain extent, and excessive deformation will easily lead to cracks, fatigue failure, and even the fracture of the blade. Therefore, the analysis of the structural stress of the impeller blade of the axial flow pump can provide a theoretical basis for the optimization of impeller blade structure and the safe and reliable operation of the axial flow pump station [10,11,12,13]. Zhu et al. [14] analyzed the operation stability of large vertical axial flow pump units based on fluid–structure coupling. The results show that the blade frequency is consistent with the pressure fluctuation frequency, and the impeller rotation is the main cause of the pressure fluctuation. The maximum deformation value of the blade occurs at the blade rim, and the maximum stress of the blade occurs at the connection between the blade and the hub. Cieślicki et al. [15] studied the problem of the pump performance decreasing with the increase in downstream pressure and the increase in pump component deformation, and the influence of pump deformation on the height of circumferential clearance was investigated. He et al. [16] studied the correlation between pressure pulsation and dynamic stress in axial flow pumps based on fluid–structure coupling. The results indicate that the maximum displacement of the axial flow pump blades under coupling occurs at the inlet edge of the blades, and the displacement at the root of the blades is relatively small. The stress and deformation of the blades decrease gradually with the increase in axial flow pump flow. The frequency of water flow pressure fluctuation before the impeller inlet is mainly determined by the frequency of the impeller blades. Li et al. [17] studied the instability characteristics of the leakage flow in impeller rim clearance. They found that blade frequency is the main pressure pulsation frequency in the impeller area, and the pressure pulsation amplitude of the inner wall of the blade inlet impeller decreases with the expansion of the clearance of the rim. Zhou et al. [18] studied the pressure pulsation and fluid–structure coupling of large axial flow pumps for reverse power generation. The results show that the total blade deformation is mainly distributed on the inlet side of the blades, and its deformation increases gradually along the direction from hub to rim. The stress is mainly concentrated at the root of the blade, and the maximum equivalent stress occurs at the root of the suction surface of the blade. The maximum equivalent stress value is within the acceptable range of the blade material and does not affect the service life and safety of the turbine operating unit.

It can be concluded that research on the lightweight design of axial flow pumps has important implications for saving materials and energy, but the parameters governing the impact of lightweight design on the hydraulic and structural performance of axial flow pumps are not yet clear. This paper takes the axial flow pump as the research object, and conducts CFD simulations and experimental studies to achieve the goal of a lightweight axial flow pump by comparing the pressure pulsation amplitude and structural stress under different conditions. The research results can provide a reference for the design of axial flow pumps and the stable and safe operation of pumping stations.

2. Numerical Calculation and Experimental Verification

2.1. Calculation Model and Grid Division

The axial flow pump model is shown in Figure 1, which mainly includes the inlet pipe, impeller, guide vanes, and outlet elbow. The pump design parameters are the design flow rate Q = 360 L/s, the design head H_des = 6.0 m, the rotational speed n = 1450 r/min, the impeller diameter D = 300 mm, and the hub ratio d_h = 0.433; the number of impeller blades and guide vane blades are 4 and 7, respectively. The impeller and guide vanes are divided into structured grids using Turbo-Grid, and the inlet and outlet pipes are divided using ICEM CFD, as shown in Figure 2. The GCI (Grid Convergence Index) criterion of the Richardson extrapolation method [19] was adopted to verify the convergence of the grid. The calculation results are shown in

Table 1. Calculation of grid discretization error.

Parameter	Numerical Value
N1, N2, N3	2,911,756, 1,316,776, 573,075
r21, r32	1.303, 1.320
ϕ1, ϕ2, ϕ3	85.12%, 84.95%, 84.46%
p	3.647
${\varphi }_{\mathrm{e}\mathrm{x}\mathrm{t}}^{21},{\varphi }_{\mathrm{e}\mathrm{x}\mathrm{t}}^{32}$	85.22%, 85.21%
$e_a^{21},e_a^{32}$	0.2%, 0.56%
$e_{\mathrm{e}\mathrm{x}\mathrm{t}}^{21}, e_{\mathrm{e}\mathrm{x}\mathrm{t}}^{32}$	0.12%, 0.31%
${\mathrm{G}\mathrm{C}\mathrm{I}}_{\mathrm{f}\mathrm{i}\mathrm{n}\mathrm{e}}^{21},{\mathrm{G}\mathrm{C}\mathrm{I}}_{\mathrm{f}\mathrm{i}\mathrm{n}\mathrm{e}}^{32}$	0.154%, 0.22%

Note: N is the total number of mesh cells; r is the mesh refinement ratio; ϕ is the mesh error evaluation variable; e_a is the approximate relative error; e_ext is the relative error of the extrapolated values; and GCI_fine is the mesh convergence index.

below. The final number of each component is obtained by grid independence analysis. Under the premise of these quantities, the number of grids does not have an excessive effect on the efficiency and head of the pump. The final number of grid units for each flow component of the axial flow pump was determined to be 541,923 for the inlet pipe, 757,101 for the impeller, 730,296 for the guide vanes, 882,436 for the outlet elbow, and 2,911,756 for the total number of grids. The grid discretization error is 0.154%, which meets the calculation requirements.

2.2. Turbulence Models and Boundary Conditions

In the process of numerical simulation, the three-dimensional Reynolds averaged N-S equation is enclosed by the SST k-ω turbulence model. The inlet and outlet boundaries adopt total pressure (1at) inlet and mass flow outlet, the wall surface adopts non slip boundary conditions, and the impeller rim adopts a reverse rotation wall surface. The influence of wall roughness in the numerical calculation of internal turbulence in axial flow pumps is neglected. In steady-state calculations, the result of 1200 iterations is used as the initial value for unsteady calculations to improve computational accuracy. In an unsteady state, the interface between the dynamic and static domains adopts “Transient Rotor Stator” with a time step of 3.4483 × 10⁻⁴ s. In the unsteady calculation, the calculation time of 8 cycles of impeller rotation is taken, and the data from the last two cycles is selected for unsteady state pressure pulsation analysis [20,21,22].

2.3. Layout of Pressure Pulsation Monitoring Points

The pressure pulsation monitoring points at each interface are uniformly spaced from the hub to the rim. Monitoring points P1~P4 are arranged at the impeller inlet, P5~P8 are arranged in the middle of the impeller, and P9~P12 are arranged at the impeller outlet. The layout of each monitoring point is shown in Figure 3. To eliminate the influence of static pressure, the rotational frequency f_n of the blades and the dimensionless pressure fluctuation coefficient C_p are used to characterize the internal pressure fluctuation size of the pump [23,24,25]. The pulsation coefficient is effective under the condition of amplitude and the influence of wave phenomenon on the amplitude of pressure pulsation is considered. The specific expression is as follows:

$$C_{\mathrm{p}}={\displaystyle \frac{p-\overline{p}}{0.5\rho u^2}}$$

(1)

$$f_{\mathrm{n}}={\displaystyle \frac{60F}{n}}$$

(2)

In the formula, p is the transient pressure value at the monitoring point, Pa, $\overline{p}$ is the average transient pressure value, Pa, ρ is the density, kg/m³, u is the circumferential velocity at the impeller outlet, m/s, F is the frequency after Fourier transform, Hz, and n is the impeller rotation speed, r/min.

2.4. Determination of the Schemes

Scheme 1 is the initial scheme of the experiment, and Scheme 2 and Scheme 3 are the lightweight schemes of the axial flow pump compared with Scheme 1. The design parameters of the axial flow pump impeller with different schemes are shown in

Table 2. Axial flow pump impeller schemes under different parameters.

Scheme	σ1	Zm	Th/mm	Tr/mm	m/kg	△m/%	Hdes/m	ηmax/%
Scheme 1	1.0	1.433	14	6	3.122	0	6.0	84.25
Scheme 2	0.85	1.45	14	6	2.586	17.2	6.0	85.28
Scheme 3	0.85	1.45	12	5	2.187	29.9	6.0	85.67

Note: σ₁ is the density of the blade tip cascade and Z_m is the multiple of the density of the cascade; m is the mass of the blade; △m is the mass decline compared to scheme 1; Th is the max thickness of the hub airfoil; Tr is the max thickness of the rim airfoil; and η_max is the maximum efficiency of the pump.

. Scheme 1 is the original impeller design scheme with relatively long blades. Scheme 2 is to change the cascade design parameters of the blade based on Scheme 1, so that the blade is shorter and the mass is lighter. Because the design head of the impeller will be reduced after the blade becomes shorter, in order to ensure that the design head of the impeller does not change, the blade placement angle of Scheme 2 is adjusted to a positive angle. On the basis of Scheme 2, Scheme 3 keeps the length of the blade unchanged, reduces the thickness of the blade, further makes the mass of the blade lighter, and further realizes the lightweight design of the impeller. It can be seen from the table that the maximum efficiency of the axial flow pump gradually increases with the decrease in mass. The highest efficiency reached 85.67%.

2.5. Test Verification

The model experiment was conducted on a high-precision hydraulic machinery test bench. The comprehensive system error of the pump or pump unit efficiency of the test bench was ±0.39%. The impeller of the axial flow pump was made of brass material, while the guide vanes were formed from steel material. The water temperature was 25 °C. The electromagnetic flowmeter was used in the axial flow pump model test, and the test accuracy was 0.22%. The differential pressure sensor is used in the head test, and the test accuracy is 0.035%. The speed measurement and torque measurement use a speed torque sensor, which has a test accuracy of 0.01%. The physical model test bench of the axial flow pump is shown in Figure 4 below; the comparison of hydraulic performance characteristics and experimental results under different schemes is shown in Figure 5.

It can be seen that, under low flow conditions, the head of lightweight Scheme 2 and Scheme 3 is slightly reduced compared to Scheme 1, but the efficiency difference is not significant. Under high flow conditions, the head of lightweight Scheme 2 and Scheme 3 has increased compared to Scheme 1, but the difference is not significant, with higher efficiency and a wider operating range in the high-efficiency zone. Under the design conditions, the head of each scheme is basically the same, and the efficiency of Scheme 2 and Scheme 3 is higher than that of Scheme 1. Under the condition of large flow rate, the numerical simulation of Scheme 1 has a smaller head than the experimental value, and the efficiency agreement is better, and the maximum head error is within 5%. Under the condition of large flow, the numerical simulation of Scheme 1 has a higher head than the experimental value. Under the design conditions, the head of Scheme 1 is 6.0 m, and the efficiency error does not exceed 1.7% compared to the experimental value.

In general, from Scheme 1 to Scheme 2, changing cascade parameters has a great influence on the hydraulic performance of the axial flow pump under off-design conditions. From Scheme 2 to Scheme 3, changing the thickness of the airfoil will not affect the hydraulic performance of the axial flow pump. The numerical simulation results are consistent with the experimental results, indicating that the numerical simulation results are reliable.

3. Pressure Pulsation Analysis

3.1. Pressure Pulsation of Impeller Inlet

Figure 6, Figure 7 and Figure 8 show the frequency domain diagrams of impeller inlet pressure pulsation under different schemes. It can be observed that the main frequency of pressure pulsation at each monitoring point in different schemes is four times the rotational frequency, which is the blade frequency. This indicates that the pressure pulsation at the inlet of the impeller is mainly affected by the rotational frequency of the impeller.

Table 3. Pressure fluctuation amplitude of the main frequency of impeller inlet under different schemes.

Scheme	Low Flow Condition				Design Condition				High Flow Condition
	Monitoring Point				Monitoring Point				Monitoring Point
	P1	P2	P3	P4	P1	P2	P3	P4	P1	P2	P3	P4
Scheme 1	0.032	0.038	0.043	0.044	0.022	0.026	0.03	0.032	0.028	0.034	0.04	0.043
Scheme 2	0.03	0.037	0.044	0.047	0.021	0.027	0.031	0.033	0.022	0.027	0.033	0.035
Scheme 3	0.029	0.036	0.043	0.045	0.02	0.025	0.029	0.031	0.02	0.025	0.03	0.033

shows the amplitude of pressure pulsation at the main frequency of each monitoring point at the inlet of the impeller in different schemes. From the table, it can be seen that the pressure pulsation amplitude of each monitoring point in different schemes gradually increases from the hub to the rim at the main frequency. Overall, the pressure pulsation amplitude of each monitoring point in Scheme 3 is relatively small. The main frequency amplitude of the pressure pulsation in the impeller inlet area is basically the same under the low flow condition and the design condition. However, the main frequency amplitude of the pressure pulsation at the impeller inlet of the lightweight axial flow pump is obviously smaller than that of the original scheme under the large flow condition.

3.2. Pressure Pulsation of Middle of Impeller

Figure 9, Figure 10 and Figure 11 show the frequency domain diagrams of pressure pulsation in the middle of the impeller under different schemes. It can be seen that under different schemes, the main frequency of pressure pulsation in the middle of the impeller is four times the rotational frequency, indicating that the pressure pulsation in the middle of the impeller is still mainly affected by the rotational frequency of the impeller.

Table 4. Pressure fluctuation amplitude of the main frequency of the impeller middle under different schemes.

Scheme	Low Flow Condition				Design Condition				High Flow Condition
	Monitoring Point				Monitoring Point				Monitoring Point
	P5	P6	P7	P8	P5	P6	P7	P8	P5	P6	P7	P8
Scheme 1	0.103	0.126	0.151	0.168	0.102	0.112	0.124	0.133	0.09	0.085	0.081	0.07
Scheme 2	0.117	0.143	0.168	0.176	0.11	0.123	0.137	0.138	0.091	0.089	0.088	0.083
Scheme 3	0.119	0.143	0.169	0.185	0.111	0.122	0.136	0.145	0.091	0.087	0.086	0.082

shows the amplitude of pressure pulsation at the main frequency of each monitoring point under different schemes. It can be seen that the pressure fluctuation amplitude at the main frequency of each monitoring point in the middle of the impeller under different schemes gradually decreases with the increase in flow rate. Under low flow and design conditions, the pressure pulsation amplitude at the main frequency of each monitoring point in different schemes gradually increases from the hub to the rim, while under high flow conditions, it shows a gradually decreasing trend. This may be due to poor water flow pattern at the hub under high flow conditions, and the unstable flow leads to greater pressure pulsation. Overall, Scheme 1 has the smallest amplitude of pressure pulsation. The results indicate that the pressure pulsation of the middle monitoring point of the impeller will increase after the blade is lightened. Comparing Table 3 and Table 4, it can be found that under design conditions, the ratio of pressure fluctuation amplitude at the main frequency in the middle of the impeller to that at the impeller inlet is the highest, with ratios of 6, 6.58, and 7.25 for each scheme, respectively. This indicates that the pressure fluctuation amplitude in the middle of the impeller is about 6–7 times that at the impeller inlet. This further shows that the pressure pulsation at the middle monitoring point of the impeller is mainly affected by the pressure difference between the front and back of the blade.

3.3. Pressure Pulsation of Impeller Outlet

Figure 12, Figure 13 and Figure 14 show the frequency domain diagrams of pressure pulsation at the impeller outlet under different schemes. Under low flow conditions, each scheme exhibits significant low-frequency pulsation, which may be influenced by blade backflow.

Table 5. Pressure fluctuation amplitude of the main frequency of impeller outlet under different schemes.

Scheme	Low Flow Condition				Design Condition				High Flow Condition
	Monitoring Point				Monitoring Point				Monitoring Point
	P9	P10	P11	P12	P9	P10	P11	P12	P9	P10	P11	P12
Scheme 1	0.0107	0.0158	0.0096	0.01	0.0072	0.0078	0.0085	0.0097	0.0091	0.0122	0.0137	0.0152
Scheme 2	0.018	0.022	0.0127	0.0076	0.0026	0.0032	0.0054	0.006	0.005	0.0092	0.0118	0.0129
Scheme 3	0.0254	0.022	0.0141	0.0102	0.0012	0.0053	0.0078	0.0084	0.0046	0.0095	0.0122	0.0142

shows the amplitude of pressure pulsation at four times the rotational frequency for each monitoring point under different schemes. It can be seen that under low flow conditions, Scheme 1 has the smallest pressure fluctuation amplitude at four times the rotational frequency, followed by Scheme 2, and Scheme 3 has the largest amplitude. The pressure pulsation amplitude on the hub side of the blade outlet is greater than that on the rim side, indicating that the degree of pulsation on the hub side is greater than that on the rim side. This may be due to the poor flow state on the hub side under low flow conditions.

Under design and high flow conditions, the pressure pulsation amplitude at four times the rotational frequency from the hub to the rim is gradually increasing at each monitoring point in each scheme. The amplitude of pressure pulsation in Scheme 2 is the smallest, Scheme 3 is the second, and Scheme 1 is the largest. This also shows that the lightweight axial flow pump has a lower pressure pulsation amplitude at the impeller outlet monitoring point under large flow conditions, and the corresponding axial flow pump has a higher efficiency and the flow field at this position is more uniform.

3.4. Pressure Distribution Analysis

Figure 15 and Figure 16 show the pressure cloud maps of the impeller pressure surface and suction surface for different schemes under design conditions. It can be seen that under the design conditions, the pressure distribution on the impeller pressure surface of each scheme is uniform, the pressure gradually increases from the hub to the rim, and the pressure gradient changes uniformly along the spanwise direction, resulting in a good flow field. Each scheme has a large low-pressure area on the suction surface, which is also the reason for the cavitation of the axial flow pump. That is to say, after the lightweight design of the axial flow pump impeller, the blade becomes shorter, the mass becomes lighter, and the cavitation performance will become worse. Overall, among the various schemes, Scheme 3 has the highest pressure difference between the pressure surface and the suction surface, the strongest impeller power generation ability, and the relatively highest efficiency of the axial flow pump.

4. Structural Analysis

4.1. Solid Domain Computing Model and Grid Division

The solid domain calculation model is shown in Figure 17. The shaft system is made of stainless steel with a density of 7850 kg/m³; Young’s modulus is 200 MPa, Poisson’s ratio is 0.3, tensile yield strength is 250 Mpa, and tensile ultimate strength is 460 Mpa. There are rubber bearing constraints at the upper and lower ends of the pump shaft.

The solid domain is divided into tetrahedral meshes. Considering the stress concentration at the root of the blade and the maximum deformation at the tip, in order to improve the calculation accuracy and precision of the entire model, all edges and surfaces of the impeller blade will be finely controlled by local mesh refinement, as shown in Figure 18. Grid independence and convergence analysis were conducted; the final size of the edge grid was determined to be 4 mm, 1 mm for the blade tip and 6 mm for the volume grid. The maximum error of stress and strain did not exceed 0.6%, and the grid reached convergence. The final determined total number of grids was 502,607, with 735,745 nodes, an average grid quality of 0.83, an aspect ratio of 1.9, a Jacques ratio of 3.95, and an average vertex angle of 96°.

4.2. Fluid–Structure Coupling Constraints

The motion constraints and loads of the axial pump shaft system are shown in Figure 19. The axial flow pump is subjected to inertia forces such as self-weight and rotating force, and applies gravity acceleration to the solid domain of the whole shaft system, with a magnitude of 9806.6 mm/s² and a direction of application along the negative z-axis. The rotation force is loaded by the rotation speed of the pump shaft, which is 1450 r/min. Fixed constraints are applied at point A at the top of the pump shaft, with rubber bearings at points B and C, using cylindrical constraints. The cylindrical surface constrains the radial motion of the shaft system, namely compression deformation, and tangential upward motion. The hydraulic pressure loads are applied at points D, E, and F. The data transmission of fluid–structure coupling occurs at the coupling interface.

4.3. Blade Stress–Strain Analysis

From Figure 20, Figure 21 and Figure 22, it can be seen that the maximum equivalent stress distribution of the blades under different working conditions and schemes is mainly concentrated near the root of the hub blade, and the maximum equivalent stress from the hub to the rim shows a gradually decreasing trend. At the same time, it can be found that under low flow rate and design conditions, the maximum equivalent stress for Scheme 1 and Scheme 2 is basically the same, while Scheme 3 has the maximum equivalent stress. This indicates that the variation in blade density has little effect on the maximum equivalent stress of the blade, while reducing blade thickness will increase the maximum equivalent stress of the blade. Overall, the maximum equivalent stress of the blades in each scheme has not yet reached the material’s tensile yield strength and ultimate strength, so the blades in each scheme can meet the structural safety requirements.

From Figure 23, Figure 24 and Figure 25, it can be seen that the maximum deformation of the blades is mainly concentrated near the inlet blade tip, and the maximum deformation from the hub to the rim blade shows a gradually increasing trend. Under high flow conditions (Q = 440 L/s), due to the high density of the blade tip cascade and complex blade vibration modes in Scheme 1, it will seriously affect the safe operation of the axial flow pump, and this is also the reason for the low efficiency of Scheme 1.

From Figure 26, it can be seen that the maximum equivalent stress and maximum deformation of the axial flow pump blades under different schemes gradually decrease with the increase in flow rate. The curves of stress and deformation with the flow are similar to the curve of the head with the flow. Comparing Scheme 1 and Scheme 2, it can be seen that reducing the density of the blade gate has a relatively small impact on the maximum equivalent stress and maximum deformation of the blade. However, comparing Scheme 2 and Scheme 3, it can be seen that reducing the thickness of the blade will increase the maximum equivalent stress and maximum deformation of the blade. This shows that when the thickness of the blade is unchanged, changing the length of the blade has little effect on the maximum stress and strain of the impeller, but reducing the thickness of the blade will make the structural characteristics of the impeller become very poor. In summary, Scheme 2 is the better option, with a 17.2% decrease in blade mass compared to Scheme 1, saving materials. Under optimal operating conditions, the efficiency of the axial flow pump has increased by 1%, and the operating range of the high-efficiency zone of the axial flow pump has been expanded.

5. Conclusions

In this paper, the lightweight design of the axial flow pump impeller is realized by changing the blade length and thickness. Compared with the initial scheme (Scheme 1), the mass of the impeller is reduced by 17.2% (Scheme 2) and 29.9% (Scheme 3), respectively. At the same time, it ensures that the design flow and the design head of the axial flow pump are the same. In this paper, the hydraulic performance, unsteady pressure pulsation characteristics, and impeller structural characteristics are compared and analyzed in detail for the above three schemes. The main conclusions are as follows:

(1)

Under low flow conditions, the head of Scheme 2 and Scheme 3 decreases compared to Scheme 1, but the efficiency difference is not significant. Under design conditions, Scheme 2 and Scheme 3 have increased efficiency by 1% compared to Scheme 1. Under high flow conditions, there is not much difference in the head of each scheme. Scheme 2 and Scheme 3 have increased efficiency compared to Scheme 1, and have expanded the operating range of the high-efficiency zone of the axial flow pump.

(2)

The amplitude of pressure pulsation at the monitoring points in the middle section of the impeller is greater than that at the inlet and outlet, because the pressure pulsation at the middle section of the impeller is mainly affected by the pressure difference between the front and back of the blade. The main frequency of pressure pulsation amplitude at monitoring points of different sections of the impeller is the blade frequency, which is four times the rotational frequency. This lightweight axial flow pump (Scheme 2 and Scheme 3) has a lower pressure pulsation amplitude at the impeller outlet monitoring point under large flow conditions. After the lightweight design of the axial flow pump impeller, the blade becomes shorter, the mass becomes lighter, and the cavitation performance will become worse.

(3)

The maximum equivalent stress and maximum deformation of axial flow pump blades gradually decrease with the increase in flow rate. The maximum equivalent stresses for Scheme 1, Scheme 2, and Scheme 3 are 31.065 MPa, 31.946 MPa, and 40.974 MPa, and the maximum deformations are 0.178 mm, 0.163 mm, and 0.226 mm, respectively. The maximum equivalent stress of Scheme 2 and Scheme 3 increased by 2.8% and 31.9%, respectively, compared to Scheme 1. The maximum deformation of Scheme 2 decreased by 8.4% compared to Scheme 1, while the maximum deformation of Scheme 3 increased by 27% compared to Scheme 1. This shows that when the blade thickness is unchanged, changing the blade length has little effect on the maximum stress and strain of the impeller, but reducing the thickness of the blade will make the structural characteristics of the impeller become very poor. All schemes can meet the safety requirements of blade structure.