Optimization Design Method for the Rotating Assembly of High-Speed and High-Pressure Axial Piston Pumps

Abstract

The present invention discloses an optimized design method for the rotating assembly of a high-speed and high-pressure axial piston pump. The specific steps are as follows: S1, input the design objectives, namely the displacement, speed and rated pressure of the piston pump; S2, design the structural parameters of the rotating assembly of the piston pump based on the target displacement and provide feedback for optimization; S3, establish a three-dimensional model of the rotating assembly, judge the strength of components under the rated pressure based on the principle of structural deformation and provide feedback for optimization; S4, judge and optimize the tilting of the slipper, tilting of the cylinder block and cavitation failure during the high-speed process of the rotating assembly based on speed limits; S5, output the structural parameters of the rotating assembly of the piston pump, as well as the materials and processes of components. The present invention can start from the design objectives and fault judgment and optimization of the piston pump, realize the rapid optimized design of the rotating assembly of the piston pump, prevent the high-speed and high-pressure failure of the piston pump during operation, shorten the design cycle of the piston pump and improve the operation reliability of the piston pump.

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 developing 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, the piston pump may experience failure risks such as slipper tilting, cylinder block tilting, and cavitation. 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 failure. When the requirements of displacement limit, pressure limit, and speed limit are met, the structural parameters of the piston pump rotating assembly and 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 properties of the rotating assembly of the piston pump. The design objectives include the maximum target displacement, the highest speed, and the rated pressure of the piston pump;

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

[0008] S3. Establish a three-dimensional model of the rotating assembly according to the structural parameters of the rotating assembly of the piston pump 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 determine whether the rated pressure is satisfied. If satisfied, 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, after optimizing the parameters, 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, valve 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 pitch 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 pitch 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 pitch circle diameter of the plungers, 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 pitch circle diameter of the plungers, 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 engagement length of the plunger, dk Where \(d\) is the plunger diameter, \(R\) is the radius of the plunger distribution circle, \(\beta\) is the swash plate angle, \(s\) is the length of the port hole connecting the plunger hole and the valve plate, i.e., the cylinder bottom thickness, \(L\) h is the length of the spline protrusion 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] Wherein, \(L\) z is the plunger length, \(l_0\) is the minimum outer extension length, and \(l_0 = 0.2d\) k \(h\) p is the plunger stroke, \(h\) p = \(2R\tan\beta\), \(l\) 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 seal band, the outer diameter of the swash plate seal 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 seal band satisfies the following formula:

[0031]

[0032] Wherein, \(d_4\) is the inner diameter of the swash plate seal band, \(k\) is the seal coefficient, \(d\) k is the plunger diameter, \(\beta\) is the swash plate angle, \(d'_6\) is the outer diameter of the swash plate seal band, and the outer diameter of the swash plate seal band satisfies \(s\) h = \((0.2 - 1)\text{mm}\);

[0033] The outer diameter of the swash plate ball cup \(d_7=(0.95 - 1)d\) k ;

[0034] The diameter of the swash plate damping hole \(d_3\) is calculated by the method of the remaining clamping force;

[0035] S24. Determine the return spring disk parameters, which include the return spring disk distribution circle diameter and the return spring disk hole diameter;

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

[0037]

[0038] Wherein, D3 is the pitch circle diameter of the return disk, R is the pitch 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 | + d7 + 2a min

[0041] Wherein, d’5 is the aperture of the return disk, |ε max | is the radial deviation between the elliptical locus of the swash plate center and the pitch circle of the return disk, d7 is the outer diameter of the ball cup of the swash plate, and a min is the clearance between the neck of the swash plate and the hole of the return disk. The outer diameter d6 of the swash plate is ≥ d’5, and the outer diameter d6 of the swash plate = 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 widths of the inner and outer sealing bands of the valve plate pair are (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] Wherein, 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 position, d k is the plunger diameter, and h max is the maximum stroke of the plunger;

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

[0048]

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

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

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

[0052]

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

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

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

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

[0057] B spmin = D + d’6 + 2A min

[0058] 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 slipper on the swash plate, d’6 is the outer diameter of the slipper sealing band, and 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, proceed to 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. Judge whether the designed displacement is less than 1.1 times the target displacement. If it is less, proceed to the next step. If not, perform the second parameter optimization and then 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 assembly, and initially set the materials and processes of the components;

[0063] S32. Judge the structural strength of the components under the rated pressure based on the principle of structural deformation, which specifically includes the following sub-steps:

[0064] S321. Divide the solid domain mesh for finite element analysis based on the component structure of the rotating assembly;

[0065] 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 structures. Apply the flow field pressure load to the parts of the rotating assembly in contact with the fluid, apply the temperature load during the operation of the plunger pump to the rotating assembly, and apply position constraints to the rotating assembly according to the actual operating conditions;

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

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

[0068] [K]×{δ}={F}

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

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

[0071]

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

[0073] S33. Extract the structural deformation and stress-strain nephogram of the rotating assembly. If there is no interference in the oil film clearance of the rotating assembly after structural deformation and the structural stress is less than the allowable stress of the material, 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 steps S2 and S3 after optimizing the third parameter;

[0074] 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; if not, re-execute steps S2 and S3 after optimizing the fourth parameter.

[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 overturning to verify the slipper overturning failure of the rotating assembly:

[0078]

[0079] Among them, ω is the rotational angular velocity of the plunger pump, and 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;

[0080] If the rotational angular velocity of the plunger pump is less than or equal to the maximum speed before slipper overturning, then execute 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, slipper end face radius, slipper mass, plunger mass, and the distance from the slipper ball head to the centroid; after optimization, re - execute step S41, and if it still does not meet the requirements after n optimizations, then continue to execute the next step;

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

[0082]

[0083] 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, 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 slipper overturning, then execute 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 block radius, slipper mass, and plunger mass; after optimization, re - execute step S42, and if it still does not meet the requirements after n optimizations, then continue to execute the next step;

[0085] S43. Construct the formula for calculating the maximum speed before 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 area of the plunger, 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, Ai The inlet area is \(A\), the number of plungers is \(N\), and the cavitation pressure of the medium is \(\rho\).

[0088] If the rotational angular velocity of the plunger 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 plunger pump as \(n\). khqs Otherwise, optimize the suction pressure, the cavitation pressure of the medium, the 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.

[0089] S44. Determine whether the maximum rotational speed \(n\) of the plunger pump pmax meets the requirements. If it meets the requirements, then proceed to the next step. If it does not meet the requirements, then optimize the fifth parameter and re-execute steps S2 - S4. If it still does not meet the requirements after executing step S44 \(n\) times, then output the highest rotational speed of the plunger pump as \(n\) under the current displacement and pressure. pmax ; \(n\) pmax The calculation formula of

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

[0091] Preferably, in step S432, the basic control equation of the fluid domain of the rotating assembly is established according to the basic law 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 equation of the fluid domain of the rotating assembly based on the design pressure and design rotational 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 characteristics of the rotating assembly.

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

[0098]

[0099] where \(|\varepsilon\) 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 for optimizing the first parameter 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 for optimizing the second parameter 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 for optimizing the third parameter in step S33 are to increase the structural strength and stiffness of the material for the area not meeting the rated pressure, reduce the thermal expansion coefficient of the material, and perform wall thickness enhancement;

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

[0104] The specific steps for optimizing the fifth parameter in step S44 are:

[0105] If the maximum rotational 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 rotational 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 rotational 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, 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 components of the plunger pump under the conditions of displacement limit, pressure limit, and rotational speed limit, can accurately obtain the deformation and stress nephograms of the key components, the cavitation and erosion nephogram of the internal flow field of the rotating components, combines with the critical condition theoretical formula of the plunger pump, discriminates whether the design parameters meet the expected requirements, and through multiple closed-loop feedback for optimization and improvement, obtains the optimal plunger pump parameters that meet the requirements.

[0110] (2) Based on the numerical simulation results obtained from the optimized design method of the rotating assembly of the high-speed high-pressure axial piston pump and the experimental results, the rotating assembly of the piston pump is improved and structurally optimized multiple times. Through experimental comparison, the design parameters and material processes of the components for reliable operation are obtained. Therefore, the present invention is an analysis method with high precision, 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. BRIEF DESCRIPTION OF THE DRAWINGS

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

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

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

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

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

[0116] Figure 6 is a curve diagram of the change of cavitation and erosion speed limit with the suction pressure in the optimized design method of the rotating assembly of the high-speed high-pressure axial piston pump of the present invention. DETAILED DESCRIPTION OF THE INVENTION

[0117] To elaborate on the technical content, achieved objectives and effects of the present invention in detail, the following will be described in detail with reference to 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, specifically including the following steps:

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

[0120] S2. Determine the structural parameters of the rotating assembly of the piston pump 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 of the cylinder block and the length of the cylinder block.

[0122] In the process of designing the cylinder block parameters, first of all, the design requirements of the displacement of the plunger pump must be met. The relationship between the plunger diameter, the diameter of the plunger distribution circle, and the pump displacement is as follows:

[0123]

[0124] Where 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 diameter of the plunger distribution circle, and β is the swash plate angle.

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

[0126]

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

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

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

[0130] L c =l k +2R tanβ+(2~3)+s+L h

[0131] Where L c is the length of the cylinder block, l k is the minimum contact 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 plunger distribution circle, β is the swash plate angle. To prevent the plunger from colliding with the cylinder block, a certain length is left, generally 2~3 mm. s is the length of the port plate hole connecting the plunger hole and the port plate, that is, the bottom thickness of the cylinder block. Generally, s=(0.4~0.6)d k ,Lh is the protruding length of the spline on the cylinder block.

[0132] S22. Determine the plunger parameters, which include the plunger length and the diameter of the plunger ball head.

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

[0134] L z = l0 + h p + l k

[0135] Where, L z is the plunger length, l0 is the minimum extension length, generally l0 = 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 diameter of the plunger ball head d2 = (0.7 - 0.8)d k .

[0137] In addition, the diameter of the damping hole of the plunger d5 = (0.5 - 2) mm. Since the swashplate is designed by the method of residual pressing force, the damping hole does not play any damping role. Therefore, the damping hole of the plunger can be appropriately taken as a larger value.

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

[0139] The diameter of the swashplate ball socket d8 is equal to the diameter of the plunger ball head d2.

[0140] The inner diameter of the swashplate sealing belt should satisfy the following formula:

[0141]

[0142] Where, d4 is the inner diameter of the swashplate sealing belt, k is the sealing coefficient, d k is the plunger diameter, β is the swashplate inclination angle, d′6 is the outer diameter of the swashplate sealing belt, and should satisfy To ensure a certain gap between the swashplates, take s h = (0.2 - 1) mm.

[0143] The outer diameter of the swashplate d6 = d′6 + (1 - 3), and the outer diameter of the swashplate ball cup d7 = (0.95 - 1)d k .

[0144] The diameter d3 of the damping hole of the slipper is calculated by the remaining 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-slipper, the pressure in the plunger chamber, and the pre-clamping force of the central spring, as shown below:

[0146]

[0147] Among them, F N is the clamping force on the slipper, F p is the pressure in the plunger chamber, F a is the inertial force of the plunger and the 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] Among them, Q s is the damping hole flow rate, p p is the pressure in the plunger chamber, p r is the pressure in the central oil chamber, C op 、C os are the flow coefficients of the plunger and slipper damping holes respectively, and the calculation formulas are as follows:

[0151]

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

[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 slipper seal band, R s is the outer diameter of the slipper seal band, p r is the pressure in the central oil chamber.

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

[0157]

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

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

[0160]

[0161] Among them, $\varPsi$ is the remaining compression force coefficient, and generally $\varPsi = 1 - 1.05$.

[0162] According to the above and the reasonable range of the oil film thickness, the diameter $d_3$ of the damping hole of the slipper can be determined.

[0163] 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.

[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 plunger, and $\beta$ is the swash plate angle.

[0167] The radial deviation between the elliptical trajectory of the slipper center and the distribution circle of the return disk is:

[0168]

[0169] Among them, $|\varepsilon$ max $|$ is the radial deviation between the elliptical trajectory of the slipper center and the distribution circle of the return disk, $R$ is the distribution circle radius of the plunger, and $\beta$ is the swash plate angle.

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

[0171] $d'_5 = 2|\varepsilon$ max $| + d_7 + 2a$ min

[0172] Among them, $d'_5$ is the aperture of the return disk, $|\varepsilon$ 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, and $a$ min is the clearance between the neck of the slipper and the hole of the return disk. Generally, $a$ min $= 0.2 - 1$ mm. At the maximum trajectory deviation, to ensure that there is a certain overlap between the maximum outer diameter of the slipper head and the return disk, the outer diameter $d_6$ of the slipper is $d_6 \geq d'_5$.

[0173] S25. Determine the parameters of the distribution plate and verify whether the distribution plate meets the requirements. The distribution plate parameters include the size and deflection angle of the distribution pair sealing strip.

[0174] The radius of the center of the waist groove of the cylinder body and the distribution plate should not be greater than the radius of the plunger distribution circle. The ratio of the two can be between 0.7 and 1.0. The width of the waist groove is 0.35 to 0.5 of the plunger diameter, and the width of the inner and outer sealing bands of the distribution plate is selected to be 0.1 to 0.2 of the plunger diameter.

[0175] After the center radius of the cylinder body and the distributor plate waist groove is determined, it should be determined whether the linear speed at the center meets the operating range. The calculation formula is as follows:

[0176]

[0177] 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 It is the maximum speed of the plunger pump.

[0178] The width of the inner and outer sealing belts ranges from (0.18 to 0.2)d k .

[0179] In order to effectively reduce pressure shock and lower the noise of the plunger pump, the waist-shaped window of the distribution plate is usually deflected to achieve pre-increase and pre-depressurization.

[0180] When the initial pressure rises 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, 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.

[0183] In order to make the necessary stroke of the plunger Δh when the compression volume is ΔV, the following formula can be obtained:

[0184]

[0185] 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 distribution plate.

[0186] Combining the above formulas, we can get the design value of the distribution plate deflection angle.

[0187] S26. Swash plate parameter design, including the swash plate inclination angle and the minimum allowable distance between two edge-cutting planes.

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

[0189] The minimum allowable distance between two edge-cutting planes shall satisfy the following formula:

[0190] B spmin = D + d’6 + 2A min

[0191] Among them, B spmin is the minimum allowable distance between two edge-cutting 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 belt, and 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 designed displacement according to the structural parameters, and judge whether the designed displacement is greater than the target displacement. If it is greater, proceed to 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.

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

[0194] The specific steps of the first parameter optimization are to increase the number of plungers, the plunger diameter, the plunger distribution circle diameter, and the swash plate inclination angle. The specific steps of the second parameter optimization are to reduce the number of plungers, the plunger diameter, the plunger distribution circle diameter, and the swash plate inclination angle.

[0195] S3. Establish a three-dimensional model of the rotating assembly according to the structural parameters of the plunger 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 plunger 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, it includes:

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

[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 plunger pump and are transmitted between structures. Apply the flow field pressure load to the part of the rotating assembly in contact with the fluid, apply the temperature field during the operation of the plunger pump to the rotating assembly, and impose position constraints on the rotating assembly according to the actual operating conditions.

[0200] S323. Conduct a structural deformation analysis of the rotating assembly under multi-field coupling, including 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] where [K] is the system stiffness matrix of the rotating assembly, {δ} is the system nodal displacement matrix of the rotating assembly, and {F} is the system force matrix of the rotating assembly.

[0204] The thermal deformation term of the rotating assembly satisfies the following formula:

[0205]

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

[0207] S324. Extract the structural deformation and stress-strain nephograms of the rotating assembly. As Figure 3 shown, if there is no interference in the oil film clearance of the rotating assembly 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 nephograms of the rotating assembly. Judge that there is no interference in the oil film clearance of the rotating assembly 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, perform the next step. If not, re-execute step S2 and step S3 after optimizing the third parameter. 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.

[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, perform the next step. Otherwise, re-execute step S2 and step S3 after optimizing the fourth parameter.

[0210] The specific steps for optimizing the third parameter are to increase the structural strength and stiffness of the material in the area where the rated pressure is not met, reduce the thermal expansion coefficient of the material, and enhance the wall thickness. The specific steps for optimizing the fourth parameter are to reduce the structural strength and stiffness of the overall material.

[0211] S4. Based on the structural parameters of the rotating assembly of the piston pump 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; if not, after parameter optimization, steps S2 - S4 are executed again.

[0212] The specific steps are as follows:

[0213] S41. The slipper is a mechanism connecting the swash plate and the piston. When it bears the axial force of the high - pressure oil from the piston chamber, at the same time, the centrifugal force caused by the circumferential movement of the piston will cause the slipper to overturn relative to the surface of the swash plate. This kind of overturning is that the slipper wears severely and even causes the oil film to fail and the phenomenon of burning the plate. Therefore, it is necessary to optimize the design of the slipper overturning fault of the rotating assembly.

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

[0215]

[0216] Among them, 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;

[0217] If the rotational angular velocity of the piston pump is less than or equal to the maximum speed before the slipper overturning, the next step is executed and the maximum speed of the rotating assembly before the slipper overturning is output as n hxqf , otherwise, 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 are optimized; after optimization, step S41 is executed again. If it still does not meet the requirement after n optimizations, the next step is continued; n is generally taken as 3. Specifically, adopting a hollow piston and other structures to reduce the piston mass can increase the maximum speed of the piston pump before the slipper overturning, increasing the spring force can increase the maximum speed of the piston pump before the slipper overturning, and the piston reverse - coating process can reduce the distance from the centroid of the slipper to the center of the ball hinge, thereby reducing the centrifugal overturning moment of the slipper. At the same time, adopting a fixed - clearance return mechanism can introduce a new anti - overturning force, further increasing the maximum speed of the piston pump before the slipper overturning. After optimization, step S4 is executed again.

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

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

[0220] F sw R s -M s Rω 2 X s =0

[0221] Wherein, F sw is the axial force of the slipper, R s is the radius of the slipper end face, 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 X s is the distance from the center of the slipper ball head to the centroid.

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

[0223]

[0224] Wherein, 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 inclination 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] Wherein, 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 inclination angle.

[0228] Combining the equations, the maximum speed calculation formula before the cylinder block overturning can be obtained.

[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. Establish the maximum speed calculation formula before the cylinder block overturns to verify the cylinder block overturning fault of the rotating assembly:

[0231]

[0232] Among them, ω is the cylinder block rotation angular velocity, R g is the outer circle 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 plunger distribution circle radius, and β is the swash plate angle;

[0233] If the cylinder block rotation angular velocity 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 overturning fault 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, then continue to execute the next step; n is generally taken as 3. Specifically, adopting structures such as hollow plungers to reduce the plunger mass can increase the maximum rotational speed before the cylinder block overturning fault of the plunger pump, increasing the spring force can increase the maximum rotational speed before the cylinder block overturning fault of the plunger pump. At the same time, installing auxiliary bearings on the outer circumference of the cylinder block and other measures to introduce new supporting forces for the cylinder block can increase the maximum rotational speed before the cylinder block overturning fault of the plunger pump. After optimization, re - execute step S4.

[0234] The construction method of the maximum speed calculation formula 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 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 plunger distribution circle radius, ω is the cylinder block rotation angular velocity, N is the number of plungers, θ n is the plunger rotation angle, 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 cylinder block center line, 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, the flow velocity in a local area increases during the oil suction and discharge process of the plunger, and the increase in kinetic energy causes the pressure potential energy to decrease. 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 assembly.

[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 assembly:

[0243]

[0244] 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 area of the plunger, 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 cavitation pressure of the medium;

[0245] If the calculated 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 cavitation and erosion of the plunger pump as n khqs , otherwise, optimize the oil suction pressure, medium cavitation pressure, inlet area, and kidney-shaped interaction area, and then re-execute step S43. If it still does not meet the requirements after n optimizations, then continue to execute the next step; specifically, the smaller the cavitation and erosion of the fluid in the plunger pump, the higher the running stability of the plunger pump. Use finite element analysis to optimize the flow channel structure to reduce the cavitation and erosion of the plunger pump.

[0246] The construction method of the calculation formula for the maximum speed before the cavitation and erosion failure of the plunger pump is as follows: The process of the oil fluid from entering the kidney-shaped interaction area to being sucked into the plunger cavity should follow the Bernoulli equation, and the balance equation is as follows:

[0247]

[0248] Among them, P i is the pressure in the kidney-shaped interaction area, P cav is the pressure in the plunger core cavity, u i is the flow velocity of the fluid entering the plunger core cavity, u k is the flow velocity of the fluid in the kidney-shaped interaction area.

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

[0250]

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

[0252]

[0253] u t = Rω

[0254] wherein, 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, 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] wherein, 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 formula, the maximum velocity calculation formula before the cavitation and erosion failure of the plunger pump can be obtained.

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

[0260] According to the basic laws of fluid transmission in the rotating component of the plunger pump, the basic control equation of the fluid domain of the rotating component is established.

[0261] The increase in the fluid mass in each grid cell in the fluid domain of the rotating component 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 component is established, and the specific expression is:

[0262]

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

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

[0265]

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

[0267] Based on the design pressure and design speed of the plunger pump, the load characteristics and motion characteristics during the use of the plunger pump are applied to the basic control equation of the fluid domain of the rotating component.

[0268] The gas volume fraction contour map of the fluid domain of the rotating component is obtained through finite element analysis to characterize the cavitation and erosion characteristics of the rotating component, 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 is the curve of the speed limit of the plunger pump under cavitation and erosion failures changing with the oil suction pressure. In addition, increasing the cavitation pressure of the medium, expanding the inlet area and the area of the kidney-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 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 the requirements are still not met after executing step S44 n times, output the highest speed of the plunger pump at the current displacement and pressure as n pmax ; the calculation formula of n pmax is:

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

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

[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, decrease the swash plate angle and reduce the plunger distribution circle radius.

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

[0275] If the maximum rotational speed of the piston pump is n pmax = n khqs , then under the basic displacement limit formula, increase the number of pistons, lower 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, distribution circle diameter, and maximum target displacement in step S21.

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

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

[0278] Structural optimization design of the piston pump rotating assembly: Through continuous cycling of the above process, obtain the structure, material, and process of the high-speed and high-pressure axial piston pump rotating assembly with a shorter 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 above-described embodiments are only descriptions of the implementation manners of the present invention, and do not limit 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. An optimized design method for the rotating assembly of a high-speed and high-pressure axial piston pump, characterized in that: Specifically, it includes the following steps: S1. Determine the design objectives and initial structural material properties of the rotating assembly of the plunger pump. The design objectives include the maximum target displacement, the highest rotational speed, and the rated pressure of the plunger pump; S2. Determine the structural parameters of the rotating assembly of the plunger pump based on the maximum target displacement; The specific steps of step S2 include: S21. Determine the cylinder block parameters, which include the diameter of the plunger hole, the pitch circle diameter, the wall thickness between the plunger holes of the cylinder block, and the length of the cylinder block; Among them, the relationship between the diameter of the plunger hole, the pitch circle diameter, and the maximum target displacement is as follows: Among them, 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; Judge whether the wall thickness between the plunger holes of the cylinder block meets the requirements by calculating the average normal stress received by the plunger holes of the cylinder block. The calculation formula for the average normal stress received by the plunger holes of the cylinder block is as follows: Among them, σ is the average normal stress of the cylinder block plunger hole, 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; The length of the cylinder block satisfies the following formula: L c = l k + 2Rtanβ+(2~3)+s+L h ; Among them, L c is the cylinder block length, l k is the minimum engagement length of the plunger, d k 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; S22. Determine the plunger parameters, which include the length of the plunger and the diameter of the plunger ball head; The length of the plunger satisfies the following formula: L z = l0 + h p + l k ; Among them, L z is the plunger length, l0 is the minimum extended length, and l0 = 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 ; The diameter d2 of the plunger ball head is (0.7 - 0.8)d k ; S23. Determine the swashplate parameters, which include the diameter of the swashplate ball socket, the inner diameter of the swashplate sealing band, the outer diameter of the swashplate sealing band, the outer diameter of the swashplate ball cup, and the diameter of the swashplate damping hole; The diameter d8 of the swashplate ball socket is equal to the diameter d2 of the plunger ball head; The inner diameter of the swashplate sealing band satisfies the following formula: Among them, d4 is the inner diameter of the shoe sealing belt, 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 shoe sealing belt, and the outer diameter of the shoe sealing belt satisfies s h = 0.2 - 1 mm; The outer diameter d7 of the slipper ball cup = (0.95 - 1)d k ; The diameter d3 of the swashplate damping hole is calculated by the method of residual clamping force; S24. Determine the return spring disc parameters, which include the pitch circle diameter of the return spring disc and the diameter of the return spring disc hole; The calculation formula for the pitch circle diameter of the return spring disc is as follows: Among them, D3 is the pitch circle diameter of the return spring disc, R is the radius of the plunger pitch circle, and β is the swashplate angle; The diameter of the return spring disc hole satisfies the following formula: d’5 = 2|ε max | + d7 + 2a min ; Among them, d’5 is the aperture diameter of the return disk, |ε max | is the radial deviation between the elliptical locus of the center of the slipper and the distribution circle of the return disk, d7 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. The outer diameter d6 of the slipper is ≥ d’5, and the outer diameter d6 of the slipper = d′6 + (1 - 3); S25. Determine the valve plate parameters and verify whether the valve plate meets the requirements. The valve plate parameters include the size of the valve plate sealing band and the deflection angle; The width of the inner and outer sealing belts of the flow distribution pair is (0.18 - 0.2)d k ; When the pressure rises from the initial pressure 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-boosting pressure, p0 is the initial pressure, E0 is the elastic modulus of the oil, V0 is the enclosed volume of the plunger chamber at the bottom dead center position, 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 radius of the plunger distribution circle, Δh is the plunger stroke, β is the swash plate angle, and is the deflected angle of the valve plate; The ratio of the radius of the center of the kidney-shaped groove of the cylinder block and the valve plate to the radius of the plunger pitch circle is 0.7 - 1.0, the width of the kidney-shaped groove is 0.35 - 0.5 times the diameter of the plunger, and the width of the inner and outer sealing bands of the valve plate is 0.1 - 0.2 times the diameter of the plunger; Judge whether the valve plate meets the requirements by calculating the linear velocity at the center of the kidney-shaped groove of the cylinder block and the valve plate. The calculation formula is as follows: Among them, V R is the center line speed of the waist-shaped groove on the flow distribution plate, and R y is the center radius of the waist-shaped groove on the flow distribution plate, and n max is the maximum rotational speed of the piston pump; S26. Determine the swashplate parameters, which include the swashplate angle and the minimum allowable distance between the two trimmed edges; The swashplate angle β is 15° - 22°; The minimum allowable distance between the two trimmed edges satisfies the following formula: B spmin = D + d’6 + 2A min ; Among them, B spmin is 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; 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 re-execute step S2; 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 re-execute step S2; S3. Establish a 3D model of the rotating assembly based on 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 structural deformation, and determine whether the rated pressure is met. If it is met, proceed to step S4; otherwise, optimize the parameters and then re - execute steps S2 - S3; The specific steps of step S3 include: S31. Establish a 3D model of the rotating assembly and initially set the materials and processes of the components; S32. Judge the structural strength of the components under the rated pressure based on the principle of structural deformation, which specifically includes the following sub - steps: S321. Divide the solid domain mesh for finite element analysis based on the structure of the rotating assembly components; 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 structures. Apply the flow field pressure load to the parts of the rotating assembly in contact with the fluid, apply the temperature load 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; 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; The force and deformation of the rotating assembly satisfy the following formula: [K]×{δ}={F}; where, [K] is the system stiffness matrix of the rotating assembly, {δ} is the system node displacement matrix of the rotating assembly, and {F} is the system force matrix of the rotating assembly; The thermal deformation of the rotating assembly 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. Extract the structural deformation and stress - strain nephogram of the rotating assembly. Judge that there is no interference in the oil film clearance of the rotating assembly after structural deformation and the structural stress is less than the allowable stress of the material. Then calculate whether the maximum pressure under the current allowable stress of the material is greater than the rated pressure. If so, proceed to the next step; if not, re - execute steps S2 and S3 after optimizing the third parameter; 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, proceed to the next step; if not, re - execute steps S2 and S3 after optimizing the fourth parameter; S4. According to the structural parameters of the piston pump rotating assembly obtained in step S2, respectively conduct parameter judgment and optimization on the faults occurring during the operation of the rotating assembly under the rated speed condition, and determine whether the maximum speed requirement is met. If it is met, proceed to step S5; otherwise, optimize the parameters and then re - execute steps S2 - S4; S5. Output the optimized structural parameters of the piston pump rotating assembly, as well as the material properties and processes of the components.

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

3. The optimized design method for the rotating assembly of a high-speed and high-pressure axial piston pump according to claim 1, wherein: If the rated pressure structure is still not met after step S33 is executed n times, output that the rated pressure requirement cannot be met under the current parameters.

4. The optimized design method for the rotating assembly of the high-speed high-pressure axial piston pump according to claim 1, characterized in that: The specific steps of step S4 include the following sub - steps: S41. Construct a formula for calculating the maximum speed before the slipper overturns to verify the slipper overturning fault of the rotating assembly; 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; If the rotational angular velocity of the plunger pump is less than or equal to the maximum velocity before the slipper tilts, then proceed to the next step and output the maximum rotational speed of the rotating assembly before the slipper tilts as n hxqf , otherwise optimize the spring force, slipper end face radius, slipper mass, plunger mass, and the distance from the slipper ball head to the centroid; after optimization, re-execute step S41, and if it still does not meet the requirements after n optimizations, then continue to execute the next step; S42. Construct a formula for calculating the maximum speed before the cylinder block overturns to verify the cylinder block overturning fault of the rotating assembly; where ω is the angular velocity of the cylinder block rotation, R g is the outer circle 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; If the rotational angular velocity of the cylinder block is less than or equal to the maximum velocity before the slipper overturns, then proceed to the next step and output the maximum rotational speed of the rotating assembly before the cylinder block overturning fault 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, then continue to execute the next step; S43. Establish the maximum speed calculation formula before the cavitation failure of the plunger pump to verify the cavitation failure of the rotating assembly: 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; If the rotational angular velocity of the plunger pump is less than or equal to the maximum velocity before cavitation occurs, then perform the next step and output that the maximum rotational speed of the rotating component before cavitation of the plunger pump is n khqs , otherwise, after optimizing the oil suction pressure, medium cavitation pressure, inlet area, and the area of the kidney-shaped interaction zone, re-perform step S43. If it still does not meet the requirements after n optimizations, then continue to perform the next step; S44. Determine the maximum speed n of the plunger pump pmax Check if it meets the requirements. If it does, proceed to the next step. If not, perform the fifth parameter optimization and then re - execute steps S2 - S4. If it still does not meet the requirements after executing S44 n times, output the highest speed of the plunger pump at the current displacement and pressure as n pmax ; n pmax The calculation formula for n is as follows: n pmax = min(n hxqf , n gtqf , n khqs ).

5. The optimized design method of the rotating assembly of the high-speed high-pressure axial piston pump according to claim 1, characterized in that: In step S432, establish the basic control equation of the fluid domain of the rotating assembly according to the basic law of fluid transmission in the rotating assembly of the plunger pump.

6. The optimized design method of the rotating assembly of the high-speed 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 assembly; S4322. Establish the Navier-Stokes equation of the compressible fluid of the rotating assembly; S4323. Based on the design pressure and design speed of the plunger pump, 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; S4324. Obtain the gas volume fraction cloud map of the fluid domain of the rotating assembly through finite element analysis to characterize the cavitation characteristics of the rotating assembly.

7. The optimized design method for the rotating assembly of the high-speed and high-pressure axial piston pump according to claim 1, wherein: In step S24, the radial deviation |ε| between the elliptical locus of the center of the slipper and the distribution circle of the return disk max is as follows: where, |ε max | is the radial deviation between the elliptical locus 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.

8. The optimized design method for the rotating assembly of the high-speed high-pressure axial plunger pump according to claim 1, wherein: 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; 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; The specific steps of the third parameter optimization in step S33 are to increase the structural strength and structural 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; The specific steps of the fourth parameter optimization in step S34 are to reduce the structural strength and structural stiffness of the overall material; The specific steps of the fifth parameter optimization in step S44 are: If the maximum speed of the piston pump is n pmax = n hxqf , then under the basic displacement limit formula, reduce the number of pistons, decrease the swash plate angle, and lower the piston distribution circle radius; If the maximum speed of the piston pump is n pmax = n gtqf , then under the basic displacement limit formula, reduce the number of pistons, lower the swash plate angle, reduce the piston distribution circle radius, and increase the cylinder block radius; If the maximum rotational speed of the piston pump is n pmax = n khqs , then under the basic displacement limit formula, increase the number of pistons, decrease the swash plate angle, and decrease the piston distribution circle radius. Among them, the basic displacement limit formula is the relational expression between the piston hole diameter, the distribution circle diameter, and the maximum target displacement in step S21.

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