A performance-driven multi-objective optimization method for a plunger pump valve plate

Abstract

A performance-driven plunger pump flow distribution plate multi-objective optimization method is applied to the vibration and noise reduction of an axial plunger pump. The fluid vibration power and mechanical excitation power are obtained through a dynamic model. The fluid dynamics model of the plunger cavity is built through the control volume method to calculate the fluid vibration power. Based on the structural analysis of the plunger pump, the system response is solved through the Newmark-beta method, and then the structural dynamics model is built to calculate the mechanical vibration power. The vibration power is converted to the frequency domain through fast Fourier transform, the root mean square of a specific harmonic is taken as the target, and the optimal flow distribution plate structure is obtained through multi-objective genetic algorithm optimization. The traditional optimization method cannot fully consider the vibration transmission relationship of the plunger pump, and cannot achieve the purpose of accurate amplitude optimization. The optimization method can solve the two problems, and is conducive to the performance stability of the optimized plunger pump.

Description

Technical Field

[0001] This invention relates to the field of vibration reduction and noise reduction of plunger pumps, and more particularly to a method for multi-objective optimization of the distribution plate of a performance-driven plunger pump. Background Technology

[0002] Hydraulic transmission offers advantages over electrical and mechanical transmissions, including greater compactness, higher reliability, and better controllability, leading to its widespread application in automotive, industrial, and aerospace fields. Axial piston pumps, due to their superior reliability, power density, and stable operation at high pressures, are frequently used as power sources in hydraulic transmission systems. However, the high pressure and large displacement of piston pumps result in high noise levels, making them a major noise source in fluid transmission systems. Furthermore, with increasing environmental awareness in recent years, the requirements for reducing permissible sound power levels in machinery are becoming more stringent, making vibration and noise reduction of piston pumps an important research direction.

[0003] The noise generated during the operation of a piston pump can be divided into fluid noise and structural noise. Fluid noise originates from fluid pulsation, generated by the limited number of pistons and the compressibility of the fluid. Structural noise originates from the internal forces and torques generated by pressure changes in the piston chamber. Over the past few decades, much research has been conducted to reduce the noise level of piston pumps, such as optimizing the casing to reduce externally radiated noise; using high-damping check valves to reduce flow pulsation in axial piston pumps; and using directional valves to control valve opening, which can be adjusted according to actual operating conditions. These methods have proven effective, but they require additional mechanisms in the hydraulic transmission system, which not only affects the stability of the hydraulic system but also increases costs. Compared to adding external mechanisms to reduce noise, modifying the internal component structure of the piston pump is more promising. Research has shown that the internal distribution plate structure of the piston pump has a significant impact on pump vibration noise.

[0004] Currently, research on the distribution plate structure is relatively mature. Numerous studies have used virtual simulation technology to build theoretical models of plunger pumps and analyze the impact of the distribution plate structure on flow turbulence. Research has found that the geometry of the damping groove in the transition region significantly affects the flow stability of the oil inside the plunger pump, leading to severe oil pulsation, generating excitation forces and torques, ultimately resulting in excessive noise levels. The rationality of the damping groove structure in the transition region of the distribution plate is crucial for vibration reduction and noise reduction in the plunger pump; therefore, the damping groove structure in the transition region of the distribution plate has always been an important optimization target. In past studies, researchers have achieved the goal of reducing flow pulsation and excitation forces and torques by optimizing the damping groove structure. However, optimization objectives are interdependent. Traditional optimization methods may result in too many optimization variables leading to poor convergence and insufficient consideration of optimization objectives. Furthermore, previous optimizations were based on the pressure pulsation at the inlet and outlet in the time domain and the resulting forces and torques, without considering the transmission relationship of vibrations between mechanisms. Therefore, optimization can only guarantee a reduction in the pulsation of the excitation source, but cannot directly ensure a reduction in vibration velocity and noise levels. In the research on vibration reduction and noise reduction of plunger pumps, more attention is paid to the amplitude of low-order harmonics in the frequency domain. Although time-domain optimization can reduce the fluctuation range of pulsation, it cannot guarantee the reduction of low-order amplitude after optimization, which affects the evaluation of subsequent optimization effects. Summary of the Invention

[0005] The purpose of this invention is to solve the above-mentioned problems in the prior art and provide a multi-objective optimization method for a performance-driven plunger pump distribution plate. This invention takes fluid vibration power as optimization objective 1, fully considering the influence of fluid pulsation on fluid vibration noise; and takes mechanical vibration power as optimization objective 2, fully considering the influence of plunger cavity pressure change on structural vibration noise. The entire optimization process of this invention is carried out in the frequency domain, which can achieve accurate optimization of the specific harmonic amplitude of the optimization objective.

[0006] The present invention will achieve the objective of the above-mentioned optimization method in the following manner:

[0007] Firstly, to obtain optimization objective 1, this invention utilizes computer software simulation technology to construct a fluid dynamics model for solving the plunger cavity pressure and fluid vibration power, including:

[0008] 1) By analyzing the positional relationship between the damping groove and the cylinder bore, the liquid fluid area is calculated analytically.

[0009] 2) By analyzing the relationship between the piston and cylinder movement during piston pump operation, the pressure change in the piston chamber is calculated using the control volume method.

[0010] Furthermore, by using a simulation model, the flow and pressure changes at the inlet and outlet ports during the piston-cylinder movement are comprehensively analyzed to obtain the flow and pressure pulsations at the inlet and outlet ports, and then the optimization objective 1 is calculated.

[0011] Furthermore, through a comprehensive analysis of the mechanics of the internal components of the plunger pump, the pressure results of the plunger cavity obtained by simulation technology were used to calculate the excitation force and torque using analytical methods, thus preparing for the construction of the structural dynamics model.

[0012] Secondly, in order to obtain optimization objective 2, the computer simulation technology of this invention constructs a multi-degree-of-freedom structural dynamics model, including:

[0013] 1) Parameterize the main components of the axial piston pump to obtain the key parameters of the structural dynamics model;

[0014] 2) The main components are reasonably simplified, and the motion coordinate system of the simplified plunger pump components is established. The simplified model includes 5 parts: the housing motion coordinate system O H -X H Y H Z H Swashplate motion coordinate system O S -X S Y S Z S The fixed connection of the plunger, slipper, and return disc assembly is simplified to a single component with the motion coordinate system O. P -X P Y P Z P The cylinder block and spindle are fixedly connected and simplified into one component, with the motion coordinate system being O. C -X C Y C Z C The end cap and distributor plate are fixedly connected and simplified into one component, with the motion coordinate system being O. EC -X EC Y EC Z EC ;

[0015] 3) Perform dynamic analysis on the main components of the axial piston pump assembly, and perform coupled dynamic analysis between the components to obtain the differential equations of the axial piston pump system;

[0016] 4) Based on the established structural dynamics differential equations, the stiffness-damping matrix of the system is defined by the main excitation forces obtained from the system differential equations and fluid dynamics model;

[0017] 5) The excitation force and torque act periodically on the axial piston pump. The system response is solved using the Newmark-β method to obtain the dynamic characteristics of the axial piston pump.

[0018] Furthermore, the vibration velocity between the axial piston pump components is calculated using a dynamic model, and then the mechanical vibration power is calculated by combining the main excitation force and torque.

[0019] Thirdly, the present invention provides computer-integrated simulation software that can implement the model building method of the first aspect; and non-computer-executable program code that can execute the model building method described in the second aspect.

[0020] Fourthly, the present invention also provides a data interaction tool that can comprehensively analyze information such as optimization variables, constraints, and optimization objectives in the dynamic model established in the first and second aspects, and automatically optimize the objective through optimization algorithms to find the optimal solution.

[0021] Compared with the prior art, the beneficial effects achieved by the technical solution of this invention are:

[0022] 1. This invention fully considers the effects of fluid vibration and structural vibration, and the number of optimization objectives does not increase. It has good convergence in the algorithm optimization process, avoiding the poor convergence caused by insufficient consideration of optimization objectives and too many objectives in traditional methods.

[0023] 2. This invention optimizes the amplitude of low-order harmonics in the target frequency domain, which can accurately reduce the amplitude of low-order harmonics that are of greater concern in the field of vibration and noise. Compared with traditional optimization that optimizes the amplitude of pulsation in the time domain, the results are more reasonable.

[0024] 3. In the process of model building, this invention involves the connection and coupling relationship between plunger pump components, and uses vibration power as the optimization target to more intuitively illustrate the impact of changes in the distribution plate structure on the overall vibration and noise of the plunger pump. Attached Figure Description

[0025] Figure 1 This is a flowchart of the optimized method of the present invention.

[0026] Figure 2 This is a schematic diagram of the distribution plate structure.

[0027] Figure 3 This is a schematic diagram of a fluid dynamics model.

[0028] Figure 4 This is a schematic diagram of the structural dynamics model.

[0029] Figure 5 The results of multi-objective optimization of the plunger pump distribution plate. Detailed Implementation

[0030] To make the technical problems, technical solutions and beneficial effects of the present invention clearer and more understandable, the present invention will be further described in detail below with reference to the accompanying drawings and embodiments.

[0031] This invention proposes a novel method for multi-objective optimization of the distribution plate of a performance-driven plunger pump, which can be applied to the field of plunger pump vibration reduction and noise reduction, and can achieve a reduction of a specific amplitude of the optimization target in the frequency domain.

[0032] Figure 1 The flowchart of the multi-objective optimization method based on the distribution plate of the driven plunger pump mainly includes the following steps:

[0033] P1. By analyzing the structure of the plunger pump's distribution plate, such as... Figure 2 As shown. The optimization variable is defined as the transition region T of the distribution panel. W and S W The damping groove angle includes the following parameters:

[0034] 1) The center angles of damping orifice 1 and damping orifice 2 and The radii r1 and r2 of the two holes;

[0035] 2) Starting angle of the oil inlet waist-shaped damping groove Oil inlet waist-shaped damping groove termination angle

[0036] 3) Starting angle of the oil outlet triangular damping groove Oil outlet waist-shaped damping groove termination angle

[0037] 4) Width angle of the triangular damping groove Triangular damping groove depth angle

[0038] P2. Simplify the axial piston pump and parameterize the structure of its main components to obtain the key parameters for building the dynamic model. Simultaneously, initialize the optimization algorithm and define the optimization range of the variables.

[0039] P3. Analyze the pressure change in the plunger cavity using the control volume method, and then build a fluid dynamics model to calculate the optimization objective 1: fluid vibration power. Based on the excitation force of the fluid dynamics model, analyze the coupling dynamics between the plunger pump components, solve the system response using the Newmark-β method, complete the construction of the structural dynamics model, and calculate the optimization objective 2: mechanical vibration power.

[0040] P4. Using a multi-objective optimization algorithm to find the optimal target, ensuring that the vibration power frequency domain harmonic amplitude is minimized while meeting the constraints, the optimal variable combination scheme can be obtained.

[0041] Figure 3 This is a schematic diagram of a fluid dynamics model. As the cylinder rotates around the main shaft, the plunger reciprocates within the cylinder bore, causing periodic changes in the plunger cavity volume. Oil enters the plunger cavity through the damping grooves of the distribution plate and the cylinder bore. When the change in the plunger cavity volume does not match the change in oil flow rate, it will cause a change in the plunger cavity pressure, which can be expressed by the following formula:

[0042]

[0043] In the formula, p represents the pressure inside the piston chamber, βe is the bulk modulus of the hydraulic oil, Vp is the volume of the piston chamber, and Q... L With Q H These are the flow rates between the piston chamber and the intake and exhaust ports, respectively, Q. X It is a leak, where t is the transient time.

[0044] Based on parametric analysis of the oil flow area, and assuming the inlet and outlet are turbulent flows through throttling orifices, Q can be calculated. L With Q H As shown below:

[0045]

[0046]

[0047] Among them, C o ρ is the flow coefficient, ρ is the oil density, and p L and p H The inlet and outlet pressures are directly related to the plunger chamber pressure, respectively. L and A H This refers to the flow area of ​​the oil inlet and outlet.

[0048] By analyzing the positional relationship between the cylinder block and the plunger during the movement of the plunger pump, the displacement of the plunger during its movement from position ODC to point A can be expressed as L. W =Rsinθtanδ0, where R is the plunger distribution radius, θ is the plunger rotation angle, and γ is the swashplate tilt angle. The change in plunger cavity volume can then be calculated:

[0049] V p =V p0 -πr 2 L W

[0050] In the formula, Vp0 is the volume of the plunger cavity when the plunger is in the ODC position, and r is the plunger radius.

[0051] When building a fluid dynamics model, the impact of leakage needs to be considered. There are three main types of leakage inside a plunger pump: leakage between the plunger and cylinder assembly (Q). XC Leakage between swashplate and slipper assembly Q XS Leakage between cylinder block and distributor plate assembly Q XV Total leaks amount to Q. X =Q XC +Q XS +Q XV The three main forms of leakage can be calculated using the following formula:

[0052]

[0053]

[0054]

[0055] In the formula, δc is the oil film thickness between the piston and the cylinder, μ is the dynamic viscosity of the hydraulic oil, ε is the piston's eccentricity relative to the piston chamber, and p h It is the pressure inside the pump casing, l p It is the contact length between the piston and the piston chamber; δ S R represents the oil film thickness of the slipper pair. st and r st These are the outer and inner radii of the slipper surface, respectively; δ V Let R1 and R2 be the inner and outer radii of the inner sealing strip of the valve disc, respectively, and R3 and R4 be the inner and outer radii of the outer sealing strip of the valve disc, respectively. α f This is the wrap angle of the piston chamber.

[0056] In summary, the plunger chamber pressure, inlet and outlet oil flow velocities and pressures during the operation of a single plunger can be obtained, and thus the fluid vibration power P can be obtained. q :

[0057] P q =p H Q H +p L Q L

[0058] The coupled dynamics model established in this study is as follows: Figure 4 As shown, the model is built based on ideal conditions, therefore some assumptions are required:

[0059] 1) The complex interactions between the housing and swashplate, the slipper and swashplate, the cylinder and piston, and the distributor plate and cylinder are simplified to massless spring-damped elements, expressed by the corresponding stiffness and damping coefficient.

[0060] 2) The bearing is modeled using a linear stiffness model. The bearing connection is replaced by a spring-damped design with stiffness damping value. The bolts are also replaced by springs with stiffness equivalent to those in the X, Y and Z directions.

[0061] 3) In this invention, since the internal rotational excitation force is not considered, it is assumed that the rotation around Z... H -axis, Z S -axis, Z C - Axis and Z EC - The shaft does not rotate, and it is assumed that the assembly of the cylinder block and the shaft rotates around Z. P - The rotational speed of the shaft is constant.

[0062] 4) All simplified assemblies described based on the mass point (LMP) were measured using 3D software to determine their mass and moment of inertia in the translational and rotational directions.

[0063] Based on the above assumptions, the axial piston pump is simplified into a model consisting of 5 lumped mass points (LMPs). The motion coordinate system of the simplified piston pump model is established, including: the housing motion coordinate system O. H -X H Y H Z H The lumped mass point is H LMP ; Swashplate motion coordinate system O S -X S Y S Z S The lumped mass point is S LMP The fixed connection of the plunger, slipper, and return disc assembly is simplified to a single component with the motion coordinate system O. P -X P Y P Z P The lumped mass point is P LMP The cylinder block and spindle are fixedly connected and simplified into one component, with the motion coordinate system being O. C -X C Y C Z C The lumped mass point is C LMP The end cap and distributor plate are fixedly connected and simplified into one component, with the motion coordinate system being O. EC -X EC Y EC Z EC The lump mass point is EC LMP .

[0064] To more specifically explain the connection relationships between components in the structural dynamics model, a more detailed simplification is provided: In reality, the plunger pump housing and swashplate are connected via bearings, which is equivalent to a spring-damped connection in the dynamics model. The direction of the spring-damped element is perpendicular to the swashplate surface and passes through the center of the swashplate sphere; the contact between the slipper and the swashplate is relatively smooth, and the coefficient of friction between their contact surfaces is small. Therefore, in this modeling process, the slipper and the swashplate move along the X-axis... P axis and Y P The shaft stiffness-damping is set to 0, and only set along the Z-axis. P Stiffness-damping in the direction; the contact between the cylinder block and the cylinder bore is also relatively smooth, and the vibration generated by the main excitation force is along the Z-axis. C Directional transmission, therefore, in the model, the contact between the plunger and the cylinder bore is simplified to be along the Z-axis. CDirectional stiffness-damping; Due to the nature of this invention, the relative rotational motion between the plunger and the slipper is ignored, and because the plunger pump operates at a low speed, the gyroscopic effect is disregarded, treating the plunger, slipper, and return plate as a single unit; the interaction between the distributor plate and the cylinder block is equivalent to four-point contact, uniformly distributed along the axis centerline with a radius of R. VP On the circle; in the process of building this dynamic model, the bolt is equivalent to the spring element, and the shell and the base, and the shell and the end cover are connected by bolts.

[0065] Based on the plunger cavity pressure calculated using the fluid dynamics model, the excitation force can be calculated. This excitation force will be used as the main excitation input to the lumped mass point P in the coupled dynamics model. LMP and C LMP The action at mass point C LMP The force and torque are calculated using the following formula:

[0066]

[0067]

[0068]

[0069] The action on mass point P LMP The force and torque are calculated using the following formula:

[0070]

[0071]

[0072]

[0073] In the formula, F CZ It acts on the concentrated mass point C LMP The force on M CX and M CY It acts on C LMP The torque on the piston, d is the diameter of the piston, p i θ is the pressure in the i-th piston chamber, R is the pitch radius, and θ is the pressure in the i-th piston chamber. i F is the rotation angle of the i-th plunger. PZ It acts on P LMP The force on M PX and M PY It acts on the concentrated mass point P LMP The torque on the swashplate, δ0 is the tilt angle of the rotating swashplate.

[0074] Based on the above assumptions and the calculated excitation force, a dynamic analysis of the five concentrated components of the axial piston pump is performed. By conducting a coupled dynamic analysis, the dynamic differential equations of each component of the axial piston pump can be obtained.

[0075] The main objective of this invention is to propose a new optimization method for distribution disks. Therefore, it will not provide a more detailed explanation of all the formulas in the dynamic model, but will only explain the main formulas used in solving the optimization objective 2: mechanical vibration power.

[0076] The plunger, slipper, and return disc assembly are arranged along the Y-axis. P and Z P The equation for directional translation can be expressed as:

[0077]

[0078] The plunger, slipper, and return disc assembly revolves around axis X. P -and Y P The equation for rotation in the - direction is:

[0079]

[0080] in: N is the number of plungers, ω0 is the principal axis angular velocity, and t is time. L SP Lumped mass point S LMP With P LMP The distance along the Z-axis is R, which is the distance from the main spindle centerline to each plunger. M P For the mass of the plunger, slipper, and return disc assembly; the assembly along axis X P -、Y P -and Z P - The translational displacement in the direction is represented by X. P Y P and Z P I Px I Py The moment of inertia of the plunger, slipper, and return disc; the rotational angles of the assembly about the XP and YP axes are represented by θ. Px and θ Py M S The mass of the swashplate assembly; the swashplate assembly along axis X S -、Y S -and Z S - The translational displacement in the direction is represented by X. S Y S and Z S , around axis X S -and Y S - The rotation angle of the axis is expressed as θ Sx and θ Sy K SW and C SW K represents the equivalent stiffness-damping parameter of the connection between the shoe and the swashplate contact surface. H and C HK represents the equivalent stiffness-damping parameter of the connection between the shell and the swashplate support contact surface. H C H These are the equivalent stiffness and damping between the plunger and the plunger cavity, respectively.

[0081] Cylinder block-spindle assembly along axis X C -、Y C -and Z C The three equations for the direction are as follows:

[0082]

[0083] Cylinder block-spindle assembly around axis X C -and Y C The equations for rotation in the - directions are as follows:

[0084]

[0085] In the formula, M C Mass of the spindle-cylinder assembly; the mass of the spindle and cylinder assembly along axis X. C -、Y C -and Z C - The translational displacement in the direction is represented by X. C Y C and Z C I Cx I Cy Main spindle-cylinder assembly around axis X C -and Y C - Moment of inertia of rotation; component about X C -and Y C - The rotation angle of the axis is expressed as θ Cx and θ Cy M EC For the mass of the distributor plate-end cap assembly; the distributor plate-end cap assembly along axis X EC -、Y EC -and Z EC - The translational displacement in the direction is represented by X. EC Y EC and Z EC , around X EC and Y EC The rotation angle of the axis is expressed as θ. ECx and θ EC y;K sh2x C sh2x These are the equivalent stiffness-damping parameters for the connection between the end cap and the main shaft; K sh1z C sh1z These are the equivalent stiffness and damping parameters of the connection between the housing and the main shaft; L C1 The position of the equivalent spring damping element at the connection between the spindle and the housing is relative to the concentrated mass point C.LMP Along Z C - Projected length of the distance along the axial direction; L C2 For lumped mass point C LMP The distance between the equivalent spring-damping element at the connection between the end cap and the spindle along the Z-axis C Projection along the - axis. L EC2 For the lumped mass point EC LMP The distance between the equivalent spring-damping element at the connection between the end cap and the spindle along the Z-axis EC Projection along the - axis. L H3 The location of the end cap and housing connecting bolts to the point of concentrated mass EC LMP Along Z EC - The projected length of the distance along the axial direction. X H Z H Y H For the housing along axis X H -、Z H -and Y H - displacement distance, around X H and Y H The rotation angle of the axis is expressed as θ. Hx and θ Hy K VP C VP R represents the equivalent stiffness and damping of the contact surfaces of the cylinder block and the distributor plate, respectively. VP K is the distance from the centerline of the main shaft to the contact point between the distributor plate and the cylinder block's equivalent stiffness damping. sh2y C sh2y The bearing connecting the end cover and the main shaft is located along the Y-axis. C- Stiffness and damping in the axial direction.

[0086] Based on the analysis of the dynamic equations, the dynamic equations of the axial piston pump can be expressed as:

[0087]

[0088] In the formula, M, C and K are the mass, damping and stiffness matrices, respectively; q and q represent the initial acceleration, velocity, and displacement, respectively; F is the excitation force.

[0089] q and F are represented as follows:

[0090] q={X H ,Y H Z H ,θ Hx ,θ Hy ,Y S Z S ,θ Sx ,θ Sy ,Y P ZP ,θ Px ,θ Py ,X C ,Y C Z C ,θ Cx ,θ Cy ,X EC ,Y EC Z EC ,θ ECx ,θ ECy} T F = {0,0,0,0,0,0,0,0,0,0,0,F} Z1 M X1 M Y1 ,0,0,F Z2 M X2 M Y2 ,0,0,0,0,0} T

[0091] In the coupled dynamic model of the plunger pump structure established in this study, the excitation force mainly acts on the concentrated mass point P. LMP and C LMP This is transmitted through the coupling connections between components. Therefore, in calculating the mechanical vibration power P... v The power radiated outward from the center of excitation force is the primary factor:

[0092]

[0093]

[0094] P v =P P +P C

[0095] In the formula, P P P represents the power radiated outward from the plunger, slipper, and return disc assembly. C F represents the power radiated outward from the cylinder block and spindle assembly. PZ F CZ T represents the axial force acting on the plunger, slipper, and return plate assembly; PX T PY T represents the torque acting in the X and Y directions of the plunger, slipper, and return plate assembly. CX T CY This represents the torque acting on the cylinder block and spindle assembly in the X and Y directions; and These represent the vibration velocity in the Z direction and the angular velocity of the plunger, slipper, and return plate in the X and Y directions, respectively. and These represent the vibration velocity in the Z direction and the angular velocity of the cylinder block and spindle assembly around the X and Y directions, respectively.

[0096] In summary, fluid vibration power and mechanical vibration power can be calculated using the established dynamic model. The vibration power in the time domain is converted to the frequency domain using the Fast Fourier Transform (FFT), and to make the target more representative, this method selects the root mean square of the first three amplitudes of the vibration power as the optimization objective.

[0097]

[0098]

[0099] Among them, P q,k (jω) represents the peak value of the kth harmonic of the fluid vibration power in the frequency domain, P v,k (jω) represents the peak value of the kth harmonic of the mechanical vibration power in the frequency domain, f1(Ф) represents the root mean square value of the peak values ​​of the first 3 fluid vibration power, and f2(Ф) represents the root mean square value of the peak values ​​of the first 3 mechanical vibration power.

[0100] To prevent severe cavitation and oil vibration during operation of the optimized distribution plate, the upper and lower limits of the plunger cavity pressure need to be constrained in the time domain during the optimization process. Two constraints are set: 1) P min ≥0.3MPa; 2)P max -P ave ≤1.5MPa, where P min P is the minimum pressure in the plunger cavity. max P is the maximum pressure in the plunger cavity. ave This represents the average pressure at the oil outlet.

[0101] After redesigning the damping groove of the distribution plate using the optimization method of this invention, the change in vibration power is as follows: Figure 5 As shown in the figure. It can be seen from the figure that the mechanical vibration power of the optimized distribution plate in the frequency domain ( Figure 5 c) and fluid vibration power ( Figure 5 d) The amplitudes of the first six harmonics are reduced compared to before optimization, with a significant reduction in the amplitudes of the first three harmonics. After optimization, the mechanical vibration power in the time domain ( Figure 5 a) and fluid vibration power ( Figure 5 b) The fluctuation range is reduced, and the plunger pump is more stable.

[0102] This invention provides computer-integrated simulation software that can implement the above-mentioned method for building fluid dynamics models; and non-destructive computer-executable program code that can execute the above-mentioned method for building structural coupled dynamics models, realizing the optimization target's conversion in the time domain and frequency domain.

[0103] This invention also provides a data interaction tool that can comprehensively analyze information such as optimization variables, constraints, and optimization objectives in fluid dynamics models and mechanical dynamics models, and automatically optimize the objective through optimization algorithms to find the optimal solution.

[0104] The optimization method of this invention reduces the vibration noise of the plunger pump by precisely reducing the amplitude of low-order harmonics in the important frequency domain, while also achieving a reduction in the target amplitude in the time domain. Furthermore, the optimization objective not only considers the influence of the excitation source but also links the coupling relationship between the excitation source and components through a dynamic model, allowing for direct reduction of the axial plunger pump vibration power.

Claims

1. A method for multi-objective optimization of a performance-driven plunger pump distribution plate, characterized in that... Includes the following steps: 1) A fluid dynamics model of the plunger cavity is built using the control volume method, and the vibration power of the oil inlet and outlet is solved. The fluid vibration power is obtained by summing the values ​​of the two, and the fluid vibration power is used as one of the optimization objectives. 2) The system response was solved using the Newmark-β method, and then a structural dynamics model was built to calculate the lumped center of mass P of the plunger, slipper, and return disk assembly. LMP The radiated power, and the centered mass C of the cylinder block and spindle assembly LMP The radiated power is calculated, and the sum of the two values ​​is taken as the mechanical vibration power, which is then used as one of the optimization objectives. 3) The fluid vibration power and mechanical vibration power are converted to the frequency domain by fast Fourier transform, and the root mean square of the first three amplitudes is taken as the optimization target. 4) The optimal distribution disk structure is obtained by using a multi-objective genetic algorithm to optimize the objective. In step 1), the equation for solving the pressure in the plunger cavity is as follows: Where p is the pressure inside the plunger cavity, β e V is the bulk modulus of hydraulic oil. p Let Q be the volume of the plunger cavity. L With Q H Q represents the flow rate between the plunger chamber and the intake and exhaust ports, respectively. X For leakage, t is the transient time; Fluid vibration power P q The calculation formula is: Where, p H and p L These represent export pressure and import pressure, respectively. H and Q L These are export flow and import flow, respectively. In step 2), the differential equations of motion for the piston pump coupled system are as follows: Where M, C, and K are the mass, damping, and stiffness matrices, respectively; , and These represent the initial acceleration, velocity, and displacement, respectively; F is the excitation force matrix. Mechanical vibration power P v The calculation formula is: Among them, P P P is the vibrational power radiated outward from the plunger, slipper, and return disc assembly. C The vibration power radiated outward from the cylinder block and spindle assembly; In step 3), the expression for the optimization objective in the frequency domain is as follows: Among them, P q,k (jω) represents the peak value of the kth harmonic of the fluid vibration power in the frequency domain, P v,k (jω) represents the peak value of the kth harmonic of the mechanical vibration power in the frequency domain, f1(Ф) represents the root mean square value of the peak values ​​of the first 3 fluid vibration power, and f2(Ф) represents the root mean square value of the peak values ​​of the first 3 mechanical vibration power.

2. The method for multi-objective optimization of a performance-driven plunger pump distribution plate as described in claim 1, characterized in that: To avoid severe pressure pulsation and cavitation during plunger pump operation, optimized constraints need to be set for the upper and lower pressure limits of the plunger chamber.

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