Influence of Distributor Structure on Through-Sea Valve Vibration Characteristics and Improvement Design

Abstract

To address the issue of excessive transient noise during the opening and closing of a sea valve, a method for reducing pressure fluctuations during the opening of the electromagnetic hydraulic distributor has been proposed by analyzing the structure and working principle of the distributor. Based on theoretical calculation and simulation analysis, the size and shape of the buffer slot of the flow hole are determined under the condition that the stable working flow rate remains unchanged. An improved electromagnetic hydraulic distributor is developed and tested. The results indicate that this method can effectively control the opening and closing transient noise of the sea valve.

1. Introduction

The through-sea valve is a critical component for submarines to interface with the external environment for water intake and release, which is essential during diving, surfacing, and cooling processes. The transient noise generated by the through-sea valve significantly impacts its stealth capabilities. Scholars both domestically and internationally have conducted a series of meaningful studies on the transient noise control of through-sea valves. Oshima and Tsnneo [1,2,3,4] initiated experimental studies on the fluid flow within the valve using a half-cone valve model, measuring the pressure distribution between the valve seat and the valve core surface. Additionally, they studied the impact of changes in valve geometry and different cone angles of the valve core on fluid flow characteristics. Johnston [5] experimentally investigated the flow within lift valves and disk valves under constant flow conditions with Reynolds numbers between 2500 and 35,000, providing a formula for fluid forces. Concurrently, Ito and Takashashi [6] simulated a valve with a conical sealing surface, deriving the flow field characteristics of such valves. Ueno [7] numerically simulated and acoustically tested the flow fields within several differently shaped valve bodies, concluding that the shape of the valve sleeve and core significantly affects fluid characteristics within the valve. They also demonstrated that using a two-stage throttling reduces the average pressure and pressure fluctuations of the fluid within the valve.

With the advancement of computer technology, CFD(computational fluid dynamics) method is extensively utilized in the analysis and calculation of valve flow fields. Palau Salvador [8] used Flunet 6.1 software for three-dimensional flow analysis within complex flow channels of globe valves, achieving simulation results consistent with experimental data. Li [9] studied the opening torque of check valves, proving that the magnitude of the opening torque is related to fluid direction, opening speed, valve disk installation angle, and the angular velocity of the valve disk opening. Amirante [10] simulated and analyzed the cavitation phenomenon at the orifice of butterfly valves. Masjedian [11] studied the cavitation phenomenon of globe valves, performing a Fourier transform on the acoustic vibration caused by cavitation within the valve, and analyzing amplitude, energy, frequency, etc. Baran [12] established a valve cavitation test device, analyzing the pressure drop, noise, and vibration issues caused by cavitation when the valve is in the cavitation phenomenon, and designed a controller based on the fuzzy algorithm to determine the cavitation flow velocity pressure of the valve. Ileana Vladu [13] conducted research on magnetorheological globe valves, providing a method for determining the wave front mathematical function, and experimented with valves with different parameter sets, determining the qualitative impact of these parameters on magnetorheological globe valves. Bhowmik [14] used experimental methods to study the flow characteristics of globe valves in the flow control and conversion system of supercritical gas turbines, plot flow characteristic curves, and calculate the ratio of measured mass flow rate to theoretical mass flow rate. Liu [15] used dynamic mesh and CFD technology to analyze and calculate the flow of liquid during the opening and closing process of the through-sea valve, providing a theoretical basis for the optimization of the internal flow channel of the through-sea valve. Cui [16] used Fluent 6.3 software for numerical simulation calculations of the optimized flow field scheme of the through-sea valve, concluding that changing the outlet flow channel structure and appropriately increasing the internal cavity of the valve body is beneficial to improving the flow capacity of the through-sea valve. Cui [17] conducted numerical simulation analysis of the three-dimensional flow of a certain ship’s through-sea valve based on the RANS method and standard turbulence model, obtaining the influence rules of key shapes and sizes of the right-angle cut-off valve on flow characteristics, and proposed an optimization design scheme for the internal flow channel of such valves. Liu [18,19], based on CFD principles, proposed a new flow channel structure, and by comparing the flow field characteristics of the through-sea valve at different installation angles of the internal support, they provided the installation method of the internal support that minimizes the turbulent kinetic energy from the perspective of preventing cavitation and reducing vortex noise. Zeng [20] processed the flow lines at the connection between the inner port and the main body, the corner, and the bulge of the titanium alloy through-sea valve, making the movement of seawater inside the valve more stable, and effectively reducing the turbulence noise during the operation of the through-sea valve. Sotoodeh [21] presented a short review of valve stem sealings and their importance, reviewing various types of sealing as well as materials including O-rings, graphite packing, thermoplastic PTFE or lip seal, Vee pack, and metallic seals. Parameters that affect the valve stem sealing such as temperature, sealing material, number of rings, and gland force were mentioned. Sotoodeh [22] examined a case study of breakaway torque (Break to Open) calculation and actuator sizing for a bare stem full-bore ball valve in pressure Class 300 and 22Cr duplex body material.

The aforementioned research on through-sea valves mainly focuses on the valve body structure and internal flow channel, with a lack of research and improvement on the through-sea valve drive system. The noise caused by the hydraulic drive mechanism is also noteworthy. This paper conducts improvement research on the control system of the through-sea valve from the perspective of the working principle of the through-sea valve opening and closing and the hydraulic drive system. The valve sleeve structure of the electro-hydraulic distributor is optimized and designed to suppress the hydraulic impact caused by rapid changes in flow, achieving effective control of the transient noise during the opening and closing of the through-sea valve.

The remainder of this paper is organized as follows. In Section 2, the structure and working principle of the electromagnetic hydraulic distributor for the sea valve are introduced. In Section 3, the impact of the electromagnetic hydraulic distributor on the transient noise of the sea valve opening and closing is analyzed using theoretical calculations and simulation methods. A prototype of the improved electromagnetic hydraulic distributor is designed, and vibration tests for the opening and closing of the sea valve are conducted in Section 4. Finally, some conclusions are drawn in Section 5.

2. The Working Principle of the Through-Sea Valve Distributor

The schematic diagram of the hydraulic drive system for the through-sea valve is shown in Figure 1. In this system, the electromagnetic hydraulic distributor plays a role in controlling the start, stop, or change in motion direction of the through-sea valve and other actuator components by using the change in the position of the valve core relative to the valve body to connect, cut off, or change the flow direction of the working medium. When the distributor switches the fluid flow by driving the pilot valve with the electromagnetic coil, a transient impact is caused due to the strong liquid pressure. To address the impact issue caused by the rapid switching of the electromagnetic hydraulic distributor, the design of the valve sleeve of the main valve core is optimized. By maintaining the stable working flow rate unchanged, the switching speed of the main valve core is slowed down by altering the flow area of the valve sleeve, thereby extending the switching time and reducing the hydraulic impact.

The structural principle of the electromagnetic hydraulic distributor is shown in Figure 2. Essentially, the electromagnetic hydraulic distributor is a pilot-operated electro-hydraulic valve, which consists of an electromagnet, a pilot valve, and a main valve. The distributor operates in two control modes: one is electrical control, which achieves the switching of the main oil route by energizing or de-energizing the electromagnet to drive the pilot valve core; the other is direct control, which manipulates the position change in the pilot valve core using a handle to achieve the switching of the main oil route.

The operational dynamics of the through-sea valve distributor are scrutinized within this discourse. A comprehensive analysis is conducted to elucidate the influence of the electromagnetic hydraulic distributor on the acoustic transients occurring during the opening and closing phases of the sea valve. This examination is executed through a synergistic application of theoretical calculations and advanced simulation methodologies, ensuring a robust and empirically grounded assessment. In the neutral position of the distributor, both the left and right electromagnets are inactive, aligning the handle and the pilot valve core to their central positions, under the influence of a spring mechanism. Concurrently, the main valve’s control cavities on either side are linked to the discharge port via the pilot valve, with the main valve core also positioned centrally due to spring tension. At this juncture, the main valve’s pressure oil port is sealed, while the opening oil port, closing oil port, and discharge oil port are interconnected.

When transitioning the distributor to the open position, the left electromagnet is activated, or the handle is manually adjusted. This action, propelled by electromagnetic force or manual operation, prompts the pilot valve core to shift to the right. The pressure oil port then feeds the right control cavity of the main valve through a throttle, while the left control cavity is connected to the discharge oil port via the pilot valve. Consequently, the main valve core moves leftward under hydraulic pressure, establishing a connection between the opening oil port and the pressure oil port, with the closing oil port simultaneously linking to the discharge oil port. If the distributor is shifted to the closed position, the right electromagnet is activated, or the handle is adjusted accordingly. The pilot valve core responds by moving leftward, due to the electromagnetic force or manual manipulation. This motion directs the pressure oil port to supply the left control cavity of the main valve through a throttle, while the right control cavity is connected to the discharge oil port via the pilot valve. The main valve core subsequently moves rightward under hydraulic pressure, aligning the closing oil port with the pressure oil port, and simultaneously connecting the opening oil port to the discharge oil port.

A thorough analysis of the pressure fluctuations at each oil port during the opening and closing process of the through-sea valve is undertaken. Subsequently, simulation calculations are performed to deduce the origins of hydraulic impact. The configuration of the oil ports is delineated in Figure 3.

Prior to the actuation of the electromagnetic valve distributor, the pressure at port P is equivalent to the oil source pressure, with ports A and B connected to port T via the pilot valve, resulting in negligible pressure at ports A, B, and T. Upon receipt of a control signal, the electromagnetic hydraulic distributor initiates the pilot valve core movement in the direction dictated by the control signal, facilitated by the electromagnetic force or manual operation. Concurrently, the main valve opens to its maximum aperture according to its inherent dynamic characteristics. As the P port and A port are initially connected—assuming the distributor opens in this orientation—the pressure at port A commences its ascent. The oil influx from port P to port A is entirely allocated to compensate for the volume of hydraulic oil compressed by the increasing pressure at port A. During this phase, ports B and T remain interconnected. The pressure differential between ports A and B exerts a force on the piston of the through-sea valve actuator hydraulic machine. Once this force surpasses the frictional resistance of the through-sea valve actuator hydraulic machine and its transmission mechanism, the valve disk of the through-sea valve begins to accelerate.

As the opening of the electromagnetic hydraulic distributor enlarges, the flow area of port A increases, bolstering the flow into port A and perpetuating the rise in pressure at port A. Consequently, the through-sea valve actuator hydraulic machine continues to accelerate. When the hydraulic machine’s velocity reaches a threshold value, the pressure loss within the pipelines and valve components of the hydraulic drive system escalates in tandem with the system flow. This progression culminates in a state where the residual pressure exerted on the cylinder after accounting for the pressure loss across the system is precisely equal to the friction force, achieving a state of equilibrium. At this juncture, the through-sea valve actuator hydraulic machine maintains a uniform motion, with the pressures at each oil port remaining essentially unchanged. When the through-sea valve actuator piston reaches a steady state of motion and the control signal remains constant, the opening of the electromagnetic valve distributor remains static, and the through-sea valve actuator hydraulic machine sustains uniform motion, with the pressures at each oil port remaining essentially unchanged. Upon full opening or closing of the through-sea valve, the actuator hydraulic machine ceases its movement, the system control signal is disconnected, the electromagnetic hydraulic distributor resets, and the two chambers of the through-sea valve actuator hydraulic machine are fully sealed.

In light of the aforementioned analysis, pressure fluctuations at the valve port occur during the opening and closing process of the through-sea valve. Prolonging the response time of the electromagnetic hydraulic distributor, effectively decelerating the response velocity of the electromagnetic valve, can mitigate the pressure fluctuations during the valve’s operation. Furthermore, optimizing the design of the distributor’s valve sleeve structure can alter the output flow characteristics of the distributor, which is instrumental in enhancing its on–off characteristics, thereby effectively reducing hydraulic impact.

3. Theoretical Calculation and Simulation Analysis

In this section, an in-depth investigation of the dynamic behavior of the through-sea valve distributor is conducted. The influence of the electromagnetic hydraulic distributor on the transient acoustic emissions associated with the actuation of the sea valve is meticulously examined through a combination of theoretical computations and computational simulation techniques.

3.1. Theoretical Calculation

Hydraulic shock occurs during the opening and closing process of the distributor. When hydraulic shock occurs, a transient fluid force is generated, which can be calculated as follows [23]:

$$F_{bt}=L{C}_{d}w\sqrt{2\rho \Delta p}{\displaystyle \frac{d{x}_v}{dt}}$$

(1)

where $L$ is the length from the center of the spool inlet to the center of the return oil outlet, $C_d$ is the flow coefficient of the valve opening, $w$ is the area gradient around the valve opening, $\rho$ is the density of the hydraulic oil flowing through the spool, and $x_v$ is the valve opening.

From Equation (1), it can be seen that the transient hydraulic force experienced by the valve core is proportional to both the moving speed of the valve core and is also proportional to the damping length of the spool. Its direction is always opposite to the acceleration of the fluid flow inside the valve. For the cylindrical spool structure, when the displacement of the main valve core is zero, the pressure in the left and right control chambers of the main valve is equal; the P port is the oil source pressure value, and ports A, B, and T are connected. The throttling slot on the main valve core is a full circumferential opening, so the flow curve of the spool depends on the shape of the flow hole in the valve sleeve. The main valve sleeve of the distributor’s electromagnetic valve has eight circumferentially arranged flow holes, each with a diameter of 3 mm. According to the form of the electromagnetic valve opening, the valve opening can be regarded as a series of eight flow areas composed of thin-walled holes in parallel. The parallel valve openings can be approximated as having consistent pressure differences before and after each small flow hole. The total flow through the valve opening is the sum of the flows through each hole. Therefore, the equivalent area of the distributor’s valve opening can be derived as follows:

$$A_d=8A_{d1}$$

(2)

where $A_d$ is the equivalent area of the valve opening, and $A_{d1}$ is the equivalent flow area of a single flow hole. During the valve opening and closing process, the flow rate through the valve opening is calculated according to the thin-edge orifice flow formula as follows:

$$q_v=C_{d}A\sqrt{2\Delta p/\rho }$$

(3)

where $q_v$ is the total flow rate through the valve opening, $\Delta p$ is the pressure difference across the valve opening, $\rho$ is the oil density, and $A$ is the area function related to the valve opening.

The equivalent throttling area of a single flow hole can be considered as consisting of two areas in series as shown in Figure 4. The calculation formula for the equivalent throttling area in series is the following:

$$A_d={\displaystyle \frac{1}{\sqrt{\frac{1}{A_1^2}+\frac{1}{A_2^2}}}}$$

(4)

where $A_1$ is the projected area of the valve sleeve flow hole on the valve core, and $A_2$ is a rectangular section.

The calculation diagram is shown in Figure 5, where x represents the valve opening, and the first throttling area at the valve opening of x is represented by the area of the segment in the diagram, where R represents the radius of the flow hole.

By integrating the segment area as a function of the valve opening and integrating from a valve opening of 0 to a valve opening of x, we obtain the following:

$$A_1={\displaystyle \underset{0}{\overset{x}{\int }}2\sqrt{R^2-{(x-R)}^2}dx}$$

(5)

Substituting R = 1.5 mm into the expression yields the integral result as follows:

$$A_1={\displaystyle \frac{9}{4}}\arcsin ({\displaystyle \frac{2x}{3}}-1)+(x-{\displaystyle \frac{3}{2}})\sqrt{{\displaystyle \frac{9}{4}}-{(x-{\displaystyle \frac{3}{2}})}^2}+C$$

(6)

where C represents any constant.

Substituting $x=0$ and $A_1=0$ into Equation (6) and solving for the constant $C={\displaystyle \frac{9}{8}}\pi$, the total throttling area of the distributor’s valve opening is as follows:

$$A_1=18\arcsin ({\displaystyle \frac{2}{3}}x-1)+9\pi +(8x-12)\sqrt{{\displaystyle \frac{9}{4}}-{(x-{\displaystyle \frac{3}{2}})}^2}$$

(7)

The second throttling area is a rectangular section, calculated as shown in Equation (8):

$$A_2=2bW$$

(8)

where b represents the valve core gap with the valve sleeve, and in the distributor, the value is 1 mm. Substituting the data, when the valve core moves to the center of the flow hole and maintains the maximum value of 3 mm², then the second equivalent throttling total area of the distributor’s valve opening is as follows:

$$A_2=\left\{\begin{array}{l}16\sqrt{{1.5}^2-{(1.5-x)}^2},0\le x\le 1.5\\ 24,1.5<x\le 3\end{array}\right.$$

(9)

According to Equations (4), (7), and (9), the valve opening flow area curve can be obtained, as shown in Figure 6. The maximum opening of the valve opening is 3 mm, and the equivalent throttling area of the distributor’s valve opening is approximately 22.09 mm². The gradient of the flow area during the opening process of the valve opening is relatively large.

To reduce the area gradient, a slot shape with a smaller area gradient than the circular hole must be chosen. A triangular buffer slot is chosen at the opening direction of the flow hole, as shown in Figure 7.

To achieve a lower area gradient, L is set to 5 mm, and h is set to 1 mm. At this point, the flow rate calculation formula for the valve opening still satisfies Equation (3). For the modified flow hole, the maximum valve opening has changed, and consequently, the maximum flow area has also changed. The overlapping area between the triangular slot and the circular hole is relatively small, so the calculation of the initial flow area is based on this. When the valve opening is from 0 to 2 mm, the flow area is calculated using the area of a triangle. When the valve opening is from 3 mm to 5 mm, it is still considered a circular hole for calculation purposes. The maximum valve opening has been adjusted from 3 mm to 5 mm, and the initial flow area has increased from 7.07 mm² to 8.07 mm². On the other hand, to ensure that the stable working flow rate of the distributor remains unchanged, the maximum opening of the flow area should be minimized. Here, only one flow hole is modified, while the other seven holes remain unchanged. The three-dimensional model of the buffer slot is shown in Figure 8.

For the flow hole with a buffer slot, the expression for the first flow area varies in two segments and can be approximated as follows:

$$A_{g1}=\left\{\begin{array}{l}0.25x^2,0\le x\le 2\\ 18\arcsin ({\displaystyle \frac{2(x-2)}{3}}-1)+9\pi +1+\\ (8x-28)\sqrt{{\displaystyle \frac{9}{4}}-{(x-2-{\displaystyle \frac{3}{2}})}^2},2<x\le 5\end{array}\right.$$

(10)

The second flow area can be expressed as follows:

$$A_{g2}=\left\{\begin{array}{l}0.5x,0\le x\le 2\\ 16\sqrt{{1.5}^2-{(3.5-x)}^2}+1,2<x\le 3.5\\ 24,3.5<x\le 5\end{array}\right.$$

(11)

Then, according to Equations (4), (10), and (11), the improved valve opening flow area curve is obtained, as shown in Figure 9.

After the distributor is improved, the maximum valve opening is 5 mm, and the maximum throttling area is approximately 22.15 mm², which is about 0.06 mm² larger than the unimproved area of 22.09 mm². The equivalent throttling area remains almost unchanged. According to Equation (3), it can be deduced that the stable working flow rate of the improved distributor remains almost unchanged. However, after the flow hole is slotted, the area gradient of the valve opening is modified, initially providing a smaller area gradient at the slot, and then gradually increasing, so the output flow is also a gradual process, thereby enhancing the on–off characteristics of the distributor. On the other hand, after the flow hole is slotted, the speed of the valve core during the distributor’s opening process is reduced, and the time for the distributor to reach the maximum opening is extended, which means the time for the distributor to reach a stable working flow rate is extended, thereby suppressing the hydraulic shock caused by the sudden change in flow rate. The flow rate rise curves before and after the improvement are shown in Figure 10.

3.2. Simulation Analysis

To study the impact of the electromagnetic hydraulic distributor’s opening time on the opening and closing process of the through-sea valve, a simulation model of the valve-controlled through-sea valve actuator cylinder is established based on the AMESim 2021.1 software, which enables system engineering modeling and simulation across multiple disciplines on a unified platform, including mechanical, hydraulic, pneumatic, thermal, electrical, and magnetic physical domains, as shown in Figure 11.

According to Equation (4), let the hydraulic oil density be $\rho =115$ kg/m³, the flow coefficient $C_d=0.61$, and the equivalent throttling area of the valve opening calculated above is 22.09 mm². Let the working flow rate of the distributor be 20 L/min, then the pressure drop across the valve opening is calculated to be 0.26 MPa. When the valve opening is fully open, the flow area between port P and port B is 22.04 mm². The simulation value is close to the calculated value of 22.09 mm². At this time, the velocity–time curve of the cylinder is shown in Figure 12.

When the response time is 0.1 s, the peak velocity of the cylinder during the opening process of the distributor is 0.06 m/s, which is twice the velocity value when the cylinder moves smoothly, and four oscillations occur during stabilization. When the response time is 0.2 s, the peak velocity of the cylinder is 0.05 m/s, and the oscillations decrease to two during stabilization, with both the amplitude and the number of oscillations decreasing. When the response time is 0.3 s, the degree of velocity fluctuation is similar. The longer the response time, the smaller the speed change when the cylinder starts, and the slower the speed decrease when the cylinder stops. That is, extending the response time of the distributor can effectively reduce the amplitude and number of velocity oscillations during the cylinder’s starting process.

The pressure change curve at port A during the opening process of the distributor is shown in Figure 13. The pressure at port A oscillates multiple times during the opening process, transitioning towards the steady-state pressure. When the response time is long, the oscillation amplitude is small and the number of oscillations is fewer. When the response time is 0.1 s, the oscillation amplitude is nearly ten times the steady-state value. During the closing of the distributor, the pressure at port A overshoots. When the response time is less than 0.3 s, the pressure has a sharp peak with a large amplitude; when the response time is greater than 0.3 s, the pressure change is more gradual.

In summary, through theoretical calculation and simulation analysis, it is known that slotting the flow hole of the main valve sleeve of the electromagnetic hydraulic distributor can extend the distributor’s response time and improve the output flow characteristics. The simulation results show that for this type of distributor, controlling the response time around 0.3 s can significantly reduce the pressure fluctuations at each valve port. Therefore, by extending the response time of the electro-hydraulic distributor, the pressure oscillations during the valve’s opening and closing processes can be mitigated, thereby reducing the transient noise.

4. Experimental Study

To verify the control effect of the improved hydraulic distributor on the transient noise of the through-sea valve during its opening and closing, a triangular slot is set at the flow hole of the main valve sleeve of the distributor according to the above simulation analysis results. An improved prototype of the electromagnetic hydraulic distributor is designed and tested.

4.1. Prototype Design and Test Rig Construction

The improved valve sleeve is shown in Figure 14a. A through-sea valve opening and closing noise experimental rig is constructed, and the rig is shown in Figure 14b. The hydraulic actuator (Hangzhou Zhijiang Hydraulic and Electrical Machinery Co., Ltd., Hangzhou, China, TY500/210-50) is directly mounted on the sea valve, and the sea valve, connected with the hydraulic actuator, is fixed to the test rig. The opening and closing of the sea valve are controlled by an electromagnetic distributor. The data acquisition system utilizes acoustic and vibration products from HBK Company, London, United Kingdom, which mainly includes a data acquisition chassis LAN-X1-3660-C-100, an acquisition module 3050, and two triaxial accelerometers 4507-B-004. The two accelerometers are positioned on the base of the sea valve to simultaneously measure the vibration acceleration levels in three directions. Simultaneously, the experimental procedures are conducted in reference to the relevant industrial standards [24,25] for noise evaluation.

After the test system stabilizes the working pressure, the values of vibration acceleration levels are measured to be around 83 dB. During the experiment, the transient time-domain graph of two measurement points is obtained (analysis frequency range: 10 Hz to 10 kHz). The statistical method for vibration acceleration levels is as follows: first, the time-domain vibration signals during the operation of the sea valve are acquired. Then, the effective period of the vibration test signals is selected. Finally, the vibration peak values at two measurement points on the base of the sea valve are read, and the average of these two peak values is processed to obtain the final vibration acceleration level. Moreover, in the test data, x represents the direction parallel to the rig and pointing towards the center of the valve seat circle, y represents the direction parallel to the rig and perpendicular to x, and z represents the direction perpendicular to the rig and upward.

To ensure the accuracy and reliability of the tests, it is necessary to avoid the following sources of error: (a) ensuring that the precision specifications of the triaxial accelerometers and data acquisition systems meet the test requirements; (b) maintaining consistent installation positions for the accelerometers under various conditions, and securing the accelerometers firmly with strong adhesive; (c) recording the background noise level of the test environment before testing, and subtracting it from the measurement results during data processing; and (d) to ensure the reliability and repeatability of the results, each test condition should be repeated at least three times.

4.2. Experiment Results and Analysis

In the opening and closing vibration test of the through-sea valve driven by the original hydraulic machine controlled by the original distributor, it can be seen from Figure 15 that the peak values of vibration acceleration levels in three directions during the opening and closing process of the through-sea valve increase with the increase in working pressure. At working pressures of 8, 9, 10, and 12 MPa, the peak values of vibration acceleration levels in three directions during the closing process of the through-sea valve all exceed 120 dB, and the peak values of vibration acceleration levels during the closing process are greater than those during the opening process.

The vibration data of the through-sea valve controlled by the original distributor prototype at different working pressures is shown in

Table 1. Vibration conditions of the through-sea valve under different working pressures controlled by the original distributor.

Working Pressure		8 MPa	9 MPa	10 MPa	12 MPa
Open vibration direction	x/dB	109	112	113	117
	y/dB	107.5	112	113	117
	z/dB	106.5	110.5	112.5	115.5
Close vibration direction	x/dB	121.5	125	122.5	129
	y/dB	123	125	124	130
	z/dB	121	125	122	126.5

.

After the improvement design of the main valve sleeve of the electromagnetic hydraulic distributor, the vibration data of the through-sea valve controlled by the improved distributor prototype at different working pressures is shown in

Table 2. Vibration conditions of the through-sea valve under different working pressures controlled by the improved distributor prototype.

Working Pressure		8 MPa	9 MPa	10 MPa	12 MPa
Open vibration direction	x/dB	107	109	108	112
	y/dB	105.5	108	106	112
	z/dB	104.5	107.5	107.5	110.5
Close vibration direction	x/dB	115.5	119	117.5	126
	y/dB	115	119	117	121
	z/dB	117	118	117	119.5

.

After the improvement design of the main valve sleeve of the electromagnetic hydraulic distributor, compared with the vibration conditions of the through-sea valve under different working pressures controlled by the original distributor, the peak values of vibration acceleration levels in the x, y, and z directions during the opening and closing processes show the following reductions at various working pressures: at 8 MPa, the reductions are 2 dB, 2 dB, and 2 dB and 6 dB, 8 dB, and 4 dB, respectively; at 9 MPa, the reductions are 3 dB, 4 dB, and 3 dB and 6 dB, 6 dB, and 7 dB, respectively; at 10 MPa, the reductions are 5 dB, 7 dB, and 5 dB and 5 dB, 7 dB, and 5 dB, respectively; at 12 MPa, the reductions are 5 dB, 5 dB, and 5 dB and 3 dB, 9 dB, and 7 dB, respectively. The improved distributor prototype exhibits better transient noise control effects on the opening process of the through-sea valve than the closing process at 8 MPa and 9 MPa working pressures. As the working pressure continues to rise, the noise control effect of the distributor during the closing process of the through-sea valve becomes more pronounced.

The pressure oscillations caused by the electro-hydraulic distributor during the opening and closing processes of the sea valve are a significant source of transient noise. Experimental results indicate that grooving the flow passage of the main valve sleeve in the distributor can effectively reduce noise during the opening and closing of the sea valve. This further confirms that the approach of extending the response time to mitigate pressure oscillations is effective and aligns with simulation results. Furthermore, it is important to note that hydraulic system-generated noise is a significant component of submarine mechanical noise. The sea valve, being a crucial part of the hydraulic system, significantly impacts the acoustic stealth performance of the vessel due to excessive vibration noise during its opening and closing. Therefore, conducting in-depth research on the vibration and noise control technology of sea valves is crucial. Analyzing the causes of noise generation during the opening and closing process from a principle-based perspective and proposing effective improvement measures are of great significance. These actions are essential for completely solving the noise issues associated with sea valve operations and further enhancing the acoustic stealth performance of submarines.

5. Conclusions

This study focuses on the electromagnetic hydraulic distributor and its influence on the transient noise produced during the actuation of the through-sea valve. The research concludes with the following findings:

(1) The distributor’s structural analysis reveals the operational mechanism and identifies the source of high noise during the through-sea valve’s opening and closing cycles. To attenuate pressure fluctuations during the distributor’s opening phase, a strategy to introduce a slot in the main valve sleeve’s flow port is suggested, aiming to prolong the distributor’s response time.

(2) Theoretical calculations and simulations are conducted to determine the original distributor valve’s opening flow area, which is calculated to be 22.09 mm². To effectively mitigate the pressure impact from the distributor’s opening and closing while maintaining the operational flow, a triangular slot is proposed at one of the main valve sleeve’s flow holes, with specific dimensions established.

(3) Transient noise control efficacy testing on the distributor prototype is performed. The results indicate that slotting the flow hole significantly reduces the transient noise during the through-sea valve’s operation while keeping the working flow constant, with noise reduction becoming more significant at higher operating pressures. Additionally, the design approach for the structural improvement of the electro-hydraulic distributor can be applied to other types of valves.