Underwater vehicle and high-pressure hydraulic system and vibration reduction method thereof

Abstract

The invention discloses an underwater vehicle and a high-pressure hydraulic system and a vibration reduction method thereof, and the high-pressure hydraulic system of the underwater vehicle comprises a rack which comprises a horizontal fixing plate; the motor is fixedly mounted above the horizontal fixing plate; the booster pump is connected with the motor and located below the horizontal fixing plate, and the booster pump is a positive displacement pump and comprises an oil inlet and an oil outlet; the flexible connecting pipe is connected to the oil liquid inlet; and the rigid straight pipe is connected to the oil liquid outlet. A small section of flexible connecting pipe is additionally arranged at the front end of the oil inlet, and vibration caused when medium oil is sucked into a pipeline can be buffered. A rigid straight pipe is used at the front end of an oil liquid outlet and used for reinforcing and relieving pressure pulsation generated by medium oil at the outlet.

Description

Technical Field

[0001] This invention relates to the field of hydraulic system and marine equipment design and manufacturing, and particularly to an underwater vehicle and its high-pressure hydraulic system and vibration reduction method. Background Technology

[0002] In marine engineering, such as offshore platforms, underwater equipment, and deep-water drilling, hydraulic power units and high-pressure hydraulic systems are involved. However, existing hydraulic devices are generally not suitable for 20MPa high-pressure conditions. When the hydraulic station of marine equipment is upgraded from 10MPa to 20MPa, the noise and vibration problems of existing gear / plunger pumps are significantly aggravated. Existing high-pressure hydraulic systems have large vibration and noise levels, which cannot meet the special usage requirements of current marine equipment. Summary of the Invention

[0003] In view of the above-mentioned defects or improvement needs of the existing technology, the present invention provides an underwater vehicle and its high-pressure hydraulic system, and a vibration reduction method, which can reduce vibration and noise while ensuring the stable operation of the high-pressure hydraulic system of the underwater vehicle under high pressure.

[0004] To achieve the above objectives, the present invention adopts the following technical solution. In some embodiments, a high-pressure hydraulic system for an underwater vehicle is provided, the high-pressure hydraulic system of the underwater vehicle comprising: a platform including a horizontal fixed plate; a motor fixedly mounted above the horizontal fixed plate; a booster pump connected to the motor and located below the horizontal fixed plate, the booster pump being a positive displacement pump including an oil inlet and an oil outlet; a flexible connecting pipe connected to the oil inlet; and a rigid straight pipe connected to the oil outlet.

[0005] In some embodiments, the high-pressure hydraulic system of the underwater vehicle includes a plurality of rectifiers connected to the rigid straight pipe, and each rectifier includes a rectifier grid.

[0006] In some embodiments, the rectifier grid includes two rectifier layers, each rectifier layer having a baffle plate with a plurality of flow holes, the diameter of the plurality of flow holes being distributed from the center outwards in descending order.

[0007] In some embodiments, the plurality of flow holes include a central hole and circumferential holes, the circumferential holes being evenly distributed around the central hole, the angle between each circumferential hole and the axis being 3°, and the holes being inclined toward the central hole in the flow direction.

[0008] In some embodiments, the high-pressure hydraulic system of the underwater vehicle has three high-pressure relief valves, namely a first high-pressure relief valve, a second high-pressure relief valve, and a third high-pressure relief valve connected in sequence from near to far of the rigid straight pipe; the rectifier has two devices, namely a first rectifier and a second rectifier, the first rectifier being disposed between the first high-pressure relief valve and the rigid straight pipe, and the second rectifier being disposed between the first high-pressure relief valve and the second high-pressure relief valve.

[0009] In some embodiments, the high-pressure hydraulic system of the underwater vehicle includes a connecting frame assembly, which includes a connecting frame and a mounting plate. The motor and the booster pump are connected through the connecting frame, and the connecting frame is then fixed to the mounting plate. The mounting plate is connected to a horizontal fixed plate. The connecting frame has an injection hole with vibration damping rubber inside. The operating frequency or excitation direction of the connecting frame is different from that of the high-pressure hydraulic system of the underwater vehicle.

[0010] In some embodiments, a plurality of vibration isolators are provided between the horizontal fixed plate and the mounting plate; the upper and lower planes of the vibration isolators are kept horizontal, with the upper surface as a reference and the parallelism error of the lower surface not exceeding 0.02 mm; the compression of the plurality of vibration isolators is consistent, with an error not exceeding 0.2 mm; the working compression of the plurality of vibration isolators is 65% of the ultimate compression, and the deviation of the working compression of all vibration isolators does not exceed 3%.

[0011] In some embodiments, the booster pump is a screw pump, which includes an oil return device and a sealing assembly, a pump body bushing, a driven rod bushing, a front cover, a sealing cover, and a main rod; the oil return device includes an oil return valve disposed at one end of the oil inlet of the screw pump; the sealing assembly includes a first sealing ring, a second sealing ring, a third sealing ring, and a mechanical seal assembly, wherein the first sealing ring is disposed between the pump body bushing and the front cover, the second sealing ring is disposed between the driven rod bushing and the front cover, the third sealing ring is disposed between the front cover and the sealing cover, and the mechanical seal assembly is disposed between the sealing cover and the main rod; the first sealing ring is elliptical.

[0012] In some embodiments, a vibration reduction method for the high-pressure hydraulic system of the underwater vehicle described above is also provided. The vibration reduction method includes: correcting the screw profile, wherein the correction of the screw profile includes screw profile correction for pressure pulsation and screw profile correction for screw meshing interference; the screw profile correction for screw meshing interference includes setting the end shape of the non-working surface of the main rod to an arc shape; adjusting the angle of the vibration isolators so that the angle of each vibration isolator matches the vibration energy transmission path; adjusting the preload of the fasteners of the vibration isolators, wherein the error of the preload of each fastener does not exceed ±5N / m.

[0013] In some embodiments, an underwater vehicle is also provided, the underwater vehicle including the high-pressure hydraulic system of any of the above-described underwater vehicles.

[0014] Compared to the prior art, the beneficial effects of the present invention include at least the following: In the embodiments of this application, a short flexible connecting pipe is added to the front end of the oil inlet, which can buffer the vibration caused when the medium oil is sucked into the pipeline. A rigid straight pipe is used at the front end of the oil outlet to reinforce and alleviate the pressure pulsation of the medium oil at the outlet. It should be noted that the technical effects of the embodiments of this application are not limited to these, and the specific advantages and effects are reflected in the following detailed descriptions. Attached Figure Description

[0015] Figure 1 This is a partial cross-sectional view of a screw pump in one direction according to some embodiments of this application.

[0016] Figure 2 This is a partial cross-sectional view from another direction of the screw pump in some embodiments of this application.

[0017] Figure 3 This is a schematic diagram of the front cover in some embodiments of this application.

[0018] Figure 4 This is a schematic diagram of the main rod in some embodiments of this application.

[0019] Figure 5 This is a schematic diagram of the slave rod in some embodiments of this application.

[0020] Figure 6 This is a schematic diagram of the follower bushing in some embodiments of this application.

[0021] Figure 7 This is a schematic diagram of the pump body bushing in some embodiments of this application.

[0022] Figure 8 This is an enlarged schematic diagram of a portion of the structure of a screw pump in some embodiments of this application.

[0023] Figure 9 This is a schematic diagram of the bushing from another direction in some embodiments of this application.

[0024] Figure 10 This is a schematic diagram of the return oil positioning pin in some embodiments of this application.

[0025] Figure 11 This is a cross-sectional schematic diagram of the main rod in some embodiments of this application.

[0026] Figure 12 This is a schematic diagram of the return valve in some embodiments of this application.

[0027] Figure 13 This is a schematic diagram of the assembly structure of the positioning component in some embodiments of this application.

[0028] Figure 14 This is a schematic diagram of screw profile correction in some embodiments of this application.

[0029] Figure 15 This is a schematic diagram of a bearing positioning and installation fixture in some embodiments of this application.

[0030] Figure 16 This is a schematic diagram of the high-pressure hydraulic system of an underwater vehicle in some embodiments of this application.

[0031] Figure 17 This is a cross-sectional schematic diagram of a rectifier device in some embodiments of this application.

[0032] Figure 18 This is a schematic diagram of the flow passage in some embodiments of this application.

[0033] Figure 19 This is a cross-sectional view of the flow passage in some embodiments of this application from another direction.

[0034] Figure 20 This is a schematic diagram of the connecting frame assembly in some embodiments of this application.

[0035] Figure 21 This is a cross-sectional schematic diagram of the connecting frame in some embodiments of this application.

[0036] Figure 22 This is a schematic diagram of the connecting frame in some embodiments of this application.

[0037] Figure 23 This is a schematic diagram of a vibration isolator in some embodiments of this application.

[0038] Figure 24 This is a partial structural schematic diagram of a high-pressure hydraulic system in some embodiments of this application. Detailed Implementation

[0039] To make the objectives, technical solutions, and advantages of this invention clearer, the invention will be further described in detail below with reference to the accompanying drawings and embodiments. It should be understood that the specific embodiments described herein are merely illustrative and not intended to limit the invention.

[0040] In some embodiments of this application, a screw pump is provided. The screw pump is a positive displacement pump that uses two or more rotating screws to create volume changes within the stator of the pump body, thereby continuously propelling the transported medium forward. As the screws rotate, the volume within the pump chamber changes accordingly, drawing in the medium and then pushing it towards the outlet. Screw pumps have good sealing performance and can transport media of various viscosities, thus finding wide application in chemical, petroleum, pharmaceutical, and food industries. In embodiments of this application, the screw pump, as a hydraulic power unit, can be used in underwater vehicles as part of their high-pressure hydraulic system, providing a high-pressure hydraulic source.

[0041] Figure 1 This is a partial cross-sectional view of a screw pump in one direction according to some embodiments of this application. Figure 2 This is a partial cross-sectional view from another direction of a screw pump according to some embodiments of this application. (Reference) Figure 1 and Figure 2 In some embodiments, the screw pump includes a stator 100 and a rotor 200. In some embodiments, the rotor includes a main rod 201 and two driven rods 202. In embodiments of this application, the screw pump is a three-screw pump. The two driven rods 202 are disposed on both sides of the main rod 201. In embodiments of this application, the screw pump is a high-pressure screw pump. In some embodiments, a high-pressure screw pump refers to a screw pump with a maximum operating pressure exceeding 20 MPa. In some embodiments, the maximum operating pressure of the screw pump exceeds 20 MPa. In some embodiments, the maximum operating pressure of the screw pump reaches 25 MPa.

[0042] In some embodiments, the stator is a pump body bushing 101. In embodiments of this application, the stator is a coupling between the pump body and the bushing, and the rotor consists of three screws, comprising one main screw and two driven screws. The main screw is the driving screw connected to the drive device, and the two driven screws are driven screws that mesh with the main screw. The main and driven screws, together with the pump body bushing, form the stator and rotor structure of a screw pump.

[0043] In some embodiments, the pump body bushing is made of a bimetallic material, comprising an outer layer of alloy steel and an inner layer of copper alloy. While the overall structure is made of high-strength alloy steel, the inner layer is sintered with alloy copper. In the embodiments of this application, the pump body and bushing are sintered together using a sintering process, combining the two different materials, alloy steel and copper alloy, into a single unit called the pump body bushing. This improves the pressure resistance and reliability of the equipment, saves space, and facilitates maintenance. In the embodiments of this application, the inner alloy copper has low hardness, reducing the frictional power generated when the screw rotates at high speed within the pump body bushing. The outer alloy steel material has high strength, enhancing the overall pressure resistance of the equipment. Through the application of bimetallic materials, the stability of the equipment under ultra-high pressure operation is ensured, while also improving the overall reliability of the equipment.

[0044] In some embodiments, the screw pump includes an oil chamber, which comprises an inlet chamber 601 and a high-pressure chamber 602. In some embodiments, the screw pump includes a front cover 6021 and a rear cover 6011. The front cover and the rear cover are located at opposite ends of the stator.

[0045] In some embodiments, the rear cover is fixedly connected to the stator, and the rear cover has an oil inlet 6012. The space between the rear cover and the stator and rotor forms an inlet chamber.

[0046] In some embodiments, the front cover is fixedly connected to the stator, and the space between the front cover, the stator, and the rotor forms a high-pressure chamber. The stator has an oil outlet 102 at the location corresponding to the high-pressure chamber.

[0047] In the embodiments of this application, the inlet chamber 601 is connected to the oil inlet, and the high-pressure chamber 602 is connected to the oil outlet. The medium oil enters the inlet chamber 601 through the oil inlet, and the screw creates a volume change in the stator, causing the medium oil to be continuously propelled forward to reach the high-pressure chamber 602, where it flows out from the oil outlet to provide high-pressure hydraulic oil.

[0048] In some embodiments, the rear cover has an inlet flange, and the stator has an outlet flange. The inlet flange and the outlet flange are respectively connected to the oil inlet 6012 and the oil outlet 102. Both the inlet flange and the outlet flange are SAE standard flanges, which facilitates installation and standardization. The length directions of the inlet flange and the outlet flange are perpendicular to each other, which reduces the installation size while ensuring installation stability.

[0049] In some embodiments, the oil chamber 600 further includes a sealing chamber 603. The sealing chamber 603 is located at the end of the main rod away from the inlet chamber 601. The screw pump includes a sealing cover 6031, which is fixedly connected to the front cover 6021. The space between the sealing cover, the front cover, and the main rod forms the sealing chamber.

[0050] In some embodiments, the screw pump includes a positioning assembly, which is mainly used for positioning the main rod and the driven rod. In some embodiments, the positioning assembly includes a main rod positioning assembly and a driven rod positioning assembly. In some embodiments, the positioning assembly includes a spacer sleeve 501, a balance ring 502, a bearing 503, a support ring 504, and a nut 505. The spacer sleeve 501, balance ring 502, bearing 503, support ring 504, and nut 505 are mainly used for positioning the main rod and constitute the main rod positioning assembly.

[0051] In some embodiments, the follower rod positioning assembly includes a follower rod bushing 506 and a return oil positioning pin 507. The follower rod bushing 506 is disposed at the end of the follower rod located in the high-pressure chamber. One end of the return oil positioning pin 507 is connected to the front cover, and the other end is inserted into the follower rod bushing 506. In some embodiments, the follower rod bushing 506 serves to balance the radial force of the follower rod.

[0052] Figure 3 This is a schematic diagram of the front cover in some embodiments of this application. (Reference) Figure 3 In some embodiments, the front cover 6021 has a main rod shaft hole, which includes a shaft hole sealing section 6024. In some embodiments, the front cover includes a tapered section 6022 and a flange connection section 6025, the tapered section 6022 being connected to the flange connection section 6025. The flange connection section 6025 is used for fixed installation with a stand. In the embodiments of this application, the tapered section 6022 not only enhances the overall strength of the front cover but also facilitates easier alignment during the installation of the screw pump.

[0053] Figure 4 This is a schematic diagram of the main rod in some embodiments of this application. (Reference) Figure 4 The main rod 201 includes a main rod screw section 2011, a main rod connecting section 2012, a main rod sealing section 2013, and a main rod bearing section 2014 arranged sequentially from the liquid inlet chamber to the sealing chamber. The main rod sealing section 2013 is assembled and connected to the shaft hole sealing section, and the main rod bearing section 2014 is supported and connected to the front cover 6021 through a bearing 503.

[0054] In some embodiments, the diameter Dm of the main rod sealing section, the outer diameter Dn of the main rod screw section, and the outer diameter Dc of the main rod connecting section satisfy the following: .

[0055] In the embodiments of this application, the main rod sealing section 2013 on the main rod mainly serves to balance the axial force, and the bearing 503 balances the remaining axial force and radial force, and at the same time plays an axial positioning role.

[0056] In some embodiments, the diameter Dm of the main rod sealing section, the outer diameter Dn of the main rod screw section, and the outer diameter Dc of the main rod connecting section satisfy the following: In some embodiments, bearing 503 is a deep groove ball bearing. In the embodiments of this application, the main rod sealing section substantially balances all axial forces, and the bearing is substantially unaffected by axial forces. The adjustment of the axial clearance of the main rod relies on the cooperation of the front cover, spacer sleeve, bearing, balance ring, support ring, and nut. In the embodiments of this application, no additional wave springs or other displacement compensation structures are required to meet the axial positioning requirements.

[0057] Figure 5 This is a schematic diagram of the slave rod in some embodiments of this application. Figure 6 This is a schematic diagram of the follower bushing in some embodiments of this application. (Reference) Figure 5 and Figure 6In some embodiments, the follower rod 202 includes a follower rod screw section 2021 and a follower rod shaft end section 2022. The follower rod bushing 506 has a follower rod shaft hole 5061, a bushing support section 5062, and a bushing end section 5063. The follower rod shaft end section 2022 is located within the follower rod shaft hole 5061, and the bushing support section 5062 of the follower rod bushing has a positioning hole 5064 that engages with the return oil positioning pin 507, and the positioning hole communicates with the follower rod shaft hole.

[0058] In some embodiments, the outer diameter of the driven rod screw section is Df, and the outer diameter of the driven rod shaft end section is Dt. The outer diameter Dt of the driven rod shaft end section and the outer diameter Df of the driven rod screw section satisfy the following condition: In embodiments of this application, a follower neck section 2023 is provided between the follower screw section and the follower shaft end section. The follower neck section 2023 is tapered, and its outer diameter gradually decreases from the follower screw section 2021 to the follower shaft end section 2022. The outer diameter Dj at the connection between the follower neck section 2023 and the follower shaft end section 2022 is smaller than the outer diameter Dt of the follower shaft end section. In some embodiments, the outer diameter Dj and the outer diameter Dt of the follower shaft end section satisfy the following relationship: In some embodiments, the outer diameter Dz of the bushing support section is equal to the outer diameter Df of the follower rod screw section. In some embodiments, the outer diameter Dj at the connection between the follower rod neck section 2023 and the follower rod shaft end section 2022 satisfies the condition that the outer diameter Df of the follower rod screw section is equal to the outer diameter Df of the follower rod screw section. The neck section of the 2023 adopts a tapered structure and meets the requirements. To ensure connection strength.

[0059] In the embodiments of this application, the outer diameter of the driven rod shaft end is increased by more than 15%, which basically offsets the axial force generated by the high-pressure medium oil pushing the driven rod and improves the reliability of the driven rod during operation.

[0060] Figure 7 The diagram shows a pump body bushing in some embodiments of this application. (a) is a perspective view and (b) is a cross-sectional view. Figure 8 This is an enlarged schematic diagram of a portion of the structure. Figure 9 This is a schematic diagram of the rod bushing from another direction in some embodiments of this application. (See reference) Figure 1 , Figures 6-9In some embodiments, the pump body bushing 101 has a countersunk hole 1011, and the bushing end section 5063 is located within the countersunk hole. Specifically, in some embodiments, the follower bushing 506 includes a contact boss 5065, which protrudes from the first end face 5066 of the follower bushing. The first end face is a radial surface connecting the bushing support section 5062 and the bushing end section 5063. In the embodiments of this application, the contact boss 5065 abuts against the countersunk surface of the countersunk hole 1011. In some embodiments, the contact surface 5067 where the contact boss 5065 abuts against the countersunk hole 1011 is a plane, and its flatness is required to be high.

[0061] In some embodiments, the follower sleeve 506 has a second end face 5068. The contact surface 5067 of the contact boss is parallel to the second end face 5068 and is completely perpendicular to the central axis of the follower sleeve 506. In embodiments of this application, the first end face does not contact the countersunk hole 1011, and there are no special requirements for the surface roughness and flatness of the first end face.

[0062] In the embodiments of this application, the contact surface between the driven rod bushing 506 and the countersunk hole 1011 is the protruding surface of the contact boss 5065, which greatly reduces the contact area between the driven rod bushing and the countersunk hole. This avoids the problem of poor parallelism caused by limitations in the processing technology, and ensures that after the driven rod bushing is installed in the pump body bushing, its end face fits more smoothly with the front cover, which enhances the sealing performance and prevents excessive passage of high-pressure medium oil from causing the driven rod bushing to be stressed or even rotated.

[0063] In some embodiments, the area S1 of the contact surface of the contact boss and the area S2 of the first end face satisfy the following conditions: In some embodiments, the area S1 of the contact surface of the contact boss and the area S2 of the first end face satisfy the following condition: In the embodiments of this application, the area S1 of the contact surface of the contact boss and the area S2 of the first end face satisfy the above relationship, which ensures contact strength while avoiding the difficulty in ensuring flatness and parallelism due to an excessively large contact area. In some embodiments, the radial width of the contact boss is 2-5mm, and the protrusion height of the contact boss, i.e., the axial length, is 5-8mm. This setting of the width and height of the contact boss facilitates processing, ensures contact strength, and maximizes parallelism.

[0064] In some embodiments of this application, the screw pump includes a driven rod bushing, the driven rod bushing having a bushing support section and a bushing end section, the driven rod including a driven rod screw section and a driven rod end section, and the outer diameter of the bushing support section being equal to the outer diameter of the driven rod screw section.

[0065] In some embodiments, the outer periphery of the bushing end section 5063 has an arc segment 50631 and a straight segment 50632. The outer diameter of the arc segment is larger than the outer diameter of the bushing support section, and the distance between the straight segment and the axis of the follower bushing is smaller than the outer diameter of the bushing support section. In the embodiments of this application, the straight segment 50632 is provided, which simplifies the manufacturing process. Furthermore, by setting the distance between the straight segment and the axis of the follower bushing to be smaller than the outer diameter of the bushing support section, interference can be minimized.

[0066] Figure 10 This is a schematic diagram of the return oil positioning pin in some embodiments of this application. (Reference) Figure 10 Specifically, in some embodiments, the return oil positioning pin 507 has a central connecting hole 5071. The front cover has a rod return oil hole 6023 that matches the return oil positioning pin 507.

[0067] In the embodiments of this application, a return oil positioning pin is inserted into the follower rod bushing 506 of the follower rod. The return oil positioning pin is made of alloy steel. A hole is drilled in the center of the return oil positioning pin to make the interior hollow, forming a passage for the medium oil to flow through, namely a connecting hole 5071, thereby increasing the amount of medium oil passing through and ensuring the lubrication of the bearing and mechanical seal assembly.

[0068] In some embodiments, the central axis of the rod shaft hole 5061 does not coincide with the central axis of the positioning hole 5064, i.e., the positioning hole 5064 is eccentrically positioned. Correspondingly, the central axis of the rod return oil hole 6023 coincides with the central axis of the positioning hole 5064, and the rod return oil hole 6023, the positioning hole 5064, and the return oil positioning pin 507 are positioned and engaged. In the embodiments of this application, the positioning hole 5064 is eccentrically positioned, and the outer periphery of the bushing end section 5063 has an arc segment 50631 and a straight segment 50632, which can effectively form circumferential positioning and prevent the rod bushing from rotating.

[0069] Figure 11 This is a schematic cross-sectional view of the main rod in some embodiments of this application. (Reference) Figure 11 and Figure 1 In some embodiments, the screw pump includes a return oil device, which includes a return oil hole 301 and a return oil control component 300. The return oil hole is disposed inside the main rod, with a first end connected to a sealing cavity 603 and a second end connected to an inlet cavity 601. The return oil control component is used to open when the return oil pressure reaches a preset pressure, allowing oil to be drawn back from the sealing cavity to the inlet cavity through the return oil hole.

[0070] In the embodiments of this application, a through hole is machined in the main rod to form an oil return hole. An oil return control component is installed at the end of the main rod and is opened when the oil return pressure in the sealing cavity reaches the design pressure. The medium oil is led back from the sealing cavity to the liquid inlet cavity through the oil return hole in the main rod.

[0071] Figure 12 This is a schematic diagram of the return valve in some embodiments of this application. (Reference) Figure 12 In some embodiments, the oil return control assembly includes an oil return valve disposed within the inlet chamber. The oil return valve includes an oil return seat 302, a compression spring 303, and a ball core 304. The oil return seat is disposed at one end of the main rod within the inlet chamber, and the compression spring is disposed between the oil return seat and the ball core. Under the initial elastic force of the compression spring, the ball core abuts against the second end of the oil return hole located within the inlet chamber. In the embodiments of this application, the oil return control assembly is disposed within the inlet chamber, directly utilizing the space of the inlet chamber without requiring additional space for the oil return control assembly. Furthermore, it employs the simplest oil return valve structure, resulting in a purely mechanical oil return valve structure that provides stable and reliable control.

[0072] In some embodiments of this application, the main rod has a return oil valve hole 305 at one end of the liquid inlet chamber, and the return oil valve is disposed within the return oil valve hole 305. High-pressure oil return of a three-screw pump typically uses an external copper pipe or a bushing return oil hole. While some structures have a built-in return oil hole in the screw, this application's embodiments abandon the external copper pipe approach, ensuring the integrity of the high-pressure screw pump. Simultaneously, an innovative approach is adopted: an opening is made at the location where the mechanical seal assembly is installed on the main rod, with the entire return oil hole extending from the sealing chamber to the tail end of the main rod. A return oil valve structure is added at the tail end, ensuring the medium oil always fills the return oil hole. Even if the return oil valve is open, it will not cause impact to the return oil hole, avoiding vibration caused by return oil impact. In the embodiments of this application, the return oil volume can be automatically adjusted by the return oil pressure, allowing for better control of the oil volume and pressure within the sealing chamber. Under the operating conditions of the bearings and mechanical seals, the medium oil can provide maximum lubrication and heat dissipation, improving the overall reliability of the equipment.

[0073] In some embodiments of this application, the oil return hole includes a first hole 3011, a second hole 3012, and a third hole 3013. The first hole is a through hole arranged radially along the main rod and directly communicates with the sealing cavity. The second hole is arranged axially along the main rod, with one end connected to the first hole and the other end connected to the oil return valve hole 305. The third hole is a through hole arranged radially along the main rod, with its axis coplanar and perpendicular to the axis of the oil return valve hole. The third hole is divided into two segments by the oil return valve hole, with one end of each segment connected to the oil return valve hole and the other end connected to the inlet cavity. The axis of the third hole is coplanar with the axis of the rotor. When the oil return valve is closed, the ball core 304 abuts against the second hole 3012.

[0074] In some embodiments, the screw pump includes a sealing assembly. In some embodiments, the sealing assembly is used to ensure that the medium oil of the entire screw pump does not leak out. In some embodiments, the sealing assembly includes at least a first sealing ring 701 and a second sealing ring 702. Specifically, in some embodiments, a first sealing ring 701 is provided between the front cover and the stator and the driven rod bushing.

[0075] In some embodiments, the first sealing ring 701 is non-circular in shape. In some embodiments, the first sealing ring 701 is elliptical in shape. In the embodiments of this application, the first sealing ring used for sealing between the stator and the front cover bears high pressure. Using an elliptical shape instead of a circular shape can reduce the pressure-bearing area of ​​the first sealing ring, reduce the high-pressure shear force at this point, and enhance the reliability of the first sealing ring structure. In the embodiments of this application, the elliptical first sealing ring surrounds the main rod and the two slave rod bushings, thus minimizing the size of the first sealing ring. In some embodiments, a sealing groove 1014 is provided on the stator, and the sealing groove 1014 matches the shape of the first sealing ring 701. In some embodiments, the sealing groove is also elliptical in shape.

[0076] Specifically, in some embodiments, a second sealing ring 702 is provided between the front cover, the return oil positioning pin, and the driven rod bushing. In the embodiments of this application, the screw pump operates at very high pressure, and high-pressure medium oil will flow into the space between the driven rod bushing and the front cover. By adding a second sealing ring at the return oil positioning pin, the outflow of high-pressure medium oil is prevented. At the same time, the return oil positioning pin also plays a positioning role for the driven rod bushing. Under the combined action of the two, rotation of the driven rod bushing is prevented during operation. In some embodiments, the return oil positioning pin has a positioning shoulder 5072, and the second sealing ring is located between the positioning shoulder, the front cover, and the driven rod bushing.

[0077] In some embodiments, the sealing assembly further includes a third sealing ring 703. The third sealing ring 703 is disposed between the front cover 6021 and the sealing cover 6031. In some embodiments, the sealing assembly further includes a mechanical seal assembly 704. The mechanical seal assembly 704 is disposed between the main rod 201 and the sealing cover 6031. In some embodiments, the mechanical seal assembly is a rotating sealing ring. It can ensure a reliable seal between the main rod 201 and the sealing cover 6031 while they are rotating. In the embodiments of this application, the sealing assembly includes a first sealing ring, a second sealing ring, a third sealing ring, and a mechanical seal assembly. The first sealing ring is disposed between the pump body bushing and the front cover, the second sealing ring is disposed between the driven rod bushing and the front cover, the third sealing ring is disposed between the front cover and the sealing cover, and the mechanical seal assembly is disposed between the sealing cover and the main rod. Through the first sealing ring, the second sealing ring, the third sealing ring, and the mechanical seal assembly of this application, a reliable overall seal is achieved between the high-pressure chamber and the sealing chamber of the screw pump.

[0078] In some embodiments of this application, a sealing and vibration damping structure for a screw pump is also provided, the sealing and vibration damping structure including an oil return device. The oil return device includes an oil return hole and an oil return control component. The oil return hole is disposed within the main shaft of the screw pump, with a first end connected to the sealing cavity of the screw pump and a second end connected to the inlet cavity of the screw pump. The oil return control component is used to open when the oil return pressure reaches a preset pressure, allowing oil to be drawn back from the sealing cavity to the inlet cavity through the oil return hole. In embodiments of this application, the oil return device is the oil return device of any of the above embodiments.

[0079] In some embodiments, the sealing and vibration damping structure further includes a sealing assembly. The sealing assembly includes at least a first sealing ring; the first sealing ring is disposed between the front cover and the stator of the screw pump, and the first sealing ring is non-circular in shape. In embodiments of this application, the sealing assembly is the sealing assembly of any of the above embodiments.

[0080] In some embodiments, the fitting clearance between the main rod sealing section 2013 and the shaft hole sealing section of the front cover is 0.06-0.08 mm. The length Lm of the main rod sealing section and the length Ln of the main rod screw section satisfy: In the embodiments of this application, the pressure of the conveyed medium oil is very high, and the structure of the main rod sealing section and the front cover has been specifically designed. No other components are installed on the shaft hole sealing section of the front cover; the front cover directly mates with the main rod sealing section. In this embodiment, a fitting clearance of 0.06-0.08 mm is used, which is smaller than that of a typical screw pump. This reduces the leakage of high-pressure medium oil at this point, improving the overall volumetric efficiency of the high-pressure screw pump and the overall working efficiency of the equipment to meet the 20 MPa working pressure. It also ensures that the internal oil circulation volume remains within an acceptable range during long-term operation, preventing internal overheating caused by excessive leakage. In this embodiment, the sealing performance is further enhanced by increasing the axial length of the main rod sealing section.

[0081] It is understood that the screw pump in this embodiment delivers a maximum pressure of 25 MPa. The leakage within the screw pump is precisely controlled. The gap between the main rod sealing section and the front cover is one of the main leakage points. The size of this gap determines the leakage; a larger gap results in greater leakage, which leads to a decrease in volumetric efficiency and prevents the output pressure from reaching above 20 MPa. Conversely, too small a gap results in lower pressure and flow rate entering the sealing cavity, hindering bearing heat dissipation and mechanical seal lubrication. Furthermore, an excessively small gap increases the risk of wear between the main rod sealing section and the front cover, thus affecting the reliability of the screw pump. In this embodiment, a clearance of 0.06-0.08 mm is used to achieve the maximum pressure while minimizing wear, thus meeting high reliability requirements while maintaining high output pressure.

[0082] In some embodiments, the main rod sealing section 2013 has a plurality of main rod sealing grooves arranged axially, the number of which is 2-4. In some embodiments, the number of main rod sealing grooves is 2. Each main rod sealing groove has a groove width of 2-3 mm and a groove depth of 2-4 mm. In the embodiments of this application, two main rod sealing grooves are provided on the main rod sealing section, which can form a labyrinth seal at the main rod sealing section and store oil in the main rod sealing grooves, ensuring the lubrication of the mating area between the main rod sealing section and the front cover, and preventing the risk of "shaft seizure" caused by reducing the mating clearance at this point.

[0083] In some embodiments, the gap between the bushing support section 5062 of the rod bushing 506 and the pump body bushing is 0.02 to 0.04 mm, which reduces the leakage of high-pressure medium oil from this point and facilitates installation.

[0084] In some embodiments, the length Lt of the driven shaft end section and the length Lf of the driven screw section satisfy: In the embodiments of this application, the end section of the driven rod shaft is lengthened, which improves the torsional resistance of the driven rod and allows it to withstand more than 30% more radial force than the driven rod of a conventional medium and low pressure pump. When operating at 20MPa, it can meet the requirements for normal operation.

[0085] In some embodiments, the driven rod end section has a plurality of driven rod sealing grooves arranged axially, the number of which is 2-4. In some embodiments, the number of driven rod sealing grooves is 2. Each driven rod sealing groove has a groove width of 1.5-2.5 mm and a groove depth of 2-3 mm. In the embodiments of this application, the same scheme as the main rod sealing section is adopted at the driven rod end section, with 2 sealing grooves provided to form a labyrinth seal, increasing the sealing performance at this location and preventing excessive leakage of high-pressure medium oil into the sealing cavity.

[0086] Figure 13Schematic diagram of the assembly structure of the positioning component in some embodiments of the present application. Refer to Figure 13 , in some embodiments, the spacer sleeve 501 and the balance ring 502 are both installed in the sealing cavity. The inner diameter of the spacer sleeve 501 is larger than the outer diameter of the balance ring 502. The first end of the spacer sleeve 501 contacts the front cover 6021, and the second end of the spacer sleeve 501 contacts the outer ring of the bearing. The first end of the balance ring 502 contacts the main rod sealing section 2013, and the second end of the balance ring 502 contacts the inner ring of the bearing. There are gaps between the spacer sleeve and the balance ring both axially and radially, for the dielectric oil to flow through to reach the bearing and the mechanical seal assembly, providing heat dissipation and lubrication effects for the bearing and the mechanical seal assembly.

[0087] In some embodiments, the inner diameter dimension of the spacer sleeve is 55 - 60 mm, the outer diameter dimension is 80 - 85 mm, and the axial length is 9 - 11 mm. The outer diameter dimension of the balance ring is 50 - 55 mm, the inner diameter dimension is 35 - 50 mm, and the axial length is 8 - 10 mm. In the embodiments of the present application, the difference between the inner diameter of the spacer sleeve and the outer diameter of the balance ring is DW1, the inner diameter of the return oil hole 6023 of the rod is DW2, and the difference between the outer diameter of the inner ring 5031 of the bearing and the inner diameter of the outer ring 5032 of the bearing is WD3 = 6 mm. Among them, DW1 < DW2, and DW1 < DW3. In some embodiments, the axial length of the spacer sleeve is greater than the axial length of the balance ring. In some embodiments, the difference between the axial length of the spacer sleeve and the axial length of the balance ring is 1 mm. In the embodiments of the present application, by strictly controlling the axial dimensions of the spacer sleeve and the balance ring, there is an assembly gap 5012 between the axial direction of the spacer sleeve and the axial direction of the balance ring, ensuring the smooth rotation of the main rod and the smooth flow path of the dielectric oil.

[0088] In the embodiments of the present application, by strictly controlling the dimensions of the spacer sleeve 501 and the balance ring 502, the flow rate of the dielectric oil reaching the bearing and the mechanical seal assembly is controlled, avoiding problems such as increased leakage, reduced volumetric efficiency, and insufficient output pressure caused by too large over - flow rate, and at the same time avoiding the "jamming" phenomenon between the driving screw and the driven screw caused by the sharp rise in the temperature of the pump body.

[0089] In some embodiments, the meshing gap between the main rod and the driven rod is 0.025 - 0.035 mm. In the embodiments of the present application, the meshing gap between the driving screw and the driven screw determines the thickness of the oil film. If the oil film is too thin, its stiffness is insufficient and it cannot effectively support the high - speed rotating screw. If the oil film is too thick, the meshing state will be unstable. In the embodiments of the present application, when the meshing gap between the main rod and the driven rod is 0.025 - 0.035 mm, the vibration and noise are the lowest, and the meshing state is the best and the operation is the most stable. In some embodiments, the meshing gap between the main rod and the driven rod can be 0.025 mm, 0.03 mm, or 0.035 mm.

[0090] In the embodiments of this application, since the working pressure of the high-pressure screw pump set is very high, the vibration excitation source of the stator and rotor of the high-pressure screw pump is mainly in the meshing cavity formed by the master and slave screws and the pump body bushing, and is transmitted in the pump body flow channel. The most effective method is to minimize the excitation force. In the embodiments of this application, the screw profile can be appropriately modified to solve this problem.

[0091] In some embodiments, screw profile correction includes screw profile correction for pressure pulsation and screw profile correction for screw meshing interference. Specifically, in some embodiments, a dedicated form milling cutter is designed and manufactured based on profile data calculations to ensure screw machining accuracy and efficiency, resulting in smoother screw meshing. In some embodiments, rough machining is first performed using a dedicated form milling cutter, followed by nitriding, and then finish machining to completion using a dedicated high-precision screw grinding machine, further improving the machining accuracy and surface finish of the screw profile and reducing lead error.

[0092] In some embodiments, screw profile correction for pressure pulsation includes reducing the screw pitch P. Specifically, the screw pitch of the main rod and the driven rod is 10-20 mm. In some embodiments, the screw pitch of the main rod and the driven rod is 10 mm, and the length of the main rod screw section 2011 is 20 cm. In the embodiments of this application, with a fixed length of the main rod screw section 2011, the screw pitch of the main rod and the driven rod is set to 10 mm. By setting the screw pitch to a smaller value, the total number of threads increases. By increasing the number of screw pitches, the pressure difference between the screw sealing chamber stages is reduced, thereby reducing pressure pulsation. In the embodiments of this application, although a three-screw pump theoretically should not generate pressure pulsation, actual engineering has shown that machining profile errors and operating clearances can cause internal leakage between the screw, pump body bushing, and sealing chamber, resulting in pressure pulsation. Pressure pulsation causes system vibration in the three-screw pump. Under the same conditions, the higher the pressure, the greater the pulsation value. In the embodiments of this application, the vibration caused by pressure pulsation is reduced by reasonably setting the screw pitch. In some embodiments, after the screw is machined, a special screw inspection instrument is used to inspect the profile. Based on the inspection results, an abrasive grinding process is added to further eliminate residual sharp points in the screw machining, thereby further reducing the possibility of screw meshing interference and reducing or even avoiding pressure pulsation.

[0093] Figure 14 This is a schematic diagram illustrating screw profile corrections in some embodiments of this application. (Reference) Figure 14In some embodiments, the screw profile correction for screw meshing interference includes setting the end shape of the non-working surface 2015 of the main rod to an arc shape. In the embodiments of this application, the non-working surface of the main rod is the threaded surface away from the high-pressure chamber. The end of the non-working surface 2015 is the radially outward outermost edge. In the embodiments of this application, the end shape of the non-working surface of the main rod is arc-shaped. In the embodiments of this application, the high-pressure screw pump set is a three-screw pump structure, and its helical theory forms a gear tooth profile curve for point meshing. Theoretically, the helices of the driving screw and the driven screw will not interact during meshing, and therefore will not produce vibration after physical contact. However, it was found that vibration noise still occurs during actual operation. In the embodiments of this application, after eliminating various factors and optimizing the design, vibration noise still occurs during startup and shutdown. After various tests and theoretical verifications, it was finally found that during startup and shutdown of the screw pump, the change in pressure in the inlet chamber and the high-pressure chamber reduces the torque of the medium oil driving the driven screw to rotate, which is insufficient to overcome the frictional resistance of rotation. Although the driving screw still drives the driven screw to rotate, the driving force of the driven screw has changed from relying on medium transmission to relying on the profile of the driving screw. At this time, the driving screw and the driven screw will interact with each other, and the physical contact between the driving screw and the driven screw will generate vibration noise. In the embodiments of this application, the end of the non-working surface of the main rod is rounded to make its shape "arc", so as to minimize the contact area between the driving screw and the driven screw, thereby reducing the vibration noise caused by physical contact.

[0094] In the embodiments of this application, the pump body bushing serves as the stator, forming a sealed volume chamber with the high-speed rotating rotor. The high-precision, low-noise pump body bushing is a bottleneck component that restricts the low-noise, high-pressure screw pump. Its processing quality directly affects the vibration and noise level of the entire machine, especially the processing accuracy of the three holes corresponding to the three screws in the pump body bushing.

[0095] In some embodiments, the pump body bushing meshes with the master and slave rods. The pump body bushing includes three holes 1012 corresponding to the three screws, including one master rod hole 10121 and two slave rod holes 10122, which together form the three holes 1012 of the pump body bushing. In some embodiments of this application, the stator has a master rod hole, which includes a master rod threaded mating hole section 10123 and a master rod connecting mating hole section 10124. The inner diameter of the master rod connecting mating hole section is larger than the inner diameter of the master rod threaded mating hole section.

[0096] The dimensional tolerances, geometric tolerances, and surface roughness of the three holes in the pump body bushing directly affect the smoothness and uniformity of the meshing between the pump body bushing and the master and slave screws. In some embodiments of this application, the straightness and coaxiality of the three holes in the pump body bushing are both less than or equal to 0.01 mm. Simultaneously, the machining error of the three holes is precisely controlled to 0.01 mm to 0.02 mm to ensure the clearance between the master and slave screws of the rotor and the pump body bushing of the stator, forming a stable oil film thickness and reducing vibration and noise.

[0097] In the embodiments of this application, the stator-rotor clearance of the high-pressure screw pump is much smaller than that of a general screw pump. The straightness and coaxiality of the three holes are both less than or equal to 0.01 mm, avoiding vibration caused by friction between the three holes and the rotor screw due to deviations. In the embodiments of this application, a special combination broach is designed and manufactured to process the three holes of the pump body bushing. The feed amount is kept to a minimum each time, and the machining accuracy can be ensured to the greatest extent through 5 broaching operations. Afterwards, a special grinding rod and abrasive are used for 12 hours of honing operation, and finally the three holes of the pump body bushing are honed to meet the design requirements of straightness and coaxiality ≤0.01 mm, avoiding vibration caused by friction between the three holes and the rotor screw due to deviations in positional tolerances.

[0098] In some embodiments, the master and slave screws of the rotor, due to the characteristics of high-pressure screw pumps, have twice as many screw sealing stages as those of general screw pumps, resulting in a slender screw shape and a high length-to-diameter ratio. In the embodiments of this application, by employing a dedicated screw grinding machine and a special heat treatment process, two quenching treatments are performed to ensure that the straightness of the screw is less than or equal to 0.01 mm, preventing vibration caused by friction between the screw and the pump body bushing during operation.

[0099] In some embodiments, the machining error of the screw lead is controlled within ±0.015mm. In the embodiments of this application, by specially designing the machining program and process of the screw lead, the machining error of the screw lead is controlled within ±0.015mm, while avoiding interference caused by the lead error during the meshing of the master and slave screws, avoiding mutual collisions during operation, affecting the operating state, and reducing vibration and noise.

[0100] In some embodiments, the bearing 503 is installed on the main rod 201 using a bearing positioning and installation fixture 2000 to ensure that the bearing is smoothly assembled onto the drive screw. Figure 15 This is a schematic diagram of a bearing positioning and mounting fixture in some embodiments of this application. (Reference) Figure 15In some embodiments, a special bearing positioning and installation fixture is used to smoothly assemble the bearing onto the drive screw. The special bearing positioning and installation fixture 2000 includes a base frame 2001, a support frame 2002, a hydraulic sleeve 2003, and a retaining bracket 2004. The support frame is detachably mounted to the base frame; the hydraulic sleeve is positioned between the bearing and the retaining bracket, and the retaining bracket is movably connected to the base frame. In embodiments of this application, a manufacturing fixture for a hydraulic power device is provided, the manufacturing fixture including the special bearing positioning and installation fixture.

[0101] In some embodiments, the retaining bracket includes a first support column 20041, a second support column 20042, and a support plate 20043. One end of the first support column and the second support column are fixedly connected to the base frame. One end of the support plate is rotatably connected to the other end of the first support column, and the other end of the support plate is detachably connected to the other end of the second support column. The base frame includes four columns, namely a first column 20011, a second column 20012, a third column 20013, and a fourth column 20014. Support holes 20015 are respectively provided on the first column, the second column, the third column, and the fourth column. The support frame includes a first support rod 20021, a second support rod 20022, and a support truss 20023. The first support rod is detachably passed through the support holes of the first column and the second column, the second support rod is detachably passed through the support holes of the third column and the fourth column, and the support truss is detachably placed on the first support rod and the second support rod.

[0102] The bearing positioning and installation method is as follows: Place the drive screw on the base frame 2001 and support it using the support frame 2002. Then, place the bearing on the drive screw and gently press it down using the hydraulic sleeve 2003. Finally, use the retaining bracket 2004 to fix the hydraulic sleeve. Slowly rotate the stopcock of the hydraulic sleeve to use hydraulic pressure to slowly and steadily push the bearing into the drive screw until it is in place.

[0103] In the embodiments of this application, a special tooling for bearing positioning and installation is used, which can greatly reduce wear during the bearing installation process and prevent reliability and vibration noise problems caused by bearing accuracy deviations.

[0104] In an embodiment of this application, a method for manufacturing a hydraulic power device is also provided. The manufacturing method uses the manufacturing tooling described above. The manufacturing method includes: bearing positioning and installation, which includes: placing the main rod on a base frame and supporting the main rod with a support frame; placing the bearing on the main rod and gently pressing the bearing with a hydraulic sleeve; then fixing the hydraulic sleeve with a retaining bracket; slowly rotating the stopcock of the hydraulic sleeve and using hydraulic pressure to slowly and steadily push the bearing into the main rod until it is installed in place.

[0105] In some embodiments, this application provides a hydraulic power unit for an underwater vehicle. The hydraulic power unit includes a motor and a screw pump. The motor has an output shaft. The screw pump includes a stator and a rotor. The rotor includes a main rod and two driven rods. The main rod is connected to the output shaft of the motor, and the axis of the main rod is vertical. In embodiments of this application, the screw pump can be any of the screw pumps described in the above embodiments.

[0106] Figure 16 This is a schematic diagram of the high-pressure hydraulic system of an underwater vehicle in some embodiments of this application. In the embodiments of this application, the high-pressure hydraulic system of the underwater vehicle includes a hydraulic power unit. The hydraulic power unit can be the hydraulic power unit in any of the above embodiments.

[0107] For details, please refer to Figure 16 In some embodiments, the high-pressure hydraulic system of the underwater vehicle includes: a test bench 1100, a motor 1200, a booster pump 1300, a flexible connecting pipe 1400, and a rigid straight pipe 1500.

[0108] The test stand 1100 includes a horizontal fixed plate 1101; a motor 1200 is fixedly installed above the horizontal fixed plate; a booster pump 1300 is connected to the motor and located below the horizontal fixed plate. The booster pump is a positive displacement pump, which includes an oil inlet and an oil outlet; a flexible connecting pipe 1400 is connected to the oil inlet; and a rigid straight pipe 1500 is connected to the oil outlet.

[0109] In the embodiments of this application, the booster pump can be a screw pump as described in any of the above embodiments. In the embodiments of this application, a short flexible connecting pipe is added to the front end of the oil inlet to buffer the vibration caused when the medium oil is sucked into the pipeline. A rigid straight pipe is used at the front end of the oil outlet to reinforce and alleviate the pressure pulsation of the medium oil at the outlet. Both the inlet and outlet pipelines are installed on support feet using special vibration isolators.

[0110] In some embodiments, the high-pressure hydraulic system of the underwater vehicle has three high-pressure relief valves, namely a first high-pressure relief valve 1601, a second high-pressure relief valve 1602, and a third high-pressure relief valve 1603, which are connected sequentially from near to far to a rigid straight pipe. In the embodiments of this application, multiple high-pressure relief valves are added at the outlet. Multiple high-pressure relief valves can share the pressure of the entire system. Even if the pressure is increased to 20MPa, the pressure will be evenly built up on each valve. The opening degree of each valve will be much larger than that of a single high-pressure valve, avoiding the flow noise caused by the valve opening being too small. When three high-pressure relief valves use the same opening degree to jointly build up the 20MPa pressure, the vibration noise is minimized. The working state of each valve is in the optimal range, thereby reducing the vibration noise caused by excessive medium flow velocity in the system.

[0111] In some embodiments, the high-pressure hydraulic system of the underwater vehicle includes multiple rectifiers 1700 connected to the rigid straight pipe, each rectifier including a rectifier grid. In embodiments of this application, the high-pressure screw pump operates at a pressure of 20 MPa and has a high flow rate, resulting in significant impact of the internal medium flow on the pipeline. This fluid impact causes pipeline vibration, thereby increasing the vibration acceleration of the inlet and outlet flanges. In embodiments of this application, by adding a rectifier grid in the middle of the pipeline, the flow of the medium oil is made more stable, reducing the impact on the pipeline, thereby reducing the vibration acceleration on the pipeline and lowering the vibration on the inlet and outlet flanges.

[0112] Figure 17 This is a cross-sectional schematic diagram of a rectifier device in some embodiments of this application. Figure 18 This is a schematic diagram of the flow passage in some embodiments of this application. Figure 19 This is a schematic diagram of the flow orifice in some embodiments of this application. (Reference) Figures 17-19 In some embodiments, the rectifier grid includes two rectifier layers 1703, each rectifier layer having a baffle plate with a plurality of flow holes 1704. In some embodiments, the apertures of the plurality of flow holes are distributed from the center outwards in a manner that decreases in size.

[0113] In the embodiments of this application, two rectifier layers are designed inside the rectifier grid to ensure sufficient flow stabilization of the fluid within the rectifier device and reduce fluid impact on the pipeline. The flow holes in each rectifier layer are evenly distributed from the center outwards with decreasing diameter, making the fluid distribution more uniform throughout the channel and preventing fluid accumulation at individual locations, thus avoiding fluid disturbance.

[0114] refer to Figure 18 and Figure 19In some embodiments, the plurality of flow holes include a central hole 17041 and circumferential holes, which are evenly distributed around the central hole, and the angle between each circumferential hole and the axis is 3°. In some embodiments, the diameter of the central hole is 6 mm. The circumferential holes include a first circumferential hole 17402, a second circumferential hole 17403, a third circumferential hole 17404, and a fourth circumferential hole 17405 arranged radially outward from the center. The number of the first circumferential holes 17402, the second circumferential holes 17403, the third circumferential holes 17404, and the fourth circumferential holes 17405 are 6, 12, 18, and 24, respectively, and the diameters of the circumferential holes arranged radially outward from the center are 4 mm, 3 mm, 2 mm, and 1 mm, respectively. In the embodiments of this application, apart from the central 4mm flow hole, the angles of the remaining flow holes are all deviated outward by 3° (the angle is small and is not directly marked in the figure), and are inclined towards the central hole in the flow direction, so that the entire flow surface presents a "converging" shape. This design causes the fluid model to slightly form a focused shape when the medium oil flows through this place, reducing the vibration caused by the fluid impacting the pipeline outward.

[0115] In some embodiments, the rectifier has two components, namely a first rectifier 1701 and a second rectifier 1702. The first rectifier is disposed between the first high-pressure relief valve and the rigid straight pipe, and the second rectifier is disposed between the first high-pressure relief valve and the second high-pressure relief valve.

[0116] Figure 20 This is a schematic diagram of a connecting frame assembly in some embodiments of this application. (See reference...) Figure 16 and Figure 20 In some embodiments, the high-pressure hydraulic system of the underwater vehicle includes a connecting frame assembly 1800, which includes a connecting frame 1801 and a mounting plate 1802. The motor and the booster pump are connected through the connecting frame, and the connecting frame is then fixed to the mounting plate. The mounting plate is connected to a horizontal fixing plate.

[0117] In the embodiments of this application, the booster pump's connecting frame assembly consists of a connecting frame and a mounting plate. The motor and the booster pump are connected through the connecting frame, which is then fixed to the mounting plate. Finally, the entire pump unit is connected to a horizontal fixed plate through the mounting plate. In the embodiments of this application, the mounting plate of the connecting frame is the external installation interface of the low-noise high-pressure screw pump. As a major path for the transmission of vibration sources, the connecting frame can be specially designed to achieve vibration reduction optimization of the entire pump unit through vibration reduction design.

[0118] Figure 21 This is a cross-sectional schematic diagram of the connecting frame in some embodiments of this application. Figure 22 This is a schematic diagram of the connecting frame in some embodiments of this application. (See reference) Figure 21 and Figure 22In some embodiments, the connecting frame has an injection hole 1805, and the injection hole contains vibration damping rubber. The operating frequency or excitation direction of the connecting frame is different from that of the high-pressure hydraulic system of the underwater vehicle.

[0119] In the embodiments of this application, the connecting frame, as one of the main paths for vibration transmission, is designed to reduce and absorb vibration, thereby isolating some vibration energy. In the embodiments of this application, a dedicated high-damping connecting frame is designed. Specifically, flow channels are machined inside the high-damping connecting frame for injecting dedicated damping rubber. Dedicated upper and lower cover plates are designed to prevent the rubber liquid from flowing out. After the injection is completed, the damping rubber liquid is allowed to stand for 48 hours to dry completely, thus completing the manufacturing of the high-damping connecting frame.

[0120] In the embodiments of this application, modal simulation calculations were also performed on the connecting frame to ensure that the operating frequency or excitation direction of the connecting frame is different from that of the high-pressure hydraulic system of the underwater vehicle, thus avoiding resonance. Since the high-damping connecting frame is a combination of metallic and non-metallic materials, unpredictable modal changes may occur after material fusion. In the embodiments of this application, simulated modal calculations were performed. The calculations showed that the first-order mode of the system is approximately 6Hz, and the first five modes are relatively low, significantly different from the equipment's excitation frequency. The sixth-order system mode frequency exceeds 40Hz, close to the equipment's excitation frequency, but its mode shape direction is inconsistent with the excitation direction. Therefore, the entire system's modes are misaligned from the equipment's operating frequency or excitation direction, preventing system resonance and meeting the design requirements.

[0121] In some embodiments, a vibration damping ring is provided between the motor and the connecting frame, and the vibration damping ring has multiple vibration damping columns inside. The vibration damping columns are made of high-strength alloy material and are used to absorb electromagnetic vibrations generated by the motor.

[0122] In the embodiments of this application, a vibration damping ring is added between the motor and the high-damping connecting frame to achieve a further vibration reduction effect. The vibration damping ring is designed with multiple vibration damping columns inside, which are made of high-strength alloy material. It can absorb the electromagnetic vibration generated by the motor and together with the high-damping connecting frame, it forms a vibration reduction system. Through actual measurement, it can effectively reduce vibration by more than 20dB.

[0123] In the embodiments of this application, an integrated mounting plate is specifically designed, and vibration-absorbing material is added to the internal structure of the mounting plate. At the same time, while strengthening the strength of the mounting plate, a weakening design is carried out in some positions to effectively change the vibration transmission path and reduce the vibration transmitted to the outside by the high-pressure screw pump.

[0124] In some embodiments, multiple vibration isolators 1900 are disposed between the horizontal fixed plate 1101 and the mounting plate 1802. In the embodiments of this application, the vibration isolators serve as the main connecting components of the low-noise high-pressure screw pump unit, effectively blocking some vibrations. In some embodiments, eight vibration isolators 1900 are disposed between the horizontal fixed plate 1101 and the mounting plate 1802. Specifically, the angles of the eight vibration isolators 1900 are matched with the vibration energy transmission path. During installation, the installation angles of the eight vibration isolators are adjusted so that each vibration isolator can better absorb the transmitted energy.

[0125] Figure 23 This is a schematic diagram of a vibration isolator in some embodiments of this application. (Reference) Figure 23 In some embodiments, the vibration isolator 1900 is a BE type vibration isolator. In some embodiments, the levelness and compression of the vibration isolator are also adjusted to optimize its vibration reduction effect.

[0126] In the embodiments of this application, the upper and lower planes of the vibration isolator are ensured to be horizontal. Using the upper surface as a reference, the parallelism error of the lower surface does not exceed 0.02 mm, preventing vibration exceeding tolerances due to uneven equipment installation. In some embodiments, the compression of all vibration isolators is consistent, with an error not exceeding 0.2 mm. In the embodiments of this application, after removing manufacturing deviations of the vibration isolator itself and the compression caused by the pump's own weight, the preload of the fastening bolts connecting the vibration isolator to the equipment also affects its compression. To ensure consistent compression of the vibration isolators, the preload error of their fasteners cannot exceed ±5 N / m. In some embodiments, the compression of the vibration isolator is between 5 mm and 5.5 mm.

[0127] In the embodiments of this application, the ultimate compression of the vibration isolator is 8 mm, and the working compression of the vibration isolator in this application is between 5 mm and 5.5 mm, so that the working compression is 62.5% to 68.8% of the ultimate compression, and the compression of all vibration isolators is within this range with a deviation of no more than 3%. In some embodiments, the working compression of the vibration isolator is 65% of the ultimate compression.

[0128] In the embodiments of this application, the upper and lower planes of the vibration isolator are kept horizontal, with the upper surface as a reference and the parallelism error of the lower surface not exceeding 0.02mm; the compression of the multiple vibration isolators is consistent, with an error not exceeding 0.2mm; the working compression of the multiple vibration isolators is 65% of the ultimate compression, and the deviation of the working compression of all vibration isolators does not exceed 3%. By adjusting the setting angle, horizontality, and compression of the vibration isolators, their vibration reduction effect is optimized.

[0129] Figure 24 This is a partial structural schematic diagram of a high-pressure hydraulic system in some embodiments of this application. (Reference) Figure 24In some embodiments, a sleeve structure 1102 is provided between the bottom of the screw pump and the front cover. By constraining the screw pump as a whole, the addition of a sleeve structure between the bottom of the screw pump and the front cover constrains the "swinging" bottom during operation, making the overall structure more stable.

[0130] In some embodiments, the inlet flange 6013, outlet flange 1013, and sleeve structure 1102 are all fixedly constrained, forming an over-positioning installation method. In the embodiments of this application, the over-positioning installation method strengthens the rigidity of the flanges while fixing them to the sleeve structure. The inlet flange 6013 is fixed to the rear cover and sleeve structure, and the outlet flange 1013 is fixed to the pump body bushing and sleeve structure, forming an over-positioning installation method, which minimizes the impact of the "swaying" phenomenon. It is understood that in the embodiments of this application, the screw pump operates at a very high pressure, reaching 20MPa, resulting in extremely high shaft power. The matching motor power is 55kW, and the size and mass difference between the motor and the screw pump is significant, far disrupting the balance of the mass ratio. The center of gravity of the pump set is far above the horizontal fixed plate 1101. The synchronous speed of the motor is 3000r / min. In addition, the high operating pressure of the screw pump results in a relatively long pump overall, causing the screw pump to experience a significant "swaying" phenomenon when the pump set operates at high speed and high pressure. The inlet and outlet flanges of the high-pressure screw pump are also a path for external vibration transmission. Due to the very high working pressure and the "slender" shape of the screw pump to meet the high-pressure conditions, the inlet and outlet flanges vibrate significantly, especially at the bottom, where the vibration at the inlet flange is much greater than at the top. This application avoids the impact of "swaying" to the greatest extent by using an over-positioning installation method.

[0131] In some embodiments, high-damping energy-absorbing plates 1103 can be added to the inlet and outlet flanges to absorb some of the vibration energy and reduce the energy ultimately transmitted through the inlet and outlet flanges.

[0132] In some embodiments of this application, a vibration reduction method for the high-pressure hydraulic system of the underwater vehicle described in any of the above embodiments is also provided. The vibration reduction method includes at least: correcting the screw profile, wherein the correction of the screw profile includes screw profile correction for pressure pulsation and screw profile correction for screw meshing interference; adjusting the angle of the vibration isolators so that the angle of each vibration isolator matches the vibration energy transmission path; and adjusting the preload of the fasteners of the vibration isolators, wherein the error of the preload of each fastener does not exceed ±5N / m.

[0133] It is understood that the vibration reduction methods in the embodiments of this application include vibration reduction design methods, structures, features, components, etc. in any embodiment of the screw pump, hydraulic power device, or high-pressure hydraulic system.

[0134] In some embodiments of this application, a sealing vibration reduction method for a screw pump according to any of the above embodiments is also provided. The sealing vibration reduction method includes: setting the end shape of the non-working surface of the main rod to an arc shape; using a special bearing positioning and installation fixture to smoothly assemble the bearing onto the main rod; connecting a flexible pipe at the oil inlet of the screw pump; and connecting a rigid straight pipe at the oil outlet of the screw pump.

[0135] In the embodiments of this application, an underwater vehicle is also provided, the underwater vehicle including the high-pressure hydraulic system of the underwater vehicle described in any of the above embodiments.

[0136] Those skilled in the art will readily understand that the above description is merely a preferred embodiment of the present invention and is not intended to limit the present invention. Any modifications, equivalent substitutions, and improvements made within the spirit and principles of the present invention should be included within the scope of protection of the present invention.

Claims

1. A high-pressure hydraulic system for an underwater vehicle, characterized in that, The high-pressure hydraulic system of the underwater vehicle includes: The platform includes a horizontal fixing plate; The motor is fixedly installed above the horizontal fixing plate; A booster pump, connected to the motor, is located below the horizontal fixed plate. The booster pump is a positive displacement pump, including an oil inlet and an oil outlet. A flexible connector is attached to the oil inlet. A rigid straight pipe is connected to the oil outlet.

2. The high-pressure hydraulic system of the underwater vehicle according to claim 1, characterized in that, The high-pressure hydraulic system of the underwater vehicle includes multiple rectifiers connected to the rigid straight pipe, and each rectifier includes a rectifier grid.

3. The high-pressure hydraulic system of the underwater vehicle according to claim 2, characterized in that, The rectifier grid includes two rectifier layers, each of which has a baffle plate with multiple flow holes. The diameter of the multiple flow holes is distributed from the center outwards to the smallest.

4. The high-pressure hydraulic system of the underwater vehicle according to claim 3, characterized in that, The plurality of flow passages include a central hole and circumferential holes. The circumferential holes are evenly distributed around the central hole. The angle between each circumferential hole and the axis is 3°, and the holes are inclined toward the central hole in the flow direction.

5. The high-pressure hydraulic system of the underwater vehicle according to claim 2, characterized in that, The high-pressure hydraulic system of the underwater vehicle has three high-pressure relief valves, namely the first high-pressure relief valve, the second high-pressure relief valve, and the third high-pressure relief valve, which are connected to the rigid straight pipe from near to far. The rectifier has two components: a first rectifier and a second rectifier. The first rectifier is disposed between the first high-pressure relief valve and the rigid straight pipe, and the second rectifier is disposed between the first high-pressure relief valve and the second high-pressure relief valve.

6. The high-pressure hydraulic system of the underwater vehicle according to claim 1, characterized in that, The high-pressure hydraulic system of the underwater vehicle includes a connecting frame assembly, which includes a connecting frame and a mounting plate. The motor and the booster pump are connected through the connecting frame, and the connecting frame is then fixed to the mounting plate. The mounting plate is connected to a horizontal fixed plate. The connecting frame has an injection hole with vibration damping rubber inside. The operating frequency or excitation direction of the connecting frame is different from that of the high-pressure hydraulic system of the underwater vehicle.

7. The high-pressure hydraulic system of the underwater vehicle according to claim 6, characterized in that, Multiple vibration isolators are installed between the horizontal fixed plate and the mounting plate; The upper and lower planes of the vibration isolator must be level. With the upper surface as the reference, the parallelism error of the lower surface shall not exceed 0.02mm. The compression of the multiple vibration isolators is consistent, with an error not exceeding 0.2 mm; The working compression of the plurality of vibration isolators is 65% of the ultimate compression, and the deviation of the working compression of all vibration isolators does not exceed 3%.

8. The high-pressure hydraulic system of the underwater vehicle according to any one of claims 1-7, characterized in that, The booster pump is a screw pump, which includes an oil return device and sealing components, a pump body bushing, a driven rod bushing, a front cover, a sealing cover, and a main rod. The oil return device includes an oil return valve, which is located at one end of the oil inlet of the screw pump; The sealing assembly includes a first sealing ring, a second sealing ring, a third sealing ring, and a mechanical seal assembly. The first sealing ring is disposed between the pump body bushing and the front cover, the second sealing ring is disposed between the driven rod bushing and the front cover, the third sealing ring is disposed between the front cover and the sealing cover, and the mechanical seal assembly is disposed between the sealing cover and the main rod. The first sealing ring is elliptical.

9. A vibration reduction method for a high-pressure hydraulic system of an underwater vehicle according to any one of claims 1-8, characterized in that, The vibration reduction method includes: The screw profile is modified, including screw profile modification for pressure pulsation and screw profile modification for screw meshing interference; the screw profile modification for screw meshing interference includes setting the end shape of the non-working surface of the main rod to an arc shape. Adjust the angle of the vibration isolators so that the angle of each vibration isolator matches the vibration energy transmission path; Adjust the preload of the fasteners on the vibration isolator, ensuring that the error in the preload of each fastener does not exceed 5 N / m.

10. An underwater vehicle, characterized in that, The underwater vehicle includes the high-pressure hydraulic system of the underwater vehicle according to any one of claims 1-8.

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