Improvement of Noise Reduction Structure of Direct-Acting Pressure Reducing Valve

Abstract

As a key pressure control component of a hydraulic system, the noise of the direct-acting pressure reducing valve affects the working state of the system directly. However, the existing pressure reducing valves generally have the problem of excessive pure noise. In order to solve this problem, this study explored various structural combinations with the aim of improving the noise level of a direct-acting pressure reducing valve. Firstly, the flow field model of the direct-acting pressure reducing valve was established by using FEA (Finite Element Analysis), and the relationship between the flow field state and noise state was demonstrated through CFD (Computational Fluid Dynamics) simulation. Secondly, the position, number, and diameter of the oil holes on the valve spool were comprehensively analyzed, and the sound field analysis using LMS Virtual Lab was carried out. Finally, a prototype of the pressure reducing valve was manufactured, and the noise level before and after improvement was compared. The results showed that the effective sound pressure after improvement was reduced by 6.1% compared with that before at 50% of the opening, which verified the precision of the simulation model. The research results could provide a guideline for the design and improvement of direct-acting pressure reducing valves.

1. Introduction

Pressure reducing valves are widely used in hydraulic systems in engineering machinery, aerospace, and other fields. They are indispensable pressure regulating components, and their performance has a significant impact on the noise and safety of hydraulic systems. In practical applications, the pressure reducing valve will produce noise, vibration, leakage, and other problems. The noise of the direct-acting pressure reducing valve is mainly flow-induced noise, and other problems such as leakage have little effect on the noise. The noise level shows that the noise of the pressure reducing valve is poor, which has a serious impact on the normal operation of the equipment. In recent years, the problem of vibration and noise in pressure reducing valves has received increasing attention.

Therefore, it is particularly important to conduct research on noise reduction technology for pressure reducing valves [1]. Li S. et al. concluded through experiments that the noise in servo valves is caused by high-frequency oscillation of components under high-pressure differential and predicted the possible reasons for high-frequency oscillation and noise generation. Based on this, suggestions were proposed to reduce pressure pulsation and noise of hydraulic servo valves [2]. In numerical research, many research theories are based on acoustic simulation theory, the internal flow field of the machine is calculated by using CFD (computational fluid dynamics), and then based on the calculated flow field, the flow noise was analyzed using the FW-H (Ffowcs Williams and Hawkings) equation embedded in CFD [3,4]. According to the FW-H acoustic equation and internal flow characteristics, Jin HZ et al. found that the maximum noise level is located in the turbulence generation zone after hedging [5].

From the content of the research in this paper, the research on the noise of the pressure reducing valve is mainly divided into the following three aspects:

In the study of the influence of structure on valve noise, Hou J et al. installed a noise reduction sleeve on the balance valve and improved the sleeve opening and spacing, resulting in a total sound pressure level below 85 dB [6]. Wei et al. found that the noise source of a non-perforated plate valve is located at the bottom of the valve spool, while a perforated plate can effectively reduce noise [7]. Xu W et al. found that changing the structure and position of the inlet can reduce the noise of the control valve [8]. Janzen et al. tested the structural noise of the valve and found that the chamfered edge of the internal structure has a significant impact on the noise [9]. Zhang L et al. proposed a downstream noise reduction structure with a spiral guide vane and improved the model to achieve a maximum noise reduction effect of 13 dB [10]. Qiu et al. studied the effect of valve seat cone angle on the noise level of the cone valve and found that the noise was reduced by 18.2 dB [11]. Youn C. et al. developed a radial slit structure to reduce pressure, which reduced the noise level by approximately 40 dB [12]. Kong F et al. found that porous materials have a better noise reduction effect on high-frequency noise [13]. Chang Z et al. measured cavitation flow through orifice plates with different porosities and found that the sound pressure level significantly increased in the high-frequency range of 500–8000 Hz [14]. L.N.Quaroni et al. studied the noise generated when a low Mach number air flow passes through a circular hole in a rectangular air duct, and the influence of the number and position of the orifices maintaining a constant flow area is addressed [15].

In the study of the influence of pressure difference on noise, Qian J. et al. found that as the pressure ratio increases, the energy loss increases, and so does the noise when valve opening is constant [16]. Li et al. studied the application of magnetic fluid in a hydraulic jet pipe servo valve, and the results showed that the application of magnetic fluid suppressed the noise of the servo valve [17]. He et al., based on LES and Lighthill’s acoustic theory, combined large eddy simulation and a structural finite element model to address the noise issue caused by submerged cavity flow in a water tunnel. They developed a fluid–structure coupling method that reduces modeling and computational costs [18].

In order to reduce noise, some scholars have designed new structures. Zhao Y et al. designed a U-shaped groove structure sleeve that can reduce the flow noise of a 2D valve by 10 decibels [19]. Yu R et al. developed a noise reduction control valve with a U-shaped balance cage, which can effectively reduce noise [20].

Many scholars have studied the noise in the working process of hydraulic valves, including the influence of inclusion structure on valve noise and the influence of pressure difference on noise. At the same time, they also mentioned the design of new structures to reduce noise. Several research cases are provided in the research article, covering different structural designs and methods, which help to fully understand the valve noise problem. However, most of the research methods of valve noise suppression in the current research are single-factor optimization, and some research results lack quantitative data support, which affects the reliability of the conclusion.

This article aims to reduce the flow noise of pressure reducing valves and improve different structural combinations of valve cores. On the one hand, CFD is used to simulate and analyze the flow noise under different oil orifices on the valve spool, and the influence of various structural parameters on valve flow noise is obtained. On the other hand, experimental verification was conducted on the prototype to ensure the precision of the simulated structure. The main process of this paper is shown in Figure 1.

2. Direct-Acting Pressure Reducing Valve Model

2.1. Working Principles and Simulation Boundary Condition Analysis of Direct-Acting Pressure Reducing Valve

Figure 2 illustrates the operating principle of the direct-acting pressure reducing valve. A direct-acting pressure reducing valve is a type of pressure control valve that is primarily composed of a valve body, valve spool, valve seat, coil spring, and adjusting screw. It has an oil inlet, an oil return port, and a working port. The pressure at the working port is adjusted by the pressure set by the guide coil spring to reduce it. When the pressure at the working port is greater than the coil spring force, the valve core moves upward. When moving to a certain position, the working port is connected to the return port, and the fluid is discharged from the return port. Define the inlet port as P, the working port as A, and the return port as T.

2.1.1. Flow Field Analysis Setup

This paper analyzes the internal flow field and noise situation during the pressure reducing operation of a pressure reducing valve. The simplified fluid domain model of the pressure reducing valve was established by Fluent, as shown in Figure 3. Under the premise of ensuring that the flow principle of the flow field in the pressure reducing valve is unchanged, the main valve part of the pressure reducing valve had been simplified, and the clearance of the spool valve sleeve with little influence on the main flow field was removed, and the main flow field was retained. At the same time, when using the CFD method, the assumptions are as follows: (1) the oil is continuous and incompressible; (2) there is no impurity inside the oil; and (3) there is no wall slip. The standard k-ε turbulence calculation model is used for calculation.

In order to clarify the variation law of the flow field of the pressure reducing valve, it is necessary to select characteristic planes for analysis. As shown in Figure 4, two flow field planes had been selected to characterize the flow field condition of two asymmetric surfaces in the flow channel.

The research object of this paper was a direct-acting pressure reducing valve. Its internal flow field and noise situation were studied based on actual application conditions. The diameter of the oil hole at the P port of the valve sleeve was 3 mm, with a total number of 8. The boundary conditions for the inlet pressure were set to 100 bar, 150 bar, 200 bar, and 250 bar, with no slip on the wall. A port is the oil outlet, and the boundary condition for the outlet pressure is set to 25 bar. The fluid parameters are shown in

Table 1. Fluid parameters.

Parameter	Unit	Fluid
Density	kg/m3	890
Dynamic viscosity	Pa·s	0.04094

Density: The quality of No. 46 hydraulic oil used in this paper; Dynamic viscosity: The viscosity of No. 46 hydraulic oil used in this paper.

.

This study used the standard k-ε turbulent model for steady state calculation and the SAS SST turbulent model for transient calculation. The time step for transient calculation was set to 0.0001 s and the convergence precision to 10⁻⁴. According to the performance of this valve in the actual project, four typical inlet and outlet pressure conditions were selected, and the pressure difference was gradually expanded. The specific parameters are shown in

Table 2. Boundary condition.

Boundary Condition	1	2	3	4
Inlet pressure/bar	100	150	200	250
Outlet pressure/bar	25	25	25	25

.

The fluid domain model of the pressure reducing valve was established and imported into Fluent to mesh it. In order to improve the calculation accuracy, it was considered that the structure of the valve inlet flow channel and the outlet flow channel was simple, but the structure of the high-pressure chamber, the low-pressure chamber, and the valve core flow channel was complex. Therefore, it was necessary to locally refine the mesh at the complex parts of the high-pressure chamber, low-pressure chamber, and spool flow channel. The surface network was improved, and the maximum skewness angle was 0.677. As shown in Figure 5, select geometric shapes to generate a body network only from the fluid region without void, in which the number of grids in the fluid domain of the pressure reducing valve was 161,518 and the number of nodes was 59,362. The mesh quality was 0.8.

2.1.2. Sound Field Setting

The pressure fluctuation information obtained from the simulation of the pressure reducing valve flow field was imported into the structural field to generate the boundary element grid of pressure reducing valve. The acoustic boundary element model was set as a quadrilateral mesh unit with a mesh size of 1 mm, resulting in a total of 6035 nodes and 12,114 meshes. The generated boundary element grid is shown in Figure 6.

In order to study the propagation characteristics and distribution law of the radiated sound field outside the pressure reducing valve, the sound field analysis was carried out using LMS Virtual Lab, and the sound field at 1 m from the pressure reducing valve was selected as the research object. The steady-state flow field at 50% valve opening was input into the sound field as the initial condition. A sphere with a radius of 1 m was established around the pressure reducing valve, as shown in Figure 7. Four detection points are chosen on the top, bottom, left, and right sides of the sphere to analyze the frequency response and equivalent noise valve of the external radiation sound field of pressure reducing valve. The coordinate valves of each point are shown in

Table 3. Coordinates of pressure reducing valve far field noise monitoring points.

	X	Y	Z
Monitoring Point 1	1 m	0	0
Monitoring Point 2	0	1 m	0
Monitoring Point 3	0	0	1 m
Monitoring Point 4	0	0	−1 m

.

2.2. Experimental Platform

To verify the accuracy of the simulation results, this paper used the pressure reducing valve experimental platform shown in Figure 8 to conduct performance tests. The test bench was composed of a TU23 Chufan test bench, pressure reducing valve, valve seat, relief valve, and noise detector. The specific information on the experimental equipment is shown in

Table 4. Characteristics of test equipment.

Testing Equipment	Type	Function
Test bench	TU23 (Chufan, Beijing, China)	It includes components such as motors, pumps, fuel tanks, and sensors, providing basic conditions for pressure reducing valve testing
Pressure reducing valve	-	As a component to be tested
Valve seat	Customized processing	It is used to assemble test valves as well as test system valve blocks
Relief valve	-	Safety protection
Noise detector	DLY-2201 (DELIXI, Leqing, China)	Voice test

.

One of the original direct-acting pressure reducing valves was chosen, and the valve core, valve sleeve, and other parts were installed according to the assembly requirements of 0.008–0.012 mm clearance fit to ensure that the valve core and valve sleeve do not appear stuck. Each pressure reducing valve was tested three times under the set pressure using the Chinese national standard test [21], and the average value was taken. Before the test, the relief valve was connected to the outlet of the pressure reducing valve to ensure that the pressure at the outlet was adjustable. The TU23 Chufan test bench was used to adjust the inlet pressure to ensure stable pressure. The medium used in the test is 46 # hydraulic oil. The oil temperature was monitored by the temperature detection instrument of the test bench. The oil temperature was about 25 °C. The improved valve spool is shown in Figure 9.

A direct-acting pressure reducing valve was chosen, and the spool, valve sleeve, and other parts were assembled according to the assembly requirements to ensure that the spool and valve sleeve do not appear stuck. Each pressure reducing valve was tested three times under the set pressure, and the test results were averaged. A pressure reducing valve was installed on the test block. During testing, the pressure at the P port was increased from 0 to 100 bar, 150 bar, 200 bar, and 250 bar. By adjusting the pressure of the relief valve and rotating the setting screw, the outlet pressure of the valve was maintained at 25 bar. The noise detector was used to detect the noise. After the pressure reducing valve was stabilized, the noise within ten minutes after the detection was stabilized and measured every 30 s, and the average value was taken to obtain the noise under this pressure. The pressure of the system was raised slowly, and the noise of pressure reducing valve was measured under the set operating conditions.

2.3. Numerical Method

To determine the acoustic source, it is necessary to choose an appropriate noise numerical simulation method. The Lighthill acoustic analogy method provides a feasible alternative solution. This method solves an appropriate control equation for near-field flow, and an integral solution of the acoustic equation is obtained, thereby enabling noise prediction.

The acoustic module of Fluent software (2022R1) is based on the FW-H acoustic analogy model (Ffowcs Williams and Hawkings Acoustic Analogy Model). Based on the Lighthill theory, the FW-H acoustic analogy model further considers the influence of stationary and moving solid boundaries on the sound field. The specific form is as follows [22]:

$${\displaystyle \frac{1}{c^2}}{\displaystyle \frac{{\partial }^2{p}^{\prime }}{\partial t^2}}-{\displaystyle \frac{{\partial }^2{p}^{\prime }}{\partial x_i^2}}={\displaystyle \frac{\partial }{\partial t}}\left\{\left[{\rho }_0{v}_n+\rho (u_n-v_n)\right]\delta (f)\right\}+{\displaystyle \frac{{\partial }^2\left\{T_{ij}H(f)\right\}}{\partial x_i\partial x_j}}-{\displaystyle \frac{\partial }{\partial x_i}}\left\{\left[p_{ij}{n}_i+\rho u_i(u_n-v_n)\right]\delta (f)\right\}$$

(1)

In the formula, u_i is the component of fluid velocity in the direction of x_i; u_n is the normal component of fluid velocity on a solid surface; v_n is the component of the velocity of the solid surface in the normal direction of the surface; δ(f) is the Delta function; H(f) is the Heaviside function; P_ij = pδ_ij + μ(∂u_i/∂x_j + ∂u_j/∂x_i) − 2/3 × μ(∂u_k/∂x_k)δ_ij is the fluid compressive stress tensor; f = 0 represents a fixed boundary; and f > 0 denotes an external flow field.

The terms on the right side of Equation (1) correspond to the sound source terms of aerodynamic noise, and different terms correspond to different sound source categories and different sound generation mechanisms.

$${\displaystyle \frac{\partial }{\partial t}}\left\{\left[{\rho }_0{v}_n+\rho (u_n-v_n)\right]\delta (f)\right\}$$

corresponds to the monopole noise source, which is caused by the movement of the fluid caused by the movement of the solid boundary.

$${\displaystyle \frac{\partial }{\partial x_i}}\left\{\left[p_{ij}{n}_i+\rho u_i(u_n-v_n)\right]\delta (f)\right\})$$

corresponds to the dipole noise source, which is caused by the force acting on the fluid by the solid surface. The effect of the solid boundary is equivalent to the dipole sound source spread over the solid boundary, and the strength of the dipole sound source at each point is equal to the force acting on the fluid at that point on the solid surface.

$${\displaystyle \frac{{\partial }^2\left\{T_{ij}H(f)\right\}}{\partial x_i\partial x_j}}$$

(2)

corresponds to the quadrupole noise source, which is caused by the turbulent stress tensor of the wake and the shear layer.

The FW-H model in FLUENT software stipulates that for low-speed flow (M_a ≤ 0.3), the contribution of quadrupole sound sources is negligible compared to monopole and dipole sound sources. This means that for the stationary solid boundary, FLUENT software only considers the role of the dipole sound source, that is, the noise radiated by the pulse force on the solid surface.

3. Results and Discussion

In this section, the numerical simulation of the flow field and sound field of the direct-acting pressure reducing valve has been completed. The internal pressure, flow, streamline, and internal sound field of the pressure reducing valve were analyzed, and the main position of the noise of the pressure reducing valve was obtained. Secondly, the pressure monitoring point was set in the flow channel of the pressure reducing valve as the main noise position, the sound power level distribution of the central plane of the pressure reducing valve was obtained, and the main sound source area was defined. Based on this, an improvement scheme was proposed, and the sound power level changes under different forms of spool valve sleeve combinations were analyzed.

3.1. Internal Flow Field Analysis

3.1.1. Internal Flow Velocity Analysis

When the pressure of the oil inlet of the pressure reducing valve was 100 bar, the flow velocity cloud atlas of the center plane of the pressure reducing valve was obtained as shown in Figure 10a. The figure shows the flow field cloud image under steady-state simulation. Flow rate was selected as the main indicator of this section. The fluid domain of pressure reducing valve was divided into four parts: the oil inlet channel, the cross-section of the passage, the valve core channel, and the oil outlet channel. The fluid flow velocity at the cross-section of the passage reached 108 m/s, and the maximum fluid flow velocity at the oil inlet of the lower valve spool was 136.9 m/s. After the fluid passed through the valve spool passage, a portion of it encountered the inner wall of the valve seat, and a vortex zone was thus formed. When the pressure conditions at the oil inlet of the pressure reducing valve were 100 bar, 150 bar, 200 bar, and 250 bar, it can be concluded that as the oil inlet pressure increased, the flow velocity in all parts of the pressure reducing valve generally became higher. The velocity clouds and velocity curves at different points of the center plane of pressure reducing valve are shown in Figure 10 and Figure 11 The dark blue area in the cloud image did not have an oil velocity of 0 m/s in real conditions, which was only caused by the simulation under ideal conditions.

3.1.2. Internal Pressure Analysis

When the pressure conditions of the oil inlet of pressure reducing valve were 100 bar, the pressure cloud map at the center plane of the pressure reducing valve was obtained as shown in Figure 12a. When the fluid entered the annular region between the valve spool and the high-pressure cavity valve casing through the flow section, the fluid pressure decreased due to the reduction in the flow area. After the fluid passed through it, a hydraulic impact on the valve spool was generated, and a local high pressure was formed. When the fluid passed through the lower oil inlet on the valve spool, the pressure decreased rapidly as the flow velocity increased sharply. When the pressure conditions at the oil inlet of the pressure reducing valve were 150 bar, 200 bar, and 250 bar, it can be seen that the trend is basically the same as shown in Figure 12. Meanwhile, the pressure of the three points (marked in Figure 12a) under different boundary conditions is shown in Figure 13.

3.1.3. Internal Streamline Analysis

When the pressure of the oil inlet of the pressure reducing valve was 100 bar, the flow line cloud atlas at the center plane of pressure reducing valve was obtained, as shown in Figure 14a. As the flow velocity increased, a large amount of eddy currents was generated when the fluid passed through the cross-section of the passage of the valve spool, the lower wall of the oil inlet, and the position near the valve wall. When the pressure conditions of the oil inlet of pressure reducing valve were 150 bar, 200 bar, and 250 bar, the internal flow line cloud atlas of the pressure reducing valve is shown in Figure 14. It can be concluded that the change tendency in the internal flow of the pressure reducing valve is consistent under various inlet pressures. But as the oil inlet pressure increased, the fluid flow velocity rose, and more eddy currents were generated inside the pressure reducing valve, thus leading to an increase in noise.

3.1.4. Internal Sound Field Noise Distribution

When the pressure of the oil inlet of the pressure reducing valve was 100 bar, the sound power level distribution cloud atlas at the center plane of pressure reducing valve was obtained, as shown in Figure 15a. In Figure 15a, three points are marked. Under different boundary conditions, the sound power level of the three points is shown in Figure 16. From Figure 15, it can be seen that when the fluid passes through the cross-section of the passage on the valve spool and enters the annular region between the valve spool and the high-pressure cavity valve casing, the sound power level reaches a maximum of 132 dB. The sound power level distribution of the pressure reducing valve oil inlet under pressure conditions of 100 bar, 150 bar, 200 bar, and 250 bar is shown in Figure 15. It can be seen that the change tendency of the internal sound power level of the pressure reducing valve is consistent under different import pressures. However, as the oil inlet pressure increased, the sound power level also rose. When the oil inlet pressure was 250 bar, the maximum sound power level of the sound field reached 153 dB. The main sound source regions were located at the cross-section of the passage, the flow channel of the valve core, and the oil outlet channel.

3.2. Noise Calculation of External Sound Field

When the oil inlet pressure was 100 bar, a noise simulation of the external sound field of pressure reducing valve was conducted, and the noise spectrum diagram of the four monitoring points of the pressure reducing valve was obtained, as shown in Figure 17.

The spectrum change trend of the four monitoring points of the pressure reducing valve was basically similar. With the increase in frequency, the noise sound pressure level decreased. When the frequency was low, the acoustic pressure amplitude of each monitoring point was high. Between 1000 Hz and 4000 Hz, the acoustic pressure amplitude of each monitoring point showed a downward trend.

In order to study the propagation characteristics of noise more clearly, the noise pressure level information of each monitoring point was analyzed, as shown in

Table 5. Sound pressure level of each monitoring point.

	Noise (dB)
Monitoring Point 1	85.1
Monitoring Point 2	86.1
Monitoring Point 3	84.6
Monitoring Point 4	84.1

.

According to the noise valve of the sound pressure level at each monitoring point in the table, it can be seen that under this operating condition, the effective noise valve was between 84 dB and 86 dB, and the propagation of noise was related to the location of the monitoring points. At the Y-axis, which was the oil inlet and outlet direction of pressure reducing valve, the effective valve of noise was relatively higher.

3.3. Oil Orifice Analysis

After analyzing the reasons for noise generation, improvements were made to the oil orifice on the valve spool, as shown in Figure 18. There were oil orifices with different positions, quantities, and diameters on both sides of the valve spool. Eleven combinations of different forms of the valve spool were set up, and noise simulation analysis was performed. The parameters of the 11 groups of different structures are shown in

Table 6. Structure modification parameters.

Group	Position	Quantity	Diameter (mm)
1	U	2	3
	L	2	3
2	U	2	5
	L	--	--
3	U	2	4
	L	--	--
4	U	2	2
	L	2	2
5	U	2	1.5
	L	2	1.5
6	U	3	2
	L	3	2
7	U	3	1.5
	L	3	1.5
8	U	4	2
	L	4	2
9	U	4	1.5
	L	4	1.5
10	U	8	1.5
	L	--	--
11	U	4	2
	L	2	3

, where U is the oil orifice on the upper side of the valve spool and L is the oil orifice on the lower side of the valve spool.

The noise simulation results of the above 11 groups are shown in Figure 19.

According to the simulation results, the flow noise generated by the eighth group was smaller than that of other structures, with the improved valve noise reduced by 5.2 dB compared to the original valve. The structural combination of Group 8 consisted of four oil orifices with a diameter of 2 mm on the upper and lower sides of the valve spool. The average valve of the original valve simulation was compared with the improved valve simulation.

Figure 20 shows the plan-A flow velocity cloud atlas of the direct-acting pressure reducing valve under an oil inlet pressure of 100 bar after modification. From the figure, it can be seen that due to the change in the structure of the oil orifice at the valve spool, the flow velocity at the cross-section of the passage and in the valve spool channel decreased, with the maximum flow velocity decreasing by 6.8%.

Figure 21 shows the plan-A flow line cloud atlas of the direct-acting pressure reducing valve under an oil inlet pressure of 100 bar after modification. From the figure, it can be seen that the modified structure significantly reduced eddy current formation at the oil orifice of the valve spool, leading to smoother streamlines and an overall improved flow field.

Figure 22 shows the plan-A sound power level distribution cloud atlas of the direct-acting pressure reducing valve under an oil inlet pressure of 100 bar after modification. From the figure, it can be seen that the noise distribution of the pressure reducing valve was closely related to the flow of the fluid through the valve spool. The locations with higher sound power levels appeared at the cross-section of the passage and the oil inlet of the valve spool. Inside the valve spool, due to the improved flow field of the pressure reducing valve, the eddy current and the pressure pulsation were reduced, resulting in a significant reduction in the sound power level by approximately 6 dB.

4. Experimental Validation

Table 7. Comparison of oil velocity simulation and experimental data.

Test	Simulation	Experiment	Error Rate
Oil inlet oil velocity (m/s)	100.8	99.5	1.28%
Oil outlet oil velocity (m/s)	63.0	60.1	4.8%

shows the oil flow rate of the inlet and outlet of the pressure reducing valve measured in this test. It was mainly tested under an inlet pressure of 150 bar and an outlet pressure of 25 bar. It can be seen from the table that the oil flow rate of the inlet port of the valve was 99.5 m/s, and the oil flow rate of the outlet port was 60.1 m/s, which is not much different from the simulation results.

Table 8. Original valve experiment data.

Test Number	P-Port Pressure (bar)	A-Port Pressure (bar)	Noise (dB)	Confidence Interval (dB)
1	100	25	92.5	[90.1–94.9]
2	150	25	103.6	[99.6–107.6]
3	200	25	105.5	[101.5–109.5]
4	250	25	107.5	[104.7–110.3]

shows the noise level of the original valve measured in this experiment, ranging from 92 to 108 dB. The average level of simulated noise was 75 dB. Due to the fact that the experiment measured all noise superposition of the valve, while this paper only studied flow noise and ignored vibration noise, it results in the fact that the test noise results are bigger than the simulated noise. At the same time, the confidence interval of the measured pressure under each condition after optimization is also given in Table 8.

Table 9. Improved valve experiment data.

Test Number	P-Port Pressure (bar)	A-Port Pressure (bar)	Noise (dB)	Confidence Interval (dB)
1	100	25	85.2	[81.4–89.0]
2	150	25	91.3	[86.3–96.3]
3	200	25	93.7	[90.7–96.7]
4	250	25	94.8	[91.8–97.8]

shows the noise of the improved valve measured in this experiment, ranging from 85 to 95 dB. These results indicate that the flow noise within the pressure reducing valve could be suppressed by this structure. When the oil inlet pressure was 100 bar, the noise of the improved pressure reducing valve was reduced by about 6 dB compared with the original one, which was consistent with the noise-reduced valves obtained from the simulations. Therefore, it can be concluded that this structure can effectively improve the turbulence inside the pressure reducing valve and reduce flow-induced noise. At the same time, the confidence interval of the measured pressure under each condition after optimization is also given in Table 9.

The comparison of noise levels before and after improvement is shown in Figure 23.

In this test environment, noise factors mainly included human voice, valve flow-induced noise, vibration noise, test bench equipment noise, and other noise. However, the test was carried out in a relatively quiet environment, and the operator remained silent during each test, thus ensuring that the noise of the human voice and the test bench equipment were controlled within a stable and acceptable range. Based on this premise, the effect of optimization measures can be clearly reflected in the collected data results. It can be seen from the figure that as the oil inlet pressure of the pressure reducing valve increased, the noise of it decreased significantly. When the oil inlet pressure ranged from 50 bar to 100 bar, the noise level was reduced by about 6 dB. When the oil inlet pressure ranged from 150 bar to 250 bar, the noise level was reduced by about 12 dB.

5. Conclusions

This article employed CFD software (2022 R1) to study the direct-acting pressure reducing valve, determine the influence of structural parameters of pressure reducing valve on flow-induced noise, and introduce an improved structure for suppressing the noise of the pressure reducing valve. This study provided a reference for the design of pressure reducing valves. The main conclusions of this study are as follows:

(1)

The internal flow field and sound field of the pressure reducing valve have been simulated and analyzed, and the main position of the flow-induced noise of the pressure reducing valve was obtained. The pressure monitoring points were set in the flow channel of the main noise position, and the sound pressure spectrum of the four monitoring points was obtained. Based on this, an improvement scheme was proposed.

(2)

The structure of two rows of eight holes with a diameter of 2 mm was adopted. The noise valve at each monitoring point of the improved pressure reducing valve was 69~70 dB, which was less than 74~76 dB of the original valve. The noise of the improved valve was 6 dB lower than that of the original valve, and the effective sound pressure level reduction was 8%. Its precision was verified by experiments.

The following is an explanation of the limitations and potential future developments for this method:

(1)

This paper analyzed the location and causes of the noise generated by the direct-acting pressure reducing valve, but the deep-seated causes of the noise need to be further studied.

(2)

This paper only analyzed the flow-induced noise of the pressure reducing valve, and further research is needed on the vibration mechanism of the pressure reducing valve and the noise caused by its mechanical vibration.

The pressure reducing valve is widely used in the industrial environment. Its noise level is not only directly related to the health and work efficiency of the operator but also related to the optimization of equipment performance, maintenance cost, and the overall production environment. The future research directions of cartridge valves studied in this paper include but are not limited to the following: Intelligent noise reduction technology that uses sensors, controllers, and other intelligent devices to monitor the noise level in real time and automatically adjust the operating conditions or valve body structure according to feedback to achieve dynamic noise reduction. Environmentally friendly materials and processes to develop new materials and processes with better environmental performance and noise reduction effects to meet increasingly stringent environmental requirements and industrial needs. Interdisciplinary application that combines the knowledge of acoustics, fluid mechanics, material science, and other disciplines; more in-depth research needs to be carried out to explore more efficient noise reduction methods.