Internal Flow, Vibration, and Noise Characteristics of a Magnetic Pump at Different Rotational Speeds

Abstract

A high-speed magnetic pump rated at 7800 r/min was studied. A numerical model was established, and a hydraulic, vibration, and noise testing system was set up to conduct flow simulations, noise, and vibration experiments at different speeds. The results show that increasing speed leads to a higher pressure difference between the pump chamber and the cooling circuit. Meanwhile, the turbulent kinetic energy at the impeller outlet increases. Despite an increase in energy loss, the loss ratio decreases, and overall efficiency improves. The internal flow noise collected by the outlet hydrophone mainly comes from Rotor–Stator Interference (RSI), and it can sensitively capture changes in rotational speed. The dominant frequency of the outlet noise agrees well with the blade frequency calculated from the set speed, with a maximum deviation of 0.26%. As the speed increases, the overall sound pressure level (OASPL) at the inlet and outlet and the Root Mean Square (RMS) acceleration values at the outlet and pump body generally increase, while the acceleration at the motor base shows a decreasing trend. The conclusions are helpful for the design and optimization of rotary machinery such as high-speed magnetic pumps.

1. Introduction

Magnetic pumps are specialized centrifugal pumps that use magnetic coupling instead of shaft seals to transmit torque. Therefore, they retain the hydraulic characteristics of centrifugal pumps while offering advantages such as leak-free operation, reduced maintenance, and suitability for hazardous fluids. Based on these features, magnetic pumps have been increasingly used in marine engineering, the petrochemical industry, and aerospace [1,2,3,4]. Changing the pump speed might lead to vibration and noise issues, thereby affecting pump performance and posing safety hazards [5,6]. Therefore, exploring the internal flow, vibration, and noise characteristics of high-speed magnetic pumps at different rotational speeds is crucial for ensuring their efficient and stable operation.

In recent years, scholars have conducted continuous research on the internal flow, vibration, and noise characteristics of centrifugal pumps.

In terms of internal flow in centrifugal pumps, advanced experimental techniques such as Particle Image Velocimetry (PIV) can be used to obtain velocity field information in key areas of the centrifugal pump. This provides direct experimental evidence for revealing complex internal flows, turbulent kinetic energy distribution, and rotor–stator interference characteristics in the pump [7,8]. However, factors such as the high cost of experimental systems and strict requirements for optical channels and transparent models limit its widespread application [9,10]. Meanwhile, Computational Fluid Dynamics (CFD) methods are widely used for numerical simulation analysis of internal flows in centrifugal pumps [11,12]. Al-Obaidi [13] systematically compared how ten turbulence models perform in numerical simulations of centrifugal pumps, pointing out that the shear stress transport (SST) k-ω model exhibits good comprehensive performance in flow characteristic analysis and performance prediction. Chang et al. [14,15] conducted studies on the internal flow of centrifugal pumps based on the SST k-ω model. The results showed that a sweeping jet inlet helps suppress large-scale vortex structures and reduce energy loss, whereas increased wall roughness intensifies vortex evolution and energy dissipation, thereby decreasing pump efficiency. Cordisco et al. [16] combined experiments with three-dimensional CFD models to analyze the performance of pumps, pointing out that simulations based on the SST k-ω model can accurately predict near-wall flow and cavitation development processes within the pump. The SST k-ω model demonstrates good comprehensive performance in predicting internal flows in centrifugal pumps and has become one of the most commonly used turbulence models in current CFD research for pumps.

Regarding the noise characteristics of centrifugal pumps, Guo et al. [17] conducted experimental research on the radiated noise of a centrifugal pump. The results showed that as the rotational speed increased from 1700 to 2900 r/min, the sound pressure level (SPL) of the external radiated noise from the pump exhibited an upward trend. Li et al. [18] systematically analyzed the variation patterns of flow-induced noise in centrifugal pumps under different operating conditions by combining numerical simulations with experimental measurements. They pointed out that the numerical model exhibited good consistency with the experimental results. Jin et al. [19] designed a bionic blade inspired by the tubercle effect to reduce the noise. They verified its effectiveness through numerical simulations of flow and sound fields. The bionic vane achieved a noise reduction of up to 2.53 dB without compromising hydraulic performance. Lu et al. [20] collected noise signals from centrifugal pumps during different cavitation stages through experiments and analyzed the noise characteristics in conjunction with numerical simulations. They pointed out that the typical frequency range of noise caused by cavitation onset is 2–8 kHz. Wang et al. [21] analyzed the variation trends of internal pump noise under different flow rates and cavitation conditions using hydrophones at the inlet/outlet. They noted that an increase in flow rate significantly elevates the noise level. Existing research indicates that the noise generated by internal flow in centrifugal pumps primarily originates from complex flows such as rotor–stator interaction (RSI), pressure pulsations, and turbulence.

Regarding the vibration characteristics of centrifugal pumps, Zhang et al. [22] proposed a numerical method for predicting flow-induced vibration using CFD–rotor dynamics sequential coupling. They pointed out that the blade passing frequency (BPF) is the dominant frequency of radial force fluctuations in the impeller. Wang et al. [23] studied pressure pulsations in a multistage pump through numerical simulations and experimental measurements. They noted that rotor–stator interference (RSI) can induce intense pressure pulsations, resulting in severe vibration and corresponding noise. Han et al. [24] utilized a three-axis acceleration sensor to detect off-design conditions and cavitation states, indicating that vibration is more suitable for detecting cavitation at design and high flow rates. Zhang et al. [25] combined numerical modal analysis with multi-point vibration tests to investigate the vibration frequency domain characteristics of high-pressure multistage centrifugal pumps under different flow rates. They noted that the vibration peaks occur at the shaft frequency and its multiples. Luo et al. [26] utilized pump vibration signals to achieve non-invasive speed measurement, with a measurement error rate of less than 0.27%. Current research indicates that BPF, shaft frequency, and its multiples are the main characteristic frequencies of vibration, and RSI is a key source that intensifies pressure pulsations and induces vibration.

Current research on the vibration and noise characteristics of pump equipment predominantly centers on conventional centrifugal pumps or low-speed magnetic pumps. In contrast, studies on high-speed magnetic pumps, particularly under variable speed conditions, remain relatively scarce. Few studies have reported experimental investigations on the vibration and internal flow noise of high-speed magnetic pumps. Although prior research has preliminarily established a correlation between speed variations and vibration noise, the specific influence patterns of different rotational speeds on the vibration intensity and noise sound pressure level at critical monitoring points of high-speed magnetic pumps have yet to be elucidated.

This study selects a high-speed magnetic pump with a rated speed of 7800 r/min as the research object. Firstly, using ANSYS 2021 and the SST k-ω model, a full flow field analysis model of the pump is established. The pressure distribution at the axial plane, velocity distribution at the axial plane, streamline distribution on the impeller mid-surface, and turbulent kinetic energy (TKE) distribution characteristics on the intermediate flow surface of the impeller at different rotational speeds are analyzed. Then, a hydraulic, vibration, and noise testing system is set up, and tests are conducted at five speeds ranging from 6200 to 7800 r/min. Finally, the frequency response curves of the sound pressure levels at the inlet and outlet, along with the variation in the RMS acceleration at the motor base, pump body, and pump outlet with rotational speed, are analyzed. The findings of this study can serve as a reference for optimizing magnetic pumps.

2. Materials and Methods

2.1. Numerical Simulation Method

2.1.1. Meshing

As shown in Figure 1, the high-speed magnetic pump is used in cooling and temperature reduction cycles and is characterized by leakage-free operation and a compact structure. The main structural components of the pump include the impeller, volute, and isolation sleeve. The isolation sleeve separates the pumped medium from the magnetic drive system, ensuring safe and reliable operation.

Its main design parameters are shown in

Table 1. Main design parameters.

Parameter	Unit	Value
Flow rate	m3/h	6
Rotational speed	r/min	7800
Head	m	90
Specific speed	-	39.78
Inlet diameter	mm	40
Outlet diameter	mm	27
Impeller inlet diameter	mm	33
Impeller outlet diameter	mm	100
Number of blades	-	6

.

The computational domain for the numerical simulation consists of seven components: the inlet pipe, de-swirler, impeller front cavity, impeller, impeller rear cavity, volute, and outlet pipe. Structured hexahedral meshes were generated using ICEM 2021, as shown in Figure 2.

Five different mesh models were selected, and numerical calculations were performed on these models under rated flow conditions. The relationship between hydraulic characteristics and mesh size is shown in Figure 3.

As shown in Figure 3, when the mesh size exceeds 2.29 million, the predicted values remain essentially unchanged. Therefore, a model with 2.29 million elements was selected.

2.1.2. Boundary Conditions and Turbulence Model

The numerical simulation was conducted using CFX 2021 software. The model of the pump was set with a pressure inlet and mass flow outlet, with the inlet pressure set at 1 atm. The convergence criterion was set at 10⁻⁵.

The steady-state calculation of the high-speed magnetic pump adopts the Frozen Rotor interface. The wall roughness was set to 15 μm. The transient calculation of the pump takes the steady state as the initial condition, with a time step of 4.27 × 10⁻⁵ s for every 2° rotation of the impeller.

In practical engineering contexts, all fluid flow phenomena must satisfy the fundamental conservation laws of mass, momentum, and energy. The flow investigated in the present study is a three-dimensional incompressible turbulent flow.

We employed the SST k-ω turbulence model for the simulation. It combines the strengths of the k-ε and k-ω models [27]. The governing equations [28,29]:

$${\displaystyle \frac{\partial \left(\rho k\right)}{\partial t}}+{\displaystyle \frac{\partial }{\partial x_i}}\left(\rho u_{i}k\right)={\displaystyle \frac{\partial }{\partial x_i}}\left[\left(\mu +{\displaystyle \frac{{\mu }_t}{{\sigma }_k}}\right){\displaystyle \frac{\partial k}{\partial x_i}}\right]+P_k-{\beta }^{*}\rho k\omega$$

(1)

$${\displaystyle \frac{\partial \left(\rho \omega \right)}{\partial t}}+{\displaystyle \frac{\partial }{\partial x_i}}\left(\rho u_i\omega \right)={\displaystyle \frac{\partial }{\partial x_i}}\left[\left(\mu +{\displaystyle \frac{{\mu }_t}{{\sigma }_{\omega }}}\right){\displaystyle \frac{\partial \omega }{\partial x_i}}\right]+D_{\omega }+\alpha {\displaystyle \frac{\omega }{k}}{P}_k-\beta \rho {\omega }^2$$

(2)

where $\rho$ is the fluid density, $u_i$ is the velocity component,$k$ is the turbulent kinetic energy, $\omega$ is the specific dissipation rate, $\mu$ is the molecular viscosity, ${\mu }_t$ is the turbulent viscosity, $P_k$ is the production term of turbulent kinetic energy, and $D_{\omega }$ is the cross-diffusion term. The constants $\varphi \left({\sigma }_k,{\sigma }_{\omega },\beta ,\alpha \right)$ are blended via $F_1$ as $\varphi =F_1{\varphi }_1+\left(1-F_1\right){\varphi }_2$, with inner $k\text{-}\omega$ values ${\sigma }_{k1}=1.176, {\sigma }_{\omega 1}=2.0, {\beta }_1=0.075, {\alpha }_1=5/9$; outer $k\text{-}\epsilon$ values ${\sigma }_{k2}=1.0, {\sigma }_{\omega 2}=1.168, {\beta }_2=0.0828, {\alpha }_2=0.44$; and a global ${\beta }^{*}$ = 0.09.

2.2. Test Apparatus and Test Method

2.2.1. Test Bench

A pump hydraulic performance, vibration, and noise test platform has been established, and the system composition is shown in Figure 4. The testing platform mainly consists of a pump circulation control system, a pump hydraulic performance testing system, and a high-frequency monitoring system for pump status.

The function of the pump circulation control system is to control the rotational speed, flow rate, and vacuum level. It is mainly composed of a 400 Hz frequency converter, a 400 Hz high-speed motor, a soft start control box, a low-voltage electrical distribution box, a stainless steel water storage tank, valves, pipelines, and a vacuum pump.

The function of the pump hydraulic performance testing system is to conduct external characteristic tests of pumps. It primarily consists of inlet and outlet low-frequency pressure sensors, speed sensors, flowmeters, and pump hydraulic performance testing software Magnetic Pump Test 2016.

The main function of the high-frequency monitoring system for pump status is to collect and record information from various high-frequency sensors. It primarily consists of a 16-channel high-speed data acquisition device, three 3-axis acceleration sensors, two hydrophones, and data acquisition software YE7600.

Figure 5 depicts the actual experimental setup.

The hydrophones were installed at locations four pipe diameters away from the inlet and outlet, respectively. The acceleration sensors were installed at the outlet, on the pump body, and on the motor base. Figure 6 shows the installation locations of the acceleration sensors and hydrophones.

The acceleration sensors were fixed to the measurement points using M5 set screws and adhesive to ensure a secure installation and optimize vibration transmission. As shown in Figure 6, the X direction is the normal direction perpendicular to the Y–Z plane (the lateral direction of the pump body). The Y direction is the outlet direction of the centrifugal pump. The Z direction is the inlet direction of the centrifugal pump.

The hydrophone was installed at the measurement point using a hydrophone mounting assembly with a double-layer sealing ring. The mounting assembly is threaded and arranged at the measurement point of the test pipe. The hydrophone sensing surface is flush with the inner side of the pipe to ensure accurate measurement of acoustic pressure pulsations. Figure 7 shows the installation structure of the hydrophone.

The data acquisition device and main sensor parameters used in the pump experimental system are shown in

Table 2. Data acquisition instrument and sensor parameters for high-speed magnetic pump experimental system.

Equipment	Model	Measurement Item	Measurement Range	Accuracy
Pressure sensor	YB131	Inlet pressure	−0.1 MPa–0.1 MPa	±0.25%
Pressure sensor	YB131	Outlet pressure	0–1.2 MPa	±0.25%
Flow meter	LWGY-25	Rate of flow	1–12 m3/h	±0.25%
Speed sensor	UT37R	Rotational speed	0–20,000 r/min	±0.05%
Acceleration sensor	SAE3005	Vibration	0–50 g	±0.25%
Hydrophone	RHT-10	Internal sound field	20 Hz–20 kHz	±1 dB
Data acquisition instrument	SA6216	Signal acquisition	0–48 kHz	/

.

2.2.2. Test Method

The hydraulic performance, vibration, and noise characteristics of the high-speed magnetic pump were experimentally investigated using the test bench shown in Figure 4.

Before the experiment, the pipeline connections of the test system were carefully checked to ensure reliable sealing and stable operation. The pump was then started through the frequency converter and operated under the specified working conditions. After the system reached stable operating conditions, the measurements were carried out.

The experiments were conducted by adjusting the outlet valve and the frequency converter to obtain different operating conditions. Multiple operating points were tested at different rotational speeds.

At each operating condition, the pump was allowed to run for approximately 1 min to reach a stable state before data acquisition.

The hydraulic performance parameters, including flow rate, head, and input power, were recorded through the hydraulic performance monitoring system. Meanwhile, vibration and noise signals were collected using high-frequency sensors connected to a data acquisition system.

The sampling frequency for the vibration and acoustic signals was 24 kHz, and the acquisition duration for each measurement was 10 s. To reduce measurement errors, five sets of data were collected at each measurement point. All measurement data were automatically recorded and stored by the acquisition software YE7600.

2.2.3. Data Processing for High-Speed Magnetic Pump Experiment

Four key parameters were calculated from the measured data using Equations (3)–(7) and compared with simulation data.

(1) Head [30]

$$H=\left({\displaystyle \frac{P_2}{\rho g}}-{\displaystyle \frac{P_1}{\rho g}}\right)+\left({\displaystyle \frac{v_2^2}{2g}}-{\displaystyle \frac{v_1^2}{2g}}\right)+\left(Z_2-Z_1\right)$$

(3)

In the formula, P₁ and P₂ represent the inlet and outlet pressures (Pa); $v_1$ and $v_2$ denote the flow velocities in the pipelines (m/s); Z₁ and Z₂ represent the elevation heights (m); $v_1={\displaystyle \frac{Q}{A_1}}$, $v_2={\displaystyle \frac{Q}{A_2}},$ Q represents the flow rate (m³/h); A₁ and A₂ represent the cross-sectional areas of the pipelines (m²); $g=$ 9.81 m/s².

(2) Efficiency [30]

$$\eta =\rho gQH/1000P\times 100\%$$

(4)

In the formula, P represents the input power (kW); $\rho$ represents the medium density (kg/m³).

(3) Noise [31,32]

Sound Pressure Level (SPL) is defined with reference to the human ear’s response to sound, and its formula is

$$L_p=20{\mathrm{l}\mathrm{o}\mathrm{g}}_{10}\left({\displaystyle \frac{p_e}{p_{ref}}}\right)$$

(5)

In the formula, $p_e$ represents the RMS sound pressure; $p_{\mathrm{ref} }$ denotes the reference sound pressure.

The OASPL formula is

$$L_a=10{\mathrm{l}\mathrm{o}\mathrm{g}}_{10}\left(\sum _{i=1}^n {10}^{{\displaystyle \frac{L_{p,i}}{10}}}\right)$$

(6)

In the formula,$L_{p,i}$ represents the SPL of the i-th one-third octave band.

(4) Acceleration Root Mean Square [33]

$$G_{rms}=\sqrt{{\displaystyle \frac{\sum _{i=1}^n x_i^2}{n}}}$$

(7)

In the formula, $x_i$ denotes the acceleration value of each sample (g).

3. Results and Discussion

3.1. Experimental Validation

The simulation results were compared with the experimentally measured data of the pump, as shown in Figure 8.

As shown in Figure 8, from the perspective of head variation, both the experimental and simulated head curves exhibit a slow downward trend as the flow rate increases. The deviation between the predicted and experimental heads is small, with a relative error of 2.93% at the rated flow rate, indicating that the SST k-ω model has high prediction accuracy for head. Both the experimental and simulated efficiency curves increase with flow rate, suggesting that the investigated range falls within the rising region of the efficiency curve, with the simulated values slightly higher than the experimental ones. The trend of shaft power variation is characterized by both experimental and simulated curves showing a linear upward trend as the flow rate increases. The numerical results agree well with the experimental data, confirming the reliability of the SST k-ω model.

3.2. Internal Flow Analysis

The internal flow of the pump at different speeds at the rated flow rate (6 m³/h) was analyzed, specifically at 6200, 7000, 7400, and 7800 r/min, with 7800 r/min being the design rotational speed.

(1) Pressure Distribution

The pressure distribution at the axial plane at different speeds is shown in Figure 9. As the speed increases, the work capacity of the impeller of the pump is enhanced, and the pressure in the impeller passage and at the impeller outlet increases accordingly. At the same time, the driving pressure difference between the backflow in the front pump chamber and the cooling circulation loop gradually increases. At lower rotational speeds, the pressure increase cannot meet the design head requirements.

(2) Velocity Distribution

The velocity distribution at the axial plane at different rotational speeds is shown in Figure 10. As the rotational speed increases, the relative flow velocity of the medium within the impeller changes slightly. Due to the increase in driving pressure difference caused by the increase in rotational speed, the flow velocity in the front pump chamber and the cooling circulation loop slightly increases.

(3) Streamline Distribution

Figure 11 shows the streamlines of the intermediate flow surface of the impeller at different speeds.

As the speed increases from 6200 to 7800 r/min, the streamline distribution exhibits the following trends: (1) The spiral density of streamlines gradually increases, and the flow rate transitions from being dominated by low flow rates to a significant increase in high flow rates; (2) The flow stability gradually decreases, with vortices emerging from non-existent ones, growing from small to large, and expanding from local areas to the entire flow field; (3) The regularity of streamlines weakens, gradually changing from a uniform and smooth distribution to twisting and crossing, and the flow loss increases with the increase in speed. This is consistent with the physical mechanism whereby increasing impeller speed enhances centrifugal force, fluid kinetic energy, and boundary layer separation.

When the rotational speed is low, no vortex appears in the impeller passage, and the flow lines are relatively smooth. This is due to the large flow rate condition corresponding to the low rotational speed of the impeller. The flow passage in the pump is designed to accommodate large flow rates using the increased flow rate design method. As the rotational speed increases, the corresponding design flow rate increases, but the actual flow rate decreases. At high speeds, flow separation begins to occur on the back of the blades, and a vortex appears at the blade outlet.

(4) Turbulent Kinetic Energy Distribution

Figure 12 shows the TKE distribution on the intermediate flow surface of the impeller at different rotational speeds.

The turbulent kinetic energy changes slightly at different speeds. With increasing speed, the energy loss in the impeller passage tends to increase, but the increment is small. At the outlet of the impeller, due to the collision with the fluid flow in the volute, the TKE is relatively high; with increasing speed, the TKE at the outlet of the impeller increases rapidly, indicating that the fluid flow collision becomes more intense and the energy loss increases. According to actual experimental data, at a rated flow rate of 6 m³/h, the pump efficiency at 7800 r/min is 0.25% higher than that at 6200 r/min, indicating that although the absolute value of energy loss increases at high speeds, the proportion of loss decreases, and the overall efficiency of the pump is improved.

3.3. Internal Flow Noise Characteristics

Experiments on hydraulic performance and noise characteristics were conducted at different speeds using the test bench. The hydraulic performance of the magnetic pump is shown in Figure 13.

As shown in Figure 13, with the increase in rotational speed, the head increases rapidly; the maximum efficiency point of the magnetic pump tends towards the direction of large flow rate, and the high-efficiency range is significantly expanded.

The change in pump speed will affect the vibration and noise characteristics. Figure 14 shows the sound pressure level frequency response characteristic curves of the inlet and outlet monitoring points at five different speeds: 6200, 6600, 7000, 7400, and 7800 r/min.

As shown in Figure 14, the frequency response curve of SPL at the outlet shows an overall trend of first increasing and then decreasing, which is consistent with the trend of external radiated noise reported by Guo et al. [17]. At different rotational speeds, the dominant frequencies of SPL at the inlet and outlet are 619.7 Hz, 658.3 Hz, 698.5 Hz, 739.7 Hz, and 779.1 Hz.

Table 3. Comparison of blade frequency and dominant outlet noise frequency.

Speed/(r/min)	Blade Frequency/Hz	Dominant Noise Frequency/Hz	Relative Error/%
6200	620	619.7	−0.05
6600	660	658.3	−0.26
7000	700	698.5	−0.21
7400	740	739.7	−0.04
7800	780	779.1	−0.12

compares the blade frequency calculated based on the set rotational speed with the dominant outlet noise frequency at different speeds.

Table 3 shows that at five different rotational speeds, the relative error between the dominant frequency at the outlet monitoring point and the blade frequency calculated based on the set rotational speed is small. The maximum relative error occurs in the test at 6600 r/min, which is 0.26%. The experimental data at different rotational speeds indicate that the internal flow noise collected by the hydrophone at the outlet is mainly caused by the interference between the internal dynamic and static components of the pump. The hydrophone can sensitively capture and reflect changes in rotational speed. The frequency converter and magnetic drive coupling work stably and reliably, capable of driving the pump to operate at the set rotational speed.

Figure 15 shows the comparison of the OASPL at the inlet and outlet.

As shown in Figure 15, with increasing rotational speed, the OASPLs at the inlet and outlet monitoring points of the magnetic pump gradually increase, reaching maximum values of 155.8 dB(A) and 176.2 dB(A) at 7800 r/min, respectively. The OASPL difference between the inlet and outlet gradually decreases, reaching a minimum value of 20.4 dB(A) at 7800 r/min.

3.4. Vibration Characteristics

Table 4. RMS values of acceleration at different rotational speeds.

Channel	Direction	Grms/g
		6200	6600	7000	7400	7800
1	A–Y	0.54	0.64	0.77	0.85	1.16
2	A–X	0.67	0.60	0.63	0.70	1.04
3	A–Z	0.78	0.81	0.88	0.99	1.27
4	B–Y	0.66	0.79	0.86	1.01	1.26
5	B–X	0.60	0.77	0.91	0.77	1.05
6	B–Z	0.94	1.00	1.18	1.49	1.99
7	C–X	0.23	0.45	0.25	0.52	0.39
8	C–Z	0.88	0.84	0.27	0.34	0.32
9	C–Y	1.39	1.28	0.81	0.98	0.83

shows the RMS acceleration values of the magnetic pump at the rated flow rate (6 m³/h) and five rotational speeds (6200, 6600, 7000, 7400, and 7800 r/min).

According to Table 4, at speeds of 7000, 7400, and 7800 r/min, the maximum RMS values are observed at the B–Z monitoring point on the pump body in the direction of the inlet pipe. At 6200 and 6600 r/min, the maximum values occur at the C–Y base in the direction of the outlet pipe.

Figure 16 shows the comparison of RMS acceleration values at different rotational speeds.

Figure 16a shows the variation trends of the RMS values measured by the three acceleration sensors in the X direction. As the rotational speed increases, the RMS values of Sensors A and B generally increase, whereas Sensor C exhibits fluctuations within a certain range.

Figure 16b and Figure 16c present the variation trends of the RMS values measured by the three acceleration sensors in the Y and Z directions, respectively, and the trends in these two directions are generally similar. As the rotational speed increases, the RMS values of Sensors A and B in both the Y and Z directions show an increasing trend. In contrast, the RMS values of Sensor C in these two directions show an overall decreasing trend, with a more pronounced decrease occurring as the rotational speed increases from 6600 r/min to 7000 r/min.

Taken together, Figure 16 indicates that the RMS acceleration values measured in different directions at Sensor A (pump outlet) and Sensor B (pump body) generally increase with rotational speed. Previous numerical simulation results showed that increasing rotational speed weakens the regularity of the streamlines in the impeller region, promotes the formation and development of vortical structures, and significantly increases the turbulent kinetic energy at the impeller outlet. This, in turn, enhances flow interaction and intensifies internal flow instability. In addition, previous analyses of internal flow noise showed that both the dominant frequency and the overall sound pressure level at the outlet increase with rotational speed. Since the dominant frequency corresponds to the blade-passing frequency, it indicates that rotor–stator interaction is the primary source of internal flow noise, and its energy level further increases at higher rotational speeds.

For Sensor C (motor base), as the rotational speed increases, the RMS acceleration value in the X direction (lateral direction of the pump body) shows only slight fluctuations, whereas the RMS acceleration values in the Z direction (inlet pipe direction) and Y direction (outlet pipe direction) show an overall decreasing trend. This phenomenon may be related to the operating state of the system and the vibration transmission characteristics. On the one hand, as the rotational speed increases, the motor gradually approaches its rated speed (7800 r/min), resulting in a more stable operating condition. On the other hand, because the magnetic drive pump adopts magnetic coupling transmission, the transmission of vibration from the pump body to the external structure may be reduced. As a result, the vibration responses of Sensor C in the Y and Z directions show a decreasing trend.

Similar observations have also been reported in previous studies. Reference [25] indicates that vibration responses at different locations, such as the pump, pipeline, and support structure, vary significantly with changes in flow rate. Moreover, Reference [5] demonstrates that the vibration response of a centrifugal pump depends not only on excitation intensity but also on structural transmission paths and modal participation.

4. Conclusions and Future Work

To investigate the influence of rotational speed on the internal flow, vibration, and noise characteristics of high-speed magnetic pumps, a high-speed magnetic pump with a rated rotational speed of 7800 r/min was selected as the research object. A numerical simulation analysis and calculation model for the high-speed magnetic pump was established, and a hydraulic, vibration, and noise testing system for the pump was set up. At the rated flow rate (6 m³/h), numerical simulations of the internal flow were conducted at different rotational speeds, together with noise and vibration tests. The main conclusions are as follows.

(1)

Numerical simulation results show that with increasing speed, the work capacity of the pump impeller increases, and the pressure in the impeller passage and at the impeller outlet also increases accordingly. At the same time, the driving pressure difference in the backflow in the front pump chamber and the cooling circulation loop of the pump gradually increases.

(2)

Numerical simulation and external characteristic test results indicate that the TKE at the impeller inlet changes only slightly with increasing speed, while the outlet TKE increases significantly. Although higher rotational speeds intensify flow interactions and energy dissipation, the relative loss decreases, resulting in improved overall efficiency.

(3)

The experimental results at different rotational speeds show that the outlet shows a higher SPL than the inlet. The hydrophone can sensitively capture and reflect changes in rotational speed. The deviation between the dominant outlet noise frequency and the blade frequency calculated based on the set rotational speed remains small, with a maximum deviation of 0.26% at 6600 r/min. As the speed increases, the OASPL at both the inlet and outlet monitoring points gradually increases.

(4)

The vibration test results at different rotational speeds show that as the speed rises from 6200 to 7800 r/min, the RMS values of accelerations in all directions of the pump body acceleration sensor B and the pump outlet acceleration sensor A generally exhibit an increasing trend. The RMS values of accelerations in the C–X direction at the base change slightly, while those in the C–Z and C–Y directions at the base show an overall decreasing trend. A possible reason is that, as the rotational speed increases, the motor gradually approaches its rated speed (7800 r/min), resulting in a more stable operating condition. Meanwhile, because the magnetic drive pump adopts magnetic coupling, the transmission of vibrations from the pump body to the external structure may be reduced.

(5)

The internal flow numerical simulation results at different rotational speeds can be corroborated by the experimental results obtained from pump body acceleration sensors and outlet hydrophones. Rotor–Stator Interference is an important excitation source of internal flow noise in the high-speed magnetic pump.

Future work will focus on the numerical simulation and experimental investigation of internal flow noise in magnetic pumps. In particular, the effects of different impeller design parameters and bionic surface structures on internal flow-induced noise will be studied. The objective is to further understand the mechanism of flow-induced noise generation and to provide guidance for improving the hydraulic performance and reducing vibration and noise of high-speed magnetic pumps.