Optimization design method for rotating assembly of high-speed high-pressure axial plunger pump

Abstract

The invention discloses an optimal design method for a rotating assembly of a high-speed and high-pressure axial plunger pump. The method comprises the specific steps that S1, a design target and the displacement, the rotating speed and the rated pressure of the plunger pump are input; s2, plunger pump rotating assembly structure parameters are designed based on the target displacement, and feedback optimization is carried out; s3, establishing a three-dimensional model of the rotating assembly, judging the strength of parts under rated pressure based on a structural deformation principle, and performing feedback optimization; s4, judging and optimizing sliding shoe overturning, cylinder overturning and cavitation and cavitation faults in the high-speed process of the rotating assembly based on speed limitation; and S5, the plunger pump rotating assembly structure parameters and part materials and processes are output. According to the method, rapid optimization design of the rotating assembly of the plunger pump can be achieved from the design objective and fault judgment optimization of the plunger pump, high-speed and high-pressure failure of the plunger pump in the operation process is prevented, the design period of the plunger pump is shortened, and the operation reliability of the plunger pump is improved.

Description

Technical Field

[0001] The present invention relates to the technical field of axial piston pumps, and specifically to an optimized design method for the rotating assembly of a high-speed and high-pressure axial piston pump. Background Art

[0002] As a power element of a fluid system, hydraulic pumps are widely used in fields such as aerospace, construction machinery, and national defense. With the development of high-end equipment manufacturing, piston pumps are gradually moving towards higher pressures and higher speeds to obtain higher system efficiency and power-to-weight ratios.

[0003] As the core component of a piston pump, the rotating assembly is the main location where failures occur. Under high-pressure conditions, the rotating assembly is affected by the high pressure of the fluid and is prone to component fracture or deformation failure. Under high-speed conditions, there are failure risks such as slipper tilting, cylinder block tilting, and cavitation erosion in piston pumps. Existing design methods are all designed for single elements or based on empirical formulas, resulting in problems such as design redundancy and long design cycles. Summary of the Invention

[0004] Aiming at the problems existing in the prior art, the present invention provides an optimized design method for the rotating assembly of a high-speed and high-pressure axial piston pump. Input the displacement, speed, and rated pressure of the expected piston pump, and design the basic structural parameters of the piston pump based on the displacement limit. Determine the safety and reliability of components under the rated pressure condition through finite element simulation analysis, and optimize the structural and material process parameters of the rotating assembly. Based on the speed limit, parameter judgment and structural optimization are carried out respectively for slipper tilting failure, cylinder block tilting failure, and cavitation erosion failure. When the requirements of displacement limit, pressure limit, and speed limit are met, the structural parameters of the piston pump rotating assembly, as well as the materials and processes of components, are finally output.

[0005] The present invention provides an optimized design method for the rotating assembly of a high-speed and high-pressure axial piston pump, which specifically includes the following steps:

[0006] S1. Determine the design objectives and initial structural material characteristics of the piston pump rotating assembly. The design objectives include the maximum target displacement, maximum speed, and rated pressure of the piston pump;

[0007] S2. Determine the structural parameters of the piston pump rotating assembly based on the maximum target displacement;

[0008] S3. Establish a three-dimensional model of the rotating assembly according to the structural parameters of the piston pump rotating assembly obtained in step S2, set the initial material properties and processes of components, judge the strength of components under the rated pressure condition of the piston pump based on the principle of structural deformation, and determine whether it meets the rated pressure. If it meets, proceed to step S4; otherwise, optimize the parameters and then re-execute steps S2 - S3;

[0009] S4. Based on the structural parameters of the plunger pump rotating assembly obtained in step S2, parameter judgment and optimization are respectively carried out on the faults occurring during the operation of the rotating assembly under the rated speed condition, and it is judged whether the maximum speed requirement is met. If it is met, step S5 is carried out; otherwise, the parameters are optimized and steps S2 - S4 are executed again.

[0010] S5. Output the optimized structural parameters of the plunger pump rotating assembly, as well as the material properties and processes of the components.

[0011] Preferably, the structural parameters of the plunger pump rotating assembly in step S2 include cylinder block parameters, plunger parameters, slipper parameters, return spring disk parameters, port plate parameters, and swash plate parameters.

[0012] Preferably, the specific steps of step S2 include:

[0013] S21. Determine the cylinder block parameters, which include the diameter of the plunger hole, the distribution circle diameter, the wall thickness between the cylinder block plunger holes, and the cylinder block length.

[0014] Among them, the relationship between the diameter of the plunger hole, the distribution circle diameter, and the maximum target displacement is as follows:

[0015]

[0016] Among them, V g is the designed flow rate of the plunger pump, z is the number of plungers, d k is the plunger diameter, D T is the plunger distribution circle diameter, and β is the swash plate angle;

[0017] Judge whether the wall thickness between the cylinder block plunger holes meets the requirements by calculating the average normal stress received by the cylinder block plunger holes. The calculation formula for the average normal stress received by the cylinder block plunger holes is as follows:

[0018]

[0019] Among them, σ is the average normal stress of the cylinder block plunger holes, is the pressure in the plunger cavity, ε is the resultant force angle between the plunger holes, D T is the plunger distribution circle diameter, d k is the plunger diameter, and z is the number of plungers;

[0020] The cylinder block length satisfies the following formula:

[0021] L c = l k + 2R tanβ + (2 - 3) + s + L h

[0022] Among them, L c is the cylinder block length, l k is the minimum contact length of the plunger, dk d is the plunger diameter, R is the radius of the plunger distribution circle, β is the swash plate angle, s is the length of the port plate hole connecting the plunger hole, i.e., the cylinder bottom thickness, L h is the spline protrusion length of the cylinder block;

[0023] S22. Determine the plunger parameters, which include the plunger length and the plunger ball head diameter;

[0024] The plunger length satisfies the following formula:

[0025] L z = l 0 + h p + l k

[0026] where, L z is the plunger length, l 0 is the minimum outer extension length, take l 0 = 0.2d k h p is the plunger stroke, h p = 2R tanβ, l k is the minimum engagement length, l k = (2 - 2.5)d k ;

[0027] The plunger ball head diameter d 2 = (0.7 - 0.8)d k ;

[0028] S23. Determine the swash plate parameters, which include the swash plate ball socket diameter, the inner diameter of the swash plate sealing band, the outer diameter of the swash plate sealing band, the outer diameter of the swash plate ball cup, and the diameter of the swash plate damping hole;

[0029] The swash plate ball socket diameter d 8 is equal to the plunger ball head diameter d 2 ;

[0030] The inner diameter of the swash plate sealing band satisfies the following formula:

[0031]

[0032] where, d 4 is the inner diameter of the swash plate sealing band, k is the sealing coefficient, d k is the plunger diameter, β is the swash plate angle, d′ 6 is the outer diameter of the swash plate sealing band, and the outer diameter of the swash plate sealing band satisfies s h = (0.2 - 1)mm;

[0033] The outer diameter of the swash plate ball cup d 7 = (0.95 - 1)d k ;

[0034] Diameter d of the damping hole of the slipper 3 It is calculated by the method of residual clamping force;

[0035] S24. Determine the parameters of the return disk. The parameters of the return disk include the distribution circle diameter of the return disk and the diameter of the return disk hole;

[0036] The calculation formula for the distribution circle diameter of the return disk is as follows:

[0037]

[0038] where, D 3 is the distribution circle diameter of the return disk, R is the distribution circle radius of the plunger, and β is the swash plate angle;

[0039] The aperture of the return disk satisfies the following formula:

[0040] d’ 5 = 2|ε max | + d 7 + 2a min

[0041] where, d’ 5 is the aperture of the return disk, |ε max | is the radial deviation between the elliptical trajectory of the slipper center and the distribution circle of the return disk, d 7 is the outer diameter of the ball cup of the slipper, a min is the clearance between the neck of the slipper and the hole of the return disk, and the outer diameter d of the slipper 6 ≥ d’ 5 , and the outer diameter d of the slipper 6 = d′ 6 + (1 - 3);

[0042] S25. Determine the parameters of the valve plate and verify whether the valve plate meets the requirements. The parameters of the valve plate include the size of the sealing band of the valve plate pair and the deflection angle;

[0043] The width of the inner and outer sealing bands of the valve plate pair is (0.18 - 0.2)d k ;

[0044] When the pressure rises from the initial pressure to the pre-boost pressure, the volume compression is calculated by the following formula:

[0045]

[0046] where, ΔV is the volume compression, p d is the pre-boost pressure, p 0 is the initial pressure, E 0 is the elastic modulus of the oil, V 0 is the closed volume of the plunger cavity at the bottom dead center position, d k is the plunger diameter, h maxis the maximum stroke of the plunger;

[0047] The swash plate deflection angle is calculated by the following formula:

[0048]

[0049] where ΔV is the volume compression, d k is the plunger diameter, R is the radius of the plunger distribution circle, Δh is the plunger stroke, β is the swash plate angle, is the swash plate deflection angle;

[0050] The ratio of the radius of the center of the kidney-shaped groove of the cylinder block and the swash plate to the radius of the plunger distribution circle is 0.7 - 1.0, the width of the kidney-shaped groove is 0.35 - 0.5 of the plunger diameter, and the width of the inner and outer sealing bands of the swash plate is 0.1 - 0.2 of the plunger diameter;

[0051] It is judged whether the swash plate meets the requirements by calculating the linear velocity at the center of the kidney-shaped groove of the cylinder block and the swash plate. The calculation formula is as follows:

[0052]

[0053] where V R is the linear velocity of the center line of the kidney-shaped groove of the swash plate, R y is the radius of the center of the kidney-shaped groove of the swash plate, n max is the maximum rotational speed of the plunger pump;

[0054] S26. Determine the swash plate parameters, which include the swash plate angle and the minimum allowable distance between the two chamfered planes;

[0055] The swash plate angle β is 15° - 22°;

[0056] The minimum allowable distance between the two chamfered planes satisfies the following formula:

[0057] B spmin = D + d’ 6 + 2A min

[0058] where B spmin is the minimum allowable distance between the two chamfered planes, D is the major axis of the elliptical trajectory of the slipper on the swash plate, d’ 6 is the outer diameter of the slipper sealing band, A min is the minimum distance from the edge of the slipper to the edge of the swash plate;

[0059] S27. Calculate the designed displacement according to the structural parameters, and judge whether the designed displacement is greater than the target displacement. If it is greater, execute the next step. If it is less than or equal, perform the first parameter optimization and then return to step S21 and re-execute step S2;

[0060] S28. Determine whether the design displacement is less than 1.1 times the target displacement. If it is less, proceed to the next step; otherwise, after optimizing the second parameter, return to step S21 and re - execute step S2.

[0061] Preferably, the specific steps of step S3 include:

[0062] S31. Establish a three - dimensional model of the rotating component, and initially set the materials and processes of the components.

[0063] S32. Based on the principle of structural deformation, judge the structural strength of the components under the rated pressure. Specifically, it includes the following sub - steps:

[0064] S321. Based on the structure of the rotating - component parts, divide the solid - domain mesh for finite - element analysis.

[0065] S322. The structural deformation and stress of the rotating component are affected by the temperature field and fluid field during the operation of the piston pump and are transmitted between structures. Apply the fluid - field pressure load to the parts of the rotating component in contact with the fluid, apply the temperature load during the operation of the piston pump to the rotating component, and impose position constraints on the rotating component according to the actual operating conditions.

[0066] S323. Conduct a structural - deformation analysis of the rotating component under multi - field coupling, including structural deformation and stress under temperature load and fluid load.

[0067] The force - deformation of the rotating component satisfies the following formula:

[0068] [K]×{δ}={F}

[0069] where [K] is the system stiffness matrix of the rotating component, {δ} is the system node - displacement matrix of the rotating component, and {F} is the system force matrix of the rotating component.

[0070] The thermal deformation of the rotating component satisfies the following formula:

[0071]

[0072] where f T is the structural deformation, α T is the thermal expansion coefficient of the rotating - component material, is the temperature difference;

[0073] S33. Extract the structural - deformation and stress - strain nephograms of the rotating component. If there is no interference in the oil - film clearance of the rotating component after structural deformation and the structural stress is less than the allowable stress of the material, then judge whether the maximum pressure under the current allowable stress of the material is greater than the rated pressure. If it is, proceed to the next step; otherwise, after optimizing the third parameter, re - execute step S2 and step S3.

[0074] S34. Calculate whether the maximum pressure under the required stress of the current material is less than 1.1 times the rated pressure. If so, proceed to the next step; otherwise, after optimizing the fourth parameter, re-execute steps S2 and S3.

[0075] Preferably, if the rated pressure structure is still not satisfied after step S33 is executed n times, output that the rated pressure requirement cannot be met under the current parameters.

[0076] Preferably, step S4 specifically includes the following sub-steps:

[0077] S41. Construct a formula for calculating the maximum speed before the slipper overturns to verify the slipper overturning failure of the rotating assembly:

[0078]

[0079] where ω is the rotational angular velocity of the piston pump, M s is the mass of a single slipper, X s is the distance from the center of the slipper ball head to the centroid, R s is the radius of the slipper end face, M p is the mass of a single piston, β is the swash plate angle, R is the piston distribution circle radius, F sp is the spring force;

[0080] If the rotational angular velocity of the piston pump is less than or equal to the maximum speed before the slipper overturns, proceed to the next step and output the maximum rotational speed of the rotating assembly before the slipper overturning failure as n hxqf , otherwise optimize the spring force, the radius of the slipper end face, the mass of the slipper, the mass of the piston, and the distance from the slipper ball head to the centroid; after optimization, re-execute step S41. If it is still not satisfied after n optimizations, continue to the next step;

[0081] S42. Construct a formula for calculating the maximum speed before the cylinder block overturns to verify the cylinder block overturning failure of the rotating assembly:

[0082]

[0083] where ω is the rotational angular velocity of the cylinder block, R g is the outer radius of the cylinder block, N is the number of pistons, M p is the mass of a single piston, M s is the mass of a single slipper, R is the piston distribution circle radius, β is the swash plate angle;

[0084] If the rotational angular velocity of the cylinder block is less than or equal to the maximum speed before the slipper overturns, proceed to the next step and output the maximum rotational speed of the rotating assembly before the cylinder block overturning failure as n gtqf, otherwise, optimize the spring force, cylinder radius, slipper mass, and plunger mass; after optimization, re - execute step S42. If it still does not meet the requirements after n optimizations, continue to execute the next step;

[0085] S43. Establish the maximum speed calculation formula before the cavitation and erosion failure of the plunger pump to verify the cavitation and erosion failure of the rotating assembly:

[0086]

[0087] Among them, ω is the rotational angular velocity of the plunger pump, P i is the pressure in the kidney - shaped interaction area, P cav is the pressure in the plunger core cavity, A p is the plunger area, A k is the area of the kidney - shaped interaction area, R is the radius of the plunger distribution circle, β is the swash - plate angle, A i is the inlet area, N is the number of plungers, and ρ is the medium cavitation pressure;

[0088] If the rotational angular velocity of the plunger pump is less than or equal to the maximum speed before cavitation and erosion, then execute the next step and output the maximum rotational speed of the rotating assembly before the cavitation and erosion of the plunger pump as n khqs , otherwise, optimize the suction pressure, medium cavitation pressure, inlet area, and the area of the kidney - shaped interaction area, and then re - execute step S43. If it still does not meet the requirements after n optimizations, continue to execute the next step;

[0089] S44. Judge whether the maximum rotational speed n pmax of the plunger pump meets the requirements. If it meets the requirements, execute the next step. If it does not meet the requirements, perform the fifth - parameter optimization and then re - execute steps S2 - S4. If it still does not meet the requirements after executing step S44 n times, output the maximum rotational speed of the plunger pump as n pmax ; n pmax The calculation formula of is:

[0090] n pmax = min(n hxqf , n gtqf , n khqs ).

[0091] Preferably, in step S432, establish the basic control equation of the fluid domain of the rotating assembly according to the basic laws of fluid transmission in the rotating assembly of the plunger pump.

[0092] Preferably, step S432 specifically includes the following sub - steps:

[0093] S4321. Construct the specific expression of the mass - conservation equation of the rotating assembly;

[0094] S4322. Establish the compressible fluid Navier - Stokes equation of the rotating assembly;

[0095] S4323. Apply the load characteristics and motion characteristics during the use of the plunger pump to the basic control equations of the fluid domain of the rotating assembly based on the design pressure and design speed of the plunger pump;

[0096] S4324. Obtain the gas volume fraction contour map of the fluid domain of the rotating assembly through finite element analysis to characterize the cavitation and erosion characteristics of the rotating assembly.

[0097] Preferably, in step S24, the radial deviation |ε max | between the elliptical trajectory of the center of the slipper and the distribution circle of the return disk is:

[0098]

[0099] where |ε max | is the radial deviation between the elliptical trajectory of the center of the slipper and the distribution circle of the return disk, R is the radius of the plunger distribution circle, and β is the swash plate angle.

[0100] Preferably, the specific steps of the first parameter optimization in step S27 are to increase the number of plungers, the plunger diameter, the plunger distribution circle diameter, and the swash plate angle;

[0101] The specific steps of the second parameter optimization in step S28 are to reduce the number of plungers, the plunger diameter, the plunger distribution circle diameter, and the swash plate angle;

[0102] The specific steps of the third parameter optimization in step S33 are to increase the structural strength and stiffness of the material, reduce the thermal expansion coefficient of the material, and perform wall thickness enhancement for the area that does not meet the rated pressure;

[0103] The specific steps of the fourth parameter optimization in step S34 are to reduce the structural strength and stiffness of the overall material;

[0104] The specific steps of the fifth parameter optimization in step S44 are:

[0105] If the maximum speed of the plunger pump is n pmax = n hxqf , then under the basic displacement limit formula, reduce the number of plungers, lower the swash plate angle, and reduce the plunger distribution circle radius;

[0106] If the maximum speed of the plunger pump is n pmax = n gtqf , then under the basic displacement limit formula, reduce the number of plungers, lower the swash plate angle, reduce the plunger distribution circle radius, and increase the cylinder block radius;

[0107] If the maximum speed of the plunger pump is n pmax = n khqs, under the basic displacement limit formula, increase the number of pistons, reduce the swash plate angle, and reduce the piston distribution circle radius, where the basic displacement limit formula is the relationship between the piston hole diameter, the distribution circle diameter, and the maximum target displacement in step S21.

[0108] Compared with the prior art, the present invention has the following advantages:

[0109] (1) The present invention comprehensively considers the structural design and optimization of the rotating assembly of the piston pump under the conditions of displacement limit, pressure limit, and speed limit, can accurately obtain the deformation and stress nephograms of key components, the cavitation and erosion nephogram of the internal flow field of the rotating assembly, combines with the critical condition theoretical formula of the piston pump, discriminates whether the design parameters meet the expected requirements, and optimizes and improves through multiple closed-loop feedbacks to obtain the optimal piston pump parameters that meet the requirements.

[0110] (2) Based on the numerical simulation and experimental results obtained from the optimization design method of the rotating assembly of the high-speed and high-pressure axial piston pump, the present invention improves and optimizes the rotating assembly of the piston pump multiple times, and obtains reliable operating component design parameters and material processes through experimental comparison. Therefore, the present invention is an analysis method with high accuracy, low calculation cost, short R & D cycle and in line with engineering practice established through a large number of numerical simulation calculations and experiments, and can be popularized and applied in various application scenarios. Description of the Drawings

[0111] Figure 1 is the flow chart of the optimization design method of the rotating assembly of the high-speed and high-pressure axial piston pump of the present invention;

[0112] Figure 2 is the three-dimensional structure diagram of the rotating assembly in the optimization design method of the rotating assembly of the high-speed and high-pressure axial piston pump of the present invention;

[0113] Figure 3 is the deformation and stress nephogram in the optimization design method of the rotating assembly of the high-speed and high-pressure axial piston pump of the present invention;

[0114] Figure 4 is the schematic diagram of the fluid domain in the optimization design method of the rotating assembly of the high-speed and high-pressure axial piston pump of the present invention;

[0115] Figure 5 is the cavitation and erosion nephogram in the optimization design method of the rotating assembly of the high-speed and high-pressure axial piston pump of the present invention;

[0116] Figure 6 is the curve graph of the cavitation and erosion speed limit varying with the suction pressure in the optimization design method of the rotating assembly of the high-speed and high-pressure axial piston pump of the present invention. Detailed Embodiments

[0117] To elaborate on the technical content, achieved objectives, and effects of the present invention, the following will be a detailed description in conjunction with the accompanying drawings of the specification.

[0118] The present invention provides an optimized design method for the rotating assembly of a high-speed high-pressure axial piston pump. As Figure 1 shown, it specifically includes the following steps:

[0119] S1. Determine the design objectives and initial structural material properties of the piston pump rotating assembly. The design objectives include the maximum target displacement, maximum rotational speed, and rated pressure of the piston pump.

[0120] S2. Determine the structural parameters of the piston pump rotating assembly based on the maximum target displacement. Specifically, it includes the following sub-steps:

[0121] S21. Determine the cylinder block parameters, which include the diameter of the piston hole, the distribution circle diameter, the wall thickness between the piston holes in the cylinder block, and the length of the cylinder block.

[0122] During the design process of the cylinder block parameters, first, the design requirements of the piston pump displacement need to be met. The relationship between the piston diameter, the piston distribution circle diameter, and the pump displacement is as follows:

[0123]

[0124] Where V g is the designed flow rate of the piston pump, z is the number of pistons, d k is the piston diameter, D T is the piston distribution circle diameter, and β is the swash plate angle.

[0125] At the same time, during the design process of the cylinder block parameters, the thin-wall strength between the piston holes in the cylinder block and between the cylinder block holes and the inner and outer wall surfaces needs to be considered. Through the equivalent conversion of the rated fluid pressure of the piston pump, the average normal stress received by the piston holes in the cylinder block satisfies the following formula:

[0126]

[0127] Where σ is the average normal stress of the piston holes in the cylinder block, σ should not exceed the maximum allowable stress of the material, p is the pressure in the piston cavity, ε is the resultant force angle between the piston holes, D T is the piston distribution circle diameter, d k is the piston diameter, and z is the number of pistons.

[0128] During the design process, the number of pistons and the swash plate angle are calculated using the initial design values. The piston diameter and the piston distribution circle diameter are obtained through the strength calculation of the cylinder block material, and further, the wall thickness between the piston holes in the cylinder block is obtained.

[0129] The cylinder block length is an important design parameter of the cylinder block. The cylinder block length mainly consists of the minimum engagement length of the plunger, the plunger stroke, the safety distance, and the protruding part of the spline of the cylinder block. The cylinder block length should satisfy the following formula:

[0130] L c =l k +2R tanβ+(2 - 3)+s+L h

[0131] Among them, L c is the cylinder block length, l k is the minimum engagement length of the plunger. When the working pressure is greater than 30 MPa, l k =(2 - 2.5)d k , d k is the plunger diameter, R is the radius of the distribution circle of the plungers, β is the swash plate angle. To prevent the plunger from colliding with the cylinder block, a certain length is reserved, generally 2 - 3 mm. s is the length of the port plate hole connecting the plunger hole, that is, the cylinder bottom thickness, and generally s=(0.4 - 0.6)d k , L h is the protruding length of the spline of the cylinder block.

[0132] S22. Determine the plunger parameters. The plunger parameters include the plunger length and the plunger ball head diameter.

[0133] The plunger length mainly consists of the minimum extension length, the plunger stroke, and the minimum engagement length. The plunger length should satisfy the following formula:

[0134] L z =l 0 +h p +l k

[0135] Among them, L z is the plunger length, l 0 is the minimum extension length, and generally l 0 =0.2d k , h p is the plunger stroke, h p =2Rtanβ, l k is the minimum engagement length, l k =(2 - 2.5)d k .

[0136] The plunger ball head diameter d 2 =(0.7 - 0.8)d k .

[0137] In addition, the diameter d of the plunger damping hole 5=(0.5 - 2) mm. Since the slipper is designed using the residual clamping force method, the damping hole does not play any damping role. Therefore, the damping hole of the plunger can be appropriately made larger.

[0138] S23. Determine the slipper parameters, which include the slipper ball socket diameter, the inner diameter of the slipper sealing band, the outer diameter of the slipper sealing band, the outer diameter of the slipper ball cup, and the diameter of the slipper damping hole.

[0139] The slipper ball socket diameter d 8 is equal to the plunger ball head diameter d 2 .

[0140] The inner diameter of the slipper sealing band should satisfy the following formula:

[0141]

[0142] where d 4 is the inner diameter of the slipper sealing band, k is the sealing coefficient, d k is the plunger diameter, β is the swash plate angle, d′ 6 is the outer diameter of the slipper sealing band and should satisfy To ensure there is a certain gap between the slippers, take s h =(0.2 - 1) mm.

[0143] The outer diameter of the slipper d 6 = d′ 6 +(1 - 3), the outer diameter of the slipper ball cup d 7 =(0.95 - 1)d k .

[0144] The diameter of the slipper damping hole d 3 is calculated by the residual clamping force method. The specific calculation process is as follows:

[0145] The clamping force on the slipper is composed of the inertial force of the plunger and slipper, the pressure in the plunger chamber, and the pre - clamping force of the central spring, as shown below:

[0146]

[0147] where F N is the clamping force on the slipper, F p is the plunger chamber pressure, F a is the inertial force of the plunger and slipper, F s is the pre - clamping force of the central spring on a single slipper, and β is the swash plate angle.

[0148] The flow rate through the damping hole is expressed by the following formula:

[0149]

[0150] where Q sis the damping hole flow rate, p p is the plunger chamber pressure, p r is the central oil chamber pressure, C op 、C os are the flow coefficient of the damping holes of the plunger and the slipper respectively, and the calculation formulas are as follows:

[0151]

[0152] Among them, d op is the diameter of the damping hole of the plunger, l op is the length of the damping hole of the plunger, d os is the diameter of the damping hole of the slipper, l os is the length of the damping hole of the slipper.

[0153] A parallel oil film is formed between the slipper and the swash plate. According to its pressure distribution, the oil film support force is:

[0154]

[0155] Among them, F r is the oil film support force, r s is the inner diameter of the sealing belt of the slipper, R s is the outer diameter of the sealing belt of the slipper, p r is the central oil chamber pressure.

[0156] The calculation formula for the flow rate of the parallel disk gap is as follows:

[0157]

[0158] Among them, Q r is the flow rate of the parallel disk gap, h is the oil film thickness, and μ is the dynamic viscosity of the oil.

[0159] According to the flow conservation, the flow rate of the fixed damper is equal to the flow rate of the parallel disk gap. Based on the remaining compression coefficient, the force balance of the slipper is listed:

[0160]

[0161] Among them, Ψ is the remaining compression force coefficient, and generally Ψ = 1 to 1.05 is taken.

[0162] According to the above and the reasonable range of the oil film thickness, the diameter d of the damping hole of the slipper can be determined 3 .

[0163] S24. Determine the return disk parameters. The return disk parameters include the distribution circle diameter of the return disk and the diameter of the return disk hole.

[0164] When the swash plate is tilted, the slipper makes an elliptical motion on the swash plate. To ensure that the pump does not interfere with the return disk during operation, the distribution circle diameter of the return disk is taken as the average value of the major and minor axes of the ellipse, as shown in the following formula:

[0165]

[0166] Among them, D 3 is the distribution circle diameter of the return disk, R is the distribution circle radius of the plungers, and β is the swash plate angle.

[0167] The radial deviation between the elliptical locus of the swashplate shoe center and the distribution circle of the return disk is:

[0168]

[0169] Among them, |ε max | is the radial deviation between the elliptical locus of the swashplate shoe center and the distribution circle of the return disk, R is the distribution circle radius of the plungers, and β is the swash plate angle.

[0170] Therefore, the return disk aperture should satisfy the following formula:

[0171] d’ 5 = 2|ε max | + d 7 + 2a min

[0172] Among them, d’ 5 is the return disk aperture, |ε max | is the radial deviation between the elliptical locus of the swashplate shoe center and the distribution circle of the return disk, d 7 is the outer diameter of the spherical cup of the swashplate shoe, a min is the clearance between the swashplate shoe neck and the return disk hole. Generally, a min = 0.2 - 1 mm can be taken. At the maximum locus deviation, in order to make the maximum outer diameter of the swashplate shoe head have a certain overlap with the return disk, the outer diameter d 6 ≥ d’ 5 .

[0173] S25. Determine the port plate parameters and verify whether the port plate meets the requirements. The port plate parameters include the port plate pair seal band size and the deflection angle.

[0174] The radius of the center of the kidney-shaped groove of the cylinder block and the port plate should not be greater than the distribution circle radius of the plungers. The ratio between the two can be between 0.7 and 1.0. The width of the kidney-shaped groove is 0.35 - 0.5 of the plunger diameter, and the width of the inner and outer seal bands of the port plate is selected as 0.1 - 0.2 of the plunger diameter.

[0175] After determining the radius of the center of the kidney-shaped groove of the cylinder block and the port plate, it should be judged whether the linear velocity at this center meets the use range. The calculation formula is as follows:

[0176]

[0177] Among them, V Ris the center line speed of the kidney-shaped groove on the port plate, R y is the center radius of the kidney-shaped groove on the port plate, n max is the maximum rotational speed of the piston pump.

[0178] The width range of the inner and outer sealing bands is (0.18 - 0.2)d k .

[0179] To effectively reduce pressure shock and lower the noise of the piston pump, the kidney-shaped window of the port plate is usually deflected to achieve pre-boost and pre-drop of pressure.

[0180] When the pressure rises from the initial pressure to the pre-boost pressure, the volume compression is calculated by the following formula:

[0181]

[0182] where, ΔV is the volume compression, p d is the pre-boost pressure, p 0 is the initial pressure, E 0 is the elastic modulus of the oil, V 0 is the closed volume of the piston cavity at the bottom dead center position, d k is the piston diameter, h max is the maximum stroke of the piston.

[0183] To make the necessary stroke of the piston when the compression volume is ΔV be Δh, the following formula can be obtained:

[0184]

[0185] where, ΔV is the volume compression, d k is the piston diameter, R is the piston distribution circle radius, Δh is the piston stroke, β is the swash plate angle, is the deflection angle of the port plate.

[0186] Combining the above formulas, the design value of the port plate deflection angle can be obtained.

[0187] S26. Design of the swash plate parameters, including the swash plate angle and the minimum allowable distance between the two chamfered planes.

[0188] The swash plate angle is calculated and optimized based on the initial design value. The swash plate angle β is generally 15° - 22°.

[0189] The minimum allowable distance between the two chamfered planes should satisfy the following formula:

[0190] B spmin = D + d’ 6 + 2A min

[0191] where, B spminis the minimum allowable distance between two chamfered planes, D is the major axis of the elliptical trajectory of the slipper on the swash plate, d’ 6 is the outer diameter of the slipper sealing band, A min is the minimum distance from the edge of the slipper to the edge of the swash plate (or pressure plate), generally taken as 0.2 - 1 mm.

[0192] S27. Calculate the design displacement according to the structural parameters, and judge whether the design displacement is greater than the target displacement. If it is greater, execute the next step. If it is less than or equal, perform the first parameter optimization and then return to step S21 and execute step S2 again.

[0193] S28. Judge whether the design displacement is less than 1.1 times the target displacement. If it is less, execute the next step. If not, perform the second parameter optimization and then return to step S21 and execute step S2 again.

[0194] The specific steps of the first parameter optimization are to increase the number of pistons, piston diameter, piston distribution circle diameter, and swash plate angle. The specific steps of the second parameter optimization are to decrease the number of pistons, piston diameter, piston distribution circle diameter, and swash plate angle.

[0195] S3. Establish a three-dimensional model of the rotating assembly according to the structural parameters of the piston pump rotating assembly obtained in step S2, set the initial material properties and processes of the components, judge the strength of the components under the rated pressure condition of the piston pump based on the principle of structural deformation, and judge whether it meets the rated pressure. If it meets, proceed to step S4. Otherwise, optimize the parameters and then re-execute steps S2 - S3. Specifically include:

[0196] S31. Establish a three-dimensional model of the rotating assembly, as Figure 2 shown, initially set the component materials and processes.

[0197] S32. Judge the structural strength of the components under the rated pressure based on the principle of structural deformation, specifically including the following sub-steps:

[0198] S321. Based on the structure of the rotating assembly components, divide the solid domain mesh for finite element analysis.

[0199] S322. The structural deformation and stress of the rotating assembly are affected by the temperature field and fluid field during the operation of the piston pump and are transmitted between the structures. Apply the fluid field pressure load to the parts of the rotating assembly in contact with the fluid, apply the temperature field during the operation of the piston pump to the rotating assembly, and apply position constraints to the rotating assembly according to the actual operating conditions.

[0200] S323. Conduct a structural deformation analysis of the rotating assembly under the multi-field coupling effect, including the structural deformation and stress under temperature load and fluid load.

[0201] The force deformation of the rotating assembly satisfies the following formula:

[0202] [K]×{δ} = {F}

[0203] Among them, [K] is the stiffness matrix of the rotating component system, {δ} is the nodal displacement matrix of the rotating component system, and {F} is the force matrix of the rotating component system.

[0204] The thermal deformation term of the rotating component satisfies the following equation:

[0205]

[0206] Among them, f T is the structural deformation, α T is the thermal expansion coefficient of the rotating component material, is the temperature difference.

[0207] S324. Extract the structural deformation and stress-strain nephogram of the rotating component. As Figure 3 shown, if there is no interference in the oil film clearance of the rotating component after structural deformation, it is considered that the plunger pump can operate normally. If the structural stress is less than the allowable stress of the material, it is considered that the structural components are reliable. Otherwise, re-execute S2 and S3.

[0208] S33. Extract the structural deformation and stress-strain nephogram of the rotating component. Judge that there is no interference in the oil film clearance of the rotating component after structural deformation and the structural stress is less than the allowable stress of the material, and then calculate whether the maximum pressure under the current allowable stress of the material is greater than the rated pressure. If so, execute the next step. If not, re-execute step S2 and step S3 after optimizing the third parameter. If it still does not meet the rated pressure structure after step S33 is executed n times, output that the rated pressure requirement cannot be met under the current parameters.

[0209] S34. Calculate whether the maximum pressure under the current allowable stress of the material is less than 1.1 times the rated pressure. If so, execute the next step. Otherwise, re-execute step S2 and step S3 after optimizing the fourth parameter.

[0210] The specific steps of the third parameter optimization are to increase the structural strength and stiffness of the material in the area that does not meet the rated pressure, reduce the thermal expansion coefficient of the material, and perform wall thickness enhancement. The specific steps of the fourth parameter optimization are to reduce the structural strength and stiffness of the overall material.

[0211] S4. According to the structural parameters of the plunger pump rotating component obtained in step S2, respectively perform parameter judgment and optimization on the faults that occur during the operation of the rotating component under the rated speed condition, and judge whether it meets the maximum speed requirement. If it meets, perform step S5. If it does not meet, perform parameter optimization and then re-execute steps S2 - S4.

[0212] The specific steps are as follows:

[0213] S41. The slipper is a mechanism connecting the swash plate and the plunger. While bearing the axial force of the high-pressure oil from the plunger cavity, the centrifugal force caused by the circumferential movement of the plunger will cause the slipper to tip relative to the swash plate surface. This tipping can lead to severe eccentric wear of the slipper and even cause the oil film to fail, resulting in the phenomenon of burned plates. Therefore, it is necessary to optimize the design of the slipper tipping fault of the rotating assembly.

[0214] Construct the maximum speed calculation formula before the slipper tips to verify the slipper tipping fault of the rotating assembly:

[0215]

[0216] Among them, ω is the rotational angular velocity of the plunger pump, M s is the mass of a single slipper, X s is the distance from the center of the slipper ball head to the centroid, R s is the radius of the slipper end face, M p is the mass of a single plunger, β is the swash plate angle, R is the radius of the plunger distribution circle, F sp is the spring force;

[0217] If the rotational angular velocity of the plunger pump is less than or equal to the maximum speed before the slipper tips, then execute the next step and output the maximum rotational speed of the rotating assembly before the slipper tipping fault as n hxqf , otherwise, optimize the spring force, the radius of the slipper end face, the mass of the slipper, the mass of the plunger, and the distance from the slipper ball head to the centroid; after optimization, re-execute step S41. If it still does not meet the requirements after n optimizations, then continue to execute the next step; n is generally taken as 3. Specifically, adopting a hollow plunger and other structures to reduce the mass of the plunger can increase the maximum rotational speed of the plunger pump before the slipper tipping fault. Increasing the spring force can increase the maximum rotational speed of the plunger pump before the slipper tipping fault. The plunger anti-cladding process can reduce the distance from the centroid of the slipper to the center of the ball hinge, thereby reducing the centrifugal tipping moment of the slipper. At the same time, adopting a fixed-clearance return mechanism can introduce a new anti-tipping force, further increasing the maximum rotational speed of the plunger pump before the slipper tipping fault. After optimization, re-execute step S4.

[0218] The construction method of the maximum speed calculation formula before the slipper tips is as follows:

[0219] Through force analysis, the moment balance equation perpendicular to the axis of the slipper is obtained as follows:

[0220] F sw R s -M s Rω 2 X s =0

[0221] Among them, F sw is the axial force of the slipper, R s is the radius of the slipper end face, M sis the mass of a single slipper, R is the radius of the plunger distribution circle, ω is the rotational angular velocity of the plunger pump, and X s is the distance from the center of the slipper ball head to the centroid.

[0222] According to the balance equation of the axial force of the slipper, the reaction force of the swash plate can be obtained as follows:

[0223]

[0224] Among them, F sw is the axial force of the slipper, F b is the force between the slipper and the plunger, N is the number of plungers, M s is the mass of a single slipper, R is the radius of the plunger distribution circle, ω is the rotational angular velocity of the plunger pump, and β is the swash plate angle.

[0225] Sum the circumferential forces on the plunger to obtain the force between the slipper and the plunger as:

[0226] F b = M p Rω 2 tanβ

[0227] Among them, F b is the force between the slipper and the plunger, M p is the mass of a single plunger, ω is the rotational angular velocity of the plunger pump, and β is the swash plate angle.

[0228] Combining the equations gives the calculation formula for the maximum speed before the cylinder block overturns.

[0229] S42. During the high-speed rotation of the plunger pump, the rotating cylinder block will be affected by the axial eccentric load, the plunger centrifugal force, and the deflection deformation of the main shaft, and will rotate relative to the valve plate in an overturned posture. Therefore, it is necessary to optimize the design of the cylinder block overturning fault of the rotating assembly.

[0230] S42. Construct the calculation formula for the maximum speed before the cylinder block overturns to verify the cylinder block overturning fault of the rotating assembly:

[0231]

[0232] Among them, ω is the rotational angular velocity of the cylinder block, R g is the outer radius of the cylinder block, N is the number of plungers, M p is the mass of a single plunger, M s is the mass of a single slipper, R is the radius of the plunger distribution circle, and β is the swash plate angle;

[0233] If the rotational angular velocity of the cylinder block is less than or equal to the maximum speed before the slipper overturns, then execute the next step and output the maximum rotational speed of the rotating assembly before the cylinder block overturns as n gtqf, otherwise, optimize the spring force, cylinder block radius, slipper mass, and plunger mass; after optimization, re - execute step S42. If it still does not meet the requirements after n optimizations, continue to execute the next step; generally, n is taken as 3. Specifically, adopting a structure such as a hollow plunger to reduce the plunger mass can increase the maximum rotational speed before the cylinder block of the plunger pump overturns. Increasing the spring force can increase the maximum rotational speed before the cylinder block of the plunger pump overturns. At the same time, adding auxiliary bearings to the outer circumference of the cylinder block to introduce a new supporting force for the cylinder block can increase the maximum rotational speed before the cylinder block of the plunger pump overturns. After optimization, re - execute step S4.

[0234] The construction method of the calculation formula for the maximum speed before the cylinder block overturns is as follows:

[0235] During the rotation of the cylinder block, it is subjected to the centrifugal force of the plunger, the inertial reaction force between the slipper and the plunger, and the spring force. The moment balance equation of the cylinder block is shown as follows:

[0236]

[0237] Among them, F sp is the spring force, R g is the outer - circle radius of the cylinder block, M p is the mass of a single plunger, M s is the mass of a single slipper, R is the distribution - circle radius of the plungers, ω is the rotational angular velocity of the cylinder block, N is the number of plungers, θ n is the rotational angle of the plunger, x n is the acting moment of the plunger, x n =Rtan(β)sin(θ n ), and β is the swash - plate angle.

[0238] For the plungers evenly spaced within a circular array around the center line of the cylinder block, the Lagrange trigonometric identity can be used to represent:

[0239]

[0240] Rearrange the moment balance equation to obtain the calculation formula for the maximum speed before the cylinder block overturns.

[0241] S43. During the high - speed rotation of the plunger pump, during the suction and discharge of oil by the plunger, the flow velocity in a local area increases, the kinetic energy increases, and the pressure potential energy decreases. When the pressure is lower than the saturated vapor pressure, cavitation will occur, seriously affecting the service performance of the plunger pump. Therefore, it is necessary to optimize the design of the cavitation and erosion failure of the rotating components.

[0242] Construct the calculation formula for the maximum speed before the cavitation and erosion failure of the plunger pump to verify the cavitation and erosion failure of the rotating components:

[0243]

[0244] where ω is the rotational angular velocity of the piston pump, P i is the pressure in the kidney-shaped interaction region, P cav is the pressure in the piston core cavity, A p is the piston area, A k is the area of the kidney-shaped interaction region, R is the radius of the piston distribution circle, β is the swash plate angle, A i is the inlet area, N is the number of pistons, and ρ is the medium cavitation pressure;

[0245] If the calculated rotational angular velocity of the piston pump is less than or equal to the maximum velocity before cavitation occurs, then proceed to the next step and output the maximum rotational speed of the rotating assembly before cavitation of the piston pump as n khqs , otherwise, optimize the suction pressure, medium cavitation pressure, inlet area, and the area of the kidney-shaped interaction region and then re-execute step S43. If it still does not meet the requirements after n optimizations, then continue to the next step; specifically, the smaller the cavitation of the fluid in the piston pump, the higher the running stability of the piston pump. Use finite element analysis to optimize the flow channel structure to reduce the cavitation of the piston pump.

[0246] The construction method of the formula for calculating the maximum velocity before the piston pump has a cavitation failure is as follows: The process of the oil fluid from entering the kidney-shaped interaction region to being sucked into the piston cavity should follow Bernoulli's equation, and the equilibrium equation is as follows:

[0247]

[0248] where P i is the pressure in the kidney-shaped interaction region, P cav is the pressure in the piston core cavity, u i is the flow velocity of the fluid entering the piston core cavity, u k is the flow velocity of the fluid in the kidney-shaped interaction region.

[0249] The flow velocity in the kidney-shaped interaction region can be decomposed into the normal velocity of the piston movement and the component of the tangential velocity of the cylinder block rotation:

[0250]

[0251] where u n is the normal velocity of the flow velocity of the fluid in the kidney-shaped interaction region, u t is the tangential velocity of the flow velocity of the fluid in the kidney-shaped interaction region. The calculation formula is as follows:

[0252]

[0253] u t = Rω

[0254] where A p is the piston area, A kis the area of the waist-shaped interaction region, R is the radius of the plunger distribution circle, β is the swash plate angle, and ω is the rotational angular velocity of the plunger pump.

[0255] The velocity of the fluid entering the plunger core cavity can be expressed as:

[0256]

[0257] where u i is the flow velocity of the fluid entering the plunger core cavity, N is the number of plungers, A p is the plunger area, R is the radius of the plunger distribution circle, β is the swash plate angle, ω is the rotational angular velocity of the plunger pump, and A i is the inlet area.

[0258] By integrating the above equations, the calculation formula for the maximum velocity before the cavitation and erosion failure of the plunger pump can be obtained.

[0259] Based on the three-dimensional model of the rotating assembly established in step S3, the fluid domain model of the rotating assembly is extracted by using solid space Boolean operations, as Figure 4 shown, and the fluid domain is meshed to meet the requirements of finite element analysis of the flow field.

[0260] According to the basic laws of fluid transmission in the rotating assembly of the plunger pump, the basic control equations of the fluid domain of the rotating assembly are established.

[0261] The increase in the fluid mass in each grid cell within the fluid domain of the rotating assembly is equal to the net mass of the fluid flowing into the grid cell within the same time interval. The mass conservation equation of the rotating assembly is established, and the specific expression is:

[0262]

[0263] where u, v, and w are the components of the velocity vector in the x, y, and z directions of the three coordinates in the coordinate system of the rotating assembly, t is the time, and ρ is the fluid density within the rotating assembly.

[0264] According to the fact that the change rate of the momentum of each grid cell in the rotating assembly with respect to time is equal to the sum of various forces acting on the grid cell from the outside, and the density of the compressible fluid changes with time, the Navier-Stokes equation for the compressible fluid of the rotating assembly is established, and the specific expression is:

[0265]

[0266] where μ is the dynamic viscosity of the fluid within the rotating assembly, p is the pressure of the fluid within the rotating assembly, ρ is the fluid density within the rotating assembly, and S u , S v and S w are the generalized source terms of the momentum equation respectively.

[0267] Apply the load characteristics and motion characteristics during the use of the plunger pump to the basic control equation of the fluid domain of the rotating assembly based on the design pressure and design speed of the plunger pump.

[0268] Obtain the gas volume fraction contour map of the rotating assembly fluid domain through finite element analysis to characterize the cavitation and erosion characteristics of the rotating assembly, as Figure 5 shown.

[0269] Using a supplementary oil pump or a booster impeller can increase the oil suction pressure of the plunger pump, which is the main way to increase the speed limit of the plunger pump under cavitation and erosion failures, as Figure 6 shown as the curve of the speed limit of the plunger pump under cavitation and erosion failures varying with the oil suction pressure. In addition, increasing the cavitation pressure of the medium, expanding the inlet area and the area of the waist-shaped interaction zone can also increase the speed limit of the plunger pump under cavitation and erosion failures, and then re-execute step S4 after optimization.

[0270] S44. Judge whether the maximum speed n pmax of the plunger pump meets the requirements. If it meets, execute the next step. If not, perform the fifth parameter optimization and then re-execute steps S2 - S4. If the requirements are still not met after executing step S44 for n times, output the highest speed of the plunger pump at the current displacement and pressure as n pmax ; n pmax The calculation formula of is:

[0271] n pmax = min(n hxqf , n gtqf , n khqs ).

[0272] The specific steps of the fifth parameter optimization are:

[0273] If the highest speed of the plunger pump is n pmax = n hxqf , then under the basic displacement limit formula, reduce the number of plungers, lower the swash plate angle and reduce the plunger distribution circle radius.

[0274] If the highest speed of the plunger pump is n pmax = n gtqf , then under the basic displacement limit formula, reduce the number of plungers, lower the swash plate angle, reduce the plunger distribution circle radius and increase the cylinder block radius.

[0275] If the highest speed of the plunger pump is n pmax = n khqs , then under the basic displacement limit formula, increase the number of plungers, lower the swash plate angle and reduce the plunger distribution circle radius, where the basic displacement limit formula is the relationship between the plunger hole diameter, distribution circle diameter and maximum target displacement in step S21.

[0276] S5. Output the optimized structural parameters, component material properties, and manufacturing processes of the plunger pump rotating assembly.

[0277] Optimize the design of the plunger pump rotating assembly from the aspects of displacement limit, operating pressure limit, and speed limit based on the above results.

[0278] Structural optimization design of the plunger pump rotating assembly: Through continuous cycling of the above process, obtain the structure, material, and manufacturing process of the rotating assembly of a high-speed and high-pressure axial plunger pump with a short design cycle, improving the safety, stability, and service life of the fluid transmission system. At the same time, the steps of the present invention can be adjusted, combined, and deleted according to actual needs.

[0279] The embodiments described above are only for describing the implementation manners of the present invention, not for limiting the scope of the present invention. Without departing from the design spirit of the present invention, various deformations and improvements made by those of ordinary skill in the art to the technical solutions of the present invention shall fall within the protection scope determined by the claims of the present invention.

Claims

1. A method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump, characterized in that: The specific steps include: S1. Determine the design objectives and initial structural material properties of the piston pump rotating assembly. The design objectives include the maximum target displacement, maximum speed and rated pressure of the piston pump; S2. determining structural parameters of the plunger pump rotating assembly based on the maximum target displacement; S3, establishing a three-dimensional model of the rotating assembly according to the structural parameters of the rotating assembly of the plunger pump obtained in step S2, setting the initial material properties and processes of the components, judging the strength of the components under the rated pressure condition of the plunger pump based on the structural deformation, and judging whether the rated pressure is met, if so, proceeding to step S4, otherwise, re-execute steps S2-S3 after optimizing the parameters; S4, according to the structural parameters of the rotating assembly of the plunger pump obtained in step S2, the parameters of the faults occurring during the operation of the rotating assembly under rated speed condition are judged and optimized respectively, and it is judged whether the maximum speed requirement is met. If so, step S5 is performed, otherwise, the parameters are optimized and steps S2-S4 are re-executed; S5. Output the optimized structural parameters of the piston pump rotating assembly and the material properties and processes of its components.

2. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, characterized in that: In step S2, the structural parameters of the piston pump rotating assembly include cylinder parameters, piston parameters, slipper parameters, return plate parameters, valve plate parameters and swash plate parameters.

3. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 2, characterized in that: The specific steps of step S2 include: S21, determining cylinder parameters, which include plunger hole diameter, distribution circle diameter, wall thickness between cylinder plunger holes and cylinder length; Among them, the relationship between the plunger hole diameter, the distribution circle diameter and the maximum target displacement is as follows: Among them, V g is the design flow rate of the plunger pump, z is the number of plungers, d k is the plunger diameter, D T is the plunger distribution circle diameter, β is the swash plate inclination angle; By calculating the average normal stress on the cylinder plunger hole, it can be judged whether the wall thickness between the cylinder plunger holes meets the requirements. The calculation formula for the average normal stress on the cylinder plunger hole is as follows: Among them, σ is the average normal stress of the cylinder plunger hole, is the pressure in the plunger cavity, ε is the resultant force angle between the plunger hole, D T is the plunger distribution circle diameter, d k is the plunger diameter, z is the number of plungers; The cylinder length satisfies the following formula: L c =l k +2Rtanβ+(2~3)+s+L h Among them, L c is the cylinder length, l k is the minimum connecting length of the plunger, d k is the plunger diameter, R is the plunger distribution circle radius, β is the inclination angle of the inclined plate, s is the length of the distribution hole connecting the plunger hole and the distribution plate, that is, the thickness of the cylinder bottom, L h It is the protruding length of the cylinder spline; S22, determining plunger parameters, the plunger parameters including plunger length and plunger ball head diameter; The plunger length satisfies the following formula: L z =l0+h p +l k Among them, L z is the plunger length, l0 is the minimum extension length, and l0 = 0.2d k ,h p is the plunger stroke, h p =2Rtanβ,l k is the minimum connection length, l k =(2~2.5)d k ; Plunger ball head diameter d2 = (0.7 ~ 0.8) d k ; S23, determining the parameters of the sliding shoe, which include the diameter of the sliding shoe ball socket, the inner diameter of the sliding shoe sealing belt, the outer diameter of the sliding shoe sealing belt, the outer diameter of the sliding shoe ball cup and the diameter of the sliding shoe damping hole; The diameter of the sliding shoe ball socket d8 is equal to the diameter of the plunger ball head d2; The inner diameter of the sliding shoe sealing belt satisfies the following formula: Among them, d4 is the inner diameter of the sliding shoe sealing belt, k is the sealing coefficient, d k is the plunger diameter, β is the tilt of the swash plate angle, d6 is the outer diameter of the sliding shoe sealing belt, and the outer diameter of the sliding shoe sealing belt meets s h =0.2-1mm; Shoe cup outer diameter d7 = (0.95 ~ 1) d k ; The diameter d3 of the damping hole of the sliding shoe is calculated by the residual clamping force method; S24, determining the parameters of the return disk, the return disk parameters including the return disk distribution circle diameter and the return disk hole diameter; The calculation formula of the return disk distribution circle diameter is as follows: Where D3 is the distribution circle diameter of the return plate, R is the radius of the plunger distribution circle, and β is the inclination angle of the inclined plate; The return plate aperture satisfies the following formula: d’5=2|ε max |+d7+2a min Where d'5 is the aperture of the return disk, |ε max | is the radial deviation between the elliptical trajectory of the center of the sliding shoe and the distribution circle of the return disk, d7 is the outer diameter of the ball cup of the sliding shoe, a min is the clearance between the neck of the sliding shoe and the hole of the return plate, the outer diameter of the sliding shoe d6≥d'5, the outer diameter of the sliding shoe d6=d'6+(1~3); S25, determining the parameters of the distribution plate and verifying whether the distribution plate meets the requirements, the distribution plate parameters including the size and deflection angle of the distribution pair sealing belt; The width of the inner and outer sealing bands of the distribution pair is (0.18~0.2)d k ; When the initial pressure rises to the pre-boost pressure, the volume compression is calculated by the following formula: Where ΔV is the volume compression, p d is the pre-boost pressure, p0 is the initial pressure, E0 is the elastic modulus of the oil, V0 is the closed volume of the plunger cavity at the bottom dead center, d k is the plunger diameter, h max is the maximum stroke of the plunger; The deflection angle of the valve plate is calculated by the following formula: Where ΔV is the volume compression, d k is the plunger diameter, R is the plunger distribution circle radius, Δh is the plunger stroke, β is the swash plate inclination angle, is the deflection angle of the valve plate; The ratio of the radius of the center of the waist groove of the cylinder body and the distribution plate to the radius of the plunger distribution circle is 0.7-1.0, the width of the waist groove is 0.35-0.5 of the plunger diameter, and the width of the inner and outer sealing bands of the distribution plate is 0.1-0.2 of the plunger diameter; By calculating the linear velocity at the center of the waist groove of the cylinder body and the distributor plate, it can be judged whether the distributor plate meets the requirements. The calculation formula is as follows: Among them, V R R is the centerline speed of the waist groove of the distribution plate, y n is the center radius of the waist groove of the distribution plate, max is the maximum speed of the plunger pump; S26, determining the swash plate parameters, the swash plate parameters including the swash plate inclination angle and the minimum allowable distance between the two chamfered edge planes; The swash plate inclination angle β is 15° to 22°; The minimum allowable distance between two chamfered planes satisfies the following formula: B spmin =D+d’6+2A min Among them, B spmin is the minimum allowable distance between the two chamfered planes, D is the major axis of the elliptical trajectory of the sliding shoe on the swash plate, d'6 is the outer diameter of the sliding shoe sealing band, A min is the minimum distance from the edge of the sliding shoe to the edge of the swash plate; S27, calculating the design displacement according to the structural parameters, determining whether the design displacement is greater than the target displacement, if greater, executing the next step, if less than or equal to, performing the first parameter optimization and then returning to step S21 and re-executing step S2; S28, determine whether the designed displacement is less than 1.1 times the target displacement, if so, execute the next step, if not, perform the second parameter optimization and then return to step S21 and re-execute step S2.

4. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, characterized in that: The specific steps of step S3 include: S31. Establish a three-dimensional model of the rotating component and initially set the material and process of the components; S32, judging the structural strength of components under rated pressure based on the principle of structural deformation, specifically including the following sub-steps: S321. Divide the solid domain mesh for finite element analysis based on the structure of the rotating component parts; S322. The structural deformation and stress of the rotating assembly are affected by the temperature field and fluid field during the operation of the plunger pump and are transmitted between the structures. The flow field pressure load is applied to the contact part between the rotating assembly and the fluid. The temperature load during the operation of the plunger pump is applied to the rotating assembly. The position of the rotating assembly is constrained according to the actual adaptation working conditions. S323. Perform structural deformation analysis of rotating components under multi-field coupling, including structural deformation and stress under temperature load and fluid load; The force deformation of the rotating component satisfies the following formula: [K] × {δ} = {F} Where, [K] is the stiffness matrix of the rotating component system, {δ} is the node displacement matrix of the rotating component system, and {F} is the force matrix of the rotating component system; The thermal deformation of the rotating component satisfies the following formula: Among them, f T is the structural deformation, α T is the thermal expansion coefficient of the rotating component material, is the temperature difference; S33, extracting the structural deformation and stress-strain cloud diagram of the rotating component, judging that the oil film gap of the rotating component has no interference after the structural deformation and the structural stress is less than the required stress of the material, then calculating whether the maximum pressure under the current required stress of the material is greater than the rated pressure, if so, executing the next step, if not, re-executing step S2 and step S3 after performing the third parameter optimization; S34, calculate whether the maximum pressure under the current material required stress is less than 1.1 times the rated pressure, if so, execute the next step, if not, re-execute step S2 and step S3 after performing the fourth parameter optimization.

5. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 4, characterized in that: If the rated pressure requirement is still not met after step S33 is executed n times, it is output that the rated pressure requirement cannot be met under the current parameters.

6. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, characterized in that: Step S4 specifically includes the following sub-steps: S41. Construct a maximum speed calculation formula before the sliding shoe overturns to verify the sliding shoe overturning failure of the rotating component: Where ω is the angular velocity of the piston pump, M s is the mass of a single sliding shoe, X s R is the distance from the center of the ball head to the center of mass, s M is the end radius of the sliding shoe, p is the mass of a single plunger, β is the inclination angle of the swash plate, R is the radius of the plunger distribution circle, F sp is the spring force; If the angular velocity of the plunger pump is less than or equal to the maximum velocity before the slipper tipping over, the next step is executed and the maximum speed of the rotating component before the slipper tipping over is output as n. hxqf , otherwise, optimize the spring force, the end radius of the sliding shoe, the mass of the sliding shoe, the mass of the plunger and the distance from the ball head of the sliding shoe to the center of mass; after optimization, re-execute step S41, if it is still not satisfied after optimization n times, continue to execute the next step; S42. Construct a maximum speed calculation formula before the cylinder overturns to verify the cylinder overturning failure of the rotating component: Where ω is the angular velocity of the cylinder, R g is the outer radius of the cylinder, N is the number of plungers, M p is the mass of a single plunger, M s is the mass of a single sliding shoe, R is the radius of the plunger distribution circle, and β is the inclination angle of the swash plate; If the cylinder body rotation angular velocity is less than or equal to the maximum speed before the slipper overturns, the next step is executed and the maximum speed of the rotating component before the cylinder body overturns is output as n. gtqf , otherwise, optimize the spring force, cylinder radius, shoe mass and plunger mass; after optimization, re-execute step S42, and if it is still not satisfied after optimization n times, continue to execute the next step; S43. Construct a maximum speed calculation formula before the piston pump cavitation failure occurs to verify the cavitation failure of the rotating component: Where ω is the angular velocity of the piston pump, P i is the pressure in the waist-shaped interaction zone, P cav is the plunger core chamber pressure, A p is the plunger area, A k is the waist-shaped interaction area, R is the radius of the plunger distribution circle, β is the inclination angle of the inclined plate, A i is the inlet area, N is the number of plungers, and ρ is the medium cavitation pressure; If the angular velocity of the piston pump is less than or equal to the maximum velocity before cavitation occurs, the next step is executed and the maximum speed of the rotating component before cavitation occurs is output as n. khqs Otherwise, optimize the oil suction pressure, medium cavitation pressure, inlet area and waist-shaped interaction area and re-execute step S43. If the optimization is still not satisfied after n times, continue to the next step; S44, determine the maximum speed n of the plunger pump pmax Whether the requirements are met, if so, execute the next step, if not, perform the fifth parameter optimization and re-execute steps S2-S4, if the requirements are still not met after executing S44 n times, then output the maximum speed of the plunger pump under the current displacement and pressure as n pmax ;n pmax The calculation formula is: n pmax =min(n hxqf ,n gtqf ,n khqs )。 7. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, characterized in that: In step S432, the basic control equation of the fluid domain of the rotating component is established according to the basic law of fluid transmission in the rotating component of the plunger pump.

8. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, characterized in that: Step S432 specifically includes the following sub-steps: S4321. Construct the specific expression of the mass conservation equation of the rotating component; S4322. Establish the compressible fluid Navier-Stokes equations for the rotating assembly; S4323. Apply the load characteristics and motion characteristics of the plunger pump during use to the basic control equations of the fluid domain of the rotating component based on the design pressure and design speed of the plunger pump; S4324. The gas volume fraction cloud diagram of the fluid domain of the rotating component is obtained through finite element analysis to characterize the cavitation characteristics of the rotating component.

9. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, characterized in that: In step S24, the radial deviation between the elliptical trajectory of the shoe center and the distribution circle of the return disk |ε max |For: Among them, |ε max | is the radial deviation between the elliptical trajectory of the sliding shoe center and the distribution circle of the return plate, R is the radius of the plunger distribution circle, and β is the inclination angle of the inclined plate.

10. The method for optimizing the design of a rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, characterized in that: The specific steps of the first parameter optimization in step S27 are to increase the number of plungers, plunger diameter, plunger distribution circle diameter and swash plate inclination angle; The specific steps of the second parameter optimization in step S28 are to reduce the number of plungers, plunger diameter, plunger distribution circle diameter and swash plate inclination angle; The specific steps of the third parameter optimization in step S33 are to increase the structural strength and structural rigidity of the material for the area that does not meet the rated pressure, reduce the thermal expansion coefficient of the material, and enhance the wall thickness; The specific step of optimizing the fourth parameter in step S34 is to reduce the structural strength and structural rigidity of the overall material; The specific steps of the fifth parameter optimization in step S44 are: If the maximum speed of the plunger pump is n pmax =n hxqf , then under the basic displacement limitation formula, reduce the number of plungers, reduce the inclination angle of the swash plate and reduce the radius of the plunger distribution circle; If the maximum speed of the plunger pump is n pmax =n gtqf , then under the basic displacement limitation formula, reduce the number of plungers, reduce the inclination angle of the swash plate, reduce the radius of the plunger distribution circle and increase the radius of the cylinder body; If the maximum speed of the plunger pump is n pmax =n khqs , then under the basic displacement limitation formula, increase the number of plungers, reduce the inclination angle of the swash plate and reduce the radius of the plunger distribution circle, where the basic displacement limitation formula is the relationship between the plunger hole diameter, the distribution circle diameter and the maximum target displacement in step S21.

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