Design method of electric vehicle power assembly rubber suspension system based on multi-objective optimization

Abstract

The invention discloses a design method of an electric vehicle power assembly rubber suspension system based on multi-objective optimization, and relates to the technical field of electric vehicle power assembly suspension. The method comprises the steps that Matlab and Isight software are combined to optimize the power assembly suspension system, an NSGA-II algorithm in the Isight software is used for optimizing the static stiffness of each suspension element, static stiffness parameters are reasonably adjusted, simulation analysis is conducted on the optimized suspension system, and the reasonability and feasibility of the optimized suspension system are verified; the rigidity curve of each suspension element is subjected to nonlinear design, the displacement of six degrees of freedom at the mass center of the power assembly is calculated by using Adams / View software according to the general 28 working condition, the rationality of the nonlinear rigidity curve of each suspension element in the U direction, the V direction and the W direction is verified, and the load of each suspension element is calculated.

Description

Technical Field

[0001] The present invention relates to the technical field of electric vehicle powertrain mounts, and particularly to a design method for a rubber mount system of an electric vehicle powertrain based on multi-objective optimization. Background Art

[0002] With the continuous development of electric vehicles, consumers' requirements for their performance are also increasing day by day. The NVH (Noise, Vibration, Harshness) performance of electric vehicles is the most significant vehicle dynamic quality that passengers can perceive. External excitations and internal vibrations and noises are the main vibration sources of pure electric vehicles. External excitations are mainly transmitted into the vehicle through channels such as the body structure, air, and suspension system, affecting the NVH performance of the vehicle. For example, when the vehicle is driving on an uneven road surface, vibrations and noises will be generated when the suspension system, tires, and the body interact with the air. In addition, when the vehicle is driving normally and the motor is working properly, mechanical vibrations of the motor stator and axial eccentricity between the stator and the rotor will cause vibrations and noises in the motor.

[0003] The powertrain (drive motor) of a pure electric vehicle is one of the main excitation sources of vibration and noise in electric vehicles. Compared with the powertrain of traditional fuel vehicles, the powertrain of pure electric vehicles has characteristics such as a wide speed regulation range, low mass inertia torque, and different vibration excitations. At the same time, pure electric vehicles do not have an idling condition. Therefore, the design methods and standards of the powertrain mount system for traditional fuel vehicles are no longer fully applicable to the design of the powertrain mount system for pure electric vehicles. Therefore, according to the structure and working characteristics of the powertrain of pure electric vehicles, reasonably designing its powertrain mount system has become a key technology for reducing the noise and vibration inside the vehicle and improving the ride comfort of the whole vehicle.

[0004] After retrieval, Chinese Patent CN112733265A discloses a design calculation and optimization method for an electric vehicle powertrain mount system, including the following steps: First, a centroid coordinate system of the powertrain is established, and on this basis, parameters such as the inertial parameters of the powertrain, the centroid position, and the installation position and static stiffness curve of the rubber mount are obtained; further, a dynamic model of the powertrain mount system is established, and the differential equation of motion is derived; an objective function is established according to the energy decoupling theory and transient response characteristics; selecting the highest energy decoupling rate of the mount system, the minimum longitudinal acceleration of the powertrain centroid under transient response, and the minimum impact degree amplitude as the objectives, using the linear segment stiffness of the mount as the design variable, and the reasonable distribution of the natural frequency and the change range of the mount stiffness as the constraints, and using a multi-island genetic algorithm for optimization. Finally, the feasibility of this method is verified through an example.

[0005] However, when using the elastic axis decoupling method to optimize the stiffness, there are problems such as insufficient decoupling rate, small adjacent modal frequency spacing, and inability to effectively avoid resonance. In addition, traditional design methods rely mostly on single software simulation and do not combine multidisciplinary joint optimization, resulting in a long design cycle and limited accuracy. Summary of the Invention

[0006] 1. Problems to be Solved

[0007] In view of the technical problems existing in the prior art, the present invention provides a design method for a rubber mounting system of an electric vehicle powertrain based on multi-objective optimization. This method combines the modal energy decoupling method with the NSGA-II algorithm to achieve the optimization of the natural frequency of the mounting system and the design of non-linear stiffness, and improve the vibration isolation performance and spatial motion control ability.

[0008] 2. Technical Solutions

[0009] To solve the above problems, the technical solutions adopted by the present invention are as follows:

[0010] In the first aspect of the present invention, a design method for a rubber mounting system of an electric vehicle powertrain based on multi-objective optimization is provided. The rubber mountings of the electric vehicle powertrain include a front mounting cushion assembly, a left mounting cushion assembly, and a right mounting cushion assembly (hereinafter referred to as mounting components). Among them, the mounting rubber main springs of the front mounting cushion assembly, the left mounting cushion assembly, and the right mounting cushion assembly have the same structure and material. The mounting rubber main spring includes an X-shaped bushing main spring, a metal outer sleeve, and a metal inner core. The method includes the steps of: jointly optimizing the powertrain mounting system using Matlab and Isight software, using the NSGA-II algorithm in Isight software to optimize the static stiffness of each mounting component, reasonably adjusting the static stiffness parameters, and performing simulation analysis on the optimized mounting system to verify the rationality and feasibility of the optimized mounting system; performing non-linear design on the stiffness curves of each mounting component, and using Adams / View software to calculate the displacements of six degrees of freedom at the centroid of the powertrain according to the general 28 working conditions, verifying the rationality of the non-linear stiffness curves of each mounting component in the U, V, and W directions, and calculating the loads of each mounting component.

[0011] The design method for the rubber mounting system of the electric vehicle powertrain based on multi-objective optimization specifically includes the following steps:

[0012] f. Establishing a mathematical model of the powertrain mounting system:

[0013] g. Establishing a model of the powertrain mounting system: including modal analysis of the powertrain mounting system and decoupling analysis of the powertrain mounting system;

[0014] h. Power-train mounting system simulation and optimization: The power-train mounting system is optimized by combining Matlab software and Isight software to improve the vibration isolation performance of the system;

[0015] i. Nonlinear stiffness design of the mounting system:

[0016] j. 28-condition simulation analysis of the power-train mounting system.

[0017] According to any implementation of the first aspect of the object of the present invention, in step (a), the power-train is simplified into a rigid body, each mounting element is regarded as a massless spring body with mutually perpendicular principal axis directions, the mounting system is simplified into a six-degree-of-freedom model, and the motion differential equation of the simplified model of the power-train is established;

[0018]

[0019] Where: Qi is the generalized coordinate of the mounting system;

[0020] T is the kinetic energy of the mounting system during vibration;

[0021] U is the potential energy of the mounting system during vibration;

[0022] D is the dissipated energy of the mounting system during vibration;

[0023] F i Excitation force.

[0024] According to any implementation of the first aspect of the object of the present invention, in step (c), an Isight and Matlab software combined simulation framework is built, the Matlab module is opened, input variables and output variables are established respectively, where the input variables are the stiffnesses of each mounting element, and the output variables are the modal frequencies of each order of the power-train mounting system, the interval between adjacent modal frequencies, and the energy decoupling rate in each excitation direction. The Matlab module is set by importing the written Matlab file and setting the Matlab executable file to start;

[0025] The Optimization module is opened, the ranges and optimization objectives of the input and output variables are set, the NSGA-II optimization algorithm is selected, and the parameters of the optimization algorithm are set accordingly; the built combined simulation framework is run to obtain the optimization process diagram of each optimization parameter;

[0026] After optimization, the optimal solution of the parameters of the optimized model of the mounting system output by the Isight software.

[0027] According to any embodiment of the first aspect of the object of the present invention, in step (d), when designing the non-linear stiffness of the suspension element, a multi-segment linear line segment with three segments (bc - cd - de) or five segments (ab - bc - cd - de - ef) is used to replace the non-linear stiffness curve, and the design method of the non-linear stiffness curve is judged according to the stiffness difference between adjacent line segments. If the stiffness difference is small, a three-segment line is used to design the non-linear stiffness curve of the suspension element; otherwise, an additional transition segment is added, and a five-segment line is used to design the non-linear stiffness curve of the suspension element.

[0028] According to any embodiment of the first aspect of the object of the present invention, in step (d), the non-linear stiffness curves of each direction of the designed suspension element are respectively added to the Adams / View software, and the non-linear stiffness curves are respectively imported into the U, V, and W directions of the suspension element. After the non-linear stiffness curves of each suspension element are set, a torque about the Y axis is set at the center of mass of the powertrain to establish a simulation analysis model of the suspension system working conditions. After the establishment is completed, the displacement and rotation angle at the center of mass of the powertrain are set for measurement.

[0029] According to any embodiment of the first aspect of the object of the present invention, in step (e), after the simulation analysis model of the suspension system working conditions is established, corresponding accelerations under the working conditions are applied to the suspension system according to the general 28 working conditions. In some working conditions, a driving load needs to be applied at the center of mass of the powertrain to obtain the simulation results of the suspension element under different working conditions; after the simulation analysis of the powertrain suspension system is carried out according to the general 28 working conditions to obtain the results, the simulation results of the displacement of the center of mass of the powertrain are viewed and extracted in the Adams / PostProcessor software.

[0030] In step (e), while calculating the displacement and rotation angle of the center of mass of the powertrain, it is necessary to calculate the loads of the front suspension cushion assembly, the left suspension cushion assembly, and the right suspension cushion, which provides a reference for the subsequent design and verification of the suspension bracket.

[0031] 3. Beneficial effects

[0032] Compared with the prior art, the beneficial effects of the present invention are as follows:

[0033] Multi-objective optimization design of the design method of the present invention: Combining the modal energy decoupling rate and the NSGA-II algorithm, multi-objective optimization of the stiffness of the mounting components is carried out to ensure that the decoupling rate of each degree of freedom is ≥ 90%, the natural frequency interval is ≥ 0.8 Hz, and the human sensitive frequencies (10 - 50 Hz) are avoided; Joint simulation verification: Through the joint modeling of Matlab and Adams software, the optimization results are verified in real time to ensure that the error of the natural frequency and the decoupling rate is ≤ 3%; Nonlinear stiffness segmented design: Use three-segment or five-segment linear curves to approximate the nonlinear stiffness of the rubber mount, and combine the general 28-condition simulation to ensure that the displacements of the powertrain (±11 mm / ±5 mm / ±11 mm in the X / Y / Z directions) and the angles of rotation (±0.5° / ±3° / ±1° in the RX / RY / RZ directions) meet the requirements of the extreme conditions. Brief Description of the Drawings

[0034] The technical solutions of the present invention will be further described in detail below in conjunction with the drawings and embodiments. However, it should be understood that these drawings are only designed for the purpose of explanation and are not intended to limit the scope of the present invention. In addition, unless otherwise specified, these drawings are only intended to conceptually illustrate the structural configurations described herein and are not necessarily drawn to scale.

[0035] Figure 1 It is a simplified model of the powertrain mounting system of the present invention;

[0036] Figure 2 It is the model of the powertrain mounting system of the present invention;

[0037] Figure 3 It is the optimization flow chart of the powertrain mounting system of the present invention;

[0038] Figure 4 It is the integrated flow block diagram of the present invention;

[0039] Figure 5 It is the parameter optimization process diagram of the present invention;

[0040] Figure 6 It is the schematic diagram of the mount arrangement of the present invention;

[0041] Figure 7 It is the nonlinear stiffness curve of the rubber mount of the present invention;

[0042] Figure 8 It is the nonlinear multi-segment line of the rubber mount of the present invention;

[0043] Figure 9 It is the nonlinear stiffness curve in the u direction of the front mount cushion assembly of the present invention;

[0044] Figure 10 It is the nonlinear stiffness curve in the v direction of the front mount cushion assembly of the present invention;

[0045] Figure 11 It is the w - direction non - linear stiffness curve of the front mounting cushion assembly of the present invention;

[0046] Figure 12 It is the simulation analysis model of the working conditions of the mounting system of the present invention. Detailed implementation manners

[0047] The following detailed description of the exemplary embodiments of the present invention refers to the accompanying drawings, which form a part of the description. In the drawings, the exemplary embodiments in which the present invention can be implemented are shown by way of example. Although these exemplary embodiments are described in sufficient detail to enable those skilled in the art to implement the present invention, it should be understood that other embodiments can be achieved and various changes can be made to the present invention without departing from the spirit and scope of the present invention. The following more detailed description of the embodiments of the present invention is not intended to limit the scope of the claimed present invention, but is merely for illustrative purposes and does not limit the description of the features and characteristics of the present invention, in order to present the best mode of implementing the present invention and to enable those skilled in the art to implement the present invention. Therefore, the scope of the present invention is defined only by the appended claims.

[0048] The following detailed description and exemplary embodiments of the present invention can be better understood in conjunction with the accompanying drawings, in which the elements and features of the present invention are identified by reference numerals.

[0049] It should be emphasized that there are huge differences in energy input between electric vehicles and traditional fuel vehicles, which lead to different vibration excitations of their powertrain structures. Therefore, when a pure - electric vehicle and a traditional fuel vehicle are operating normally, the vibration excitation characteristics generated by their powertrains are different.

[0050] Traditional fuel vehicles mainly convert the internal energy of fuel into mechanical energy through combustion in the engine, and transmit the power to the drive wheels through the transmission system. To adapt to different vehicle speeds and driving conditions, the clutch and gearbox are used to help adjust the torque and speed output by the fuel vehicle, so as to reduce the engine load and improve vehicle performance and fuel efficiency, etc. In order to save space and simplify the manufacturing process, thereby improving efficiency and performance, some models of fuel vehicles integrate the main reducer and differential into the transmission. At the same time, traditional vehicle engines require a cooling system to dissipate heat to maintain the engine at an appropriate operating temperature. Thus, the powertrain of traditional fuel vehicles usually has the characteristics of numerous components, large volume, and large overall mass.

[0051] However, electric vehicles use battery packs as energy storage devices. When an electric vehicle starts, the battery pack releases the stored electrical energy to supply power to auxiliary systems such as the powertrain, heating, air conditioning, and electronic devices. Since the core component of the powertrain, the electric motor, is smaller in weight and volume than the engine under the same power, and the electric motor has a wide speed range and can provide stable power output at different speeds without frequent shifting, pure electric vehicles generally do not require a traditional transmission or reducer, but only a simpler single-speed transmission or differential to optimize the efficiency and performance of the electric motor. Moreover, at low speeds and during the starting phase, the electric motor can usually provide a larger torque output. Thus, compared with traditional fuel vehicles, the powertrain of electric vehicles has a simpler structure, is lighter in weight and smaller in size, and has characteristics such as a large starting torque.

[0052] When traditional fuel vehicles and electric vehicles are running normally, due to the differences in their powertrain structures, the causes of vibration excitation are different. When the powertrain of a traditional fuel vehicle is working normally, its vibration excitation is mainly composed of the ignition excitation generated by the engine and the vibration generated by the reciprocating motion of the crankshaft and piston. The causes of the vibration excitation of the powertrain of traditional fuel vehicles are complex and variable, related to the engine structure, the balance of the crankshaft and connecting rod, and the engine combustion process. The interaction of these factors leads to the complexity of the vibration excitation of the powertrain of traditional fuel vehicles.

[0053] When the powertrain of a pure electric vehicle is working normally, its vibration is mainly generated by the electric motor and the transmission (differential). Due to the differences in the structures and working processes of the electric motor and the engine, the vibration excitation they generate is also different. When the electric motor is working normally, vibrations are generated due to the interaction between the magnet and the coil, rotor imbalance, magnetic force vibrations caused by alternating magnetic fields, and the external structure of the electric motor. Therefore, the vibration excitation generated when the powertrain of an electric vehicle is working has characteristics such as a relatively high and stable vibration frequency and a relatively small vibration amplitude. At the same time, the vibration generated by the electric motor mainly affects the motor itself and the surrounding systems.

[0054] In summary, by comparing and analyzing the structures and excitation forces of the powertrain of pure electric vehicles and traditional fuel vehicles, the following two differences can be summarized: 1) Compared with traditional fuel vehicles, there is no traditional transmission or reducer in the powertrain of electric vehicles. Only a single transmission or differential with a simpler structure is required. At the same time, under the same power, the weight and volume of the electric motor are smaller. Therefore, the powertrain of electric vehicles has the characteristics of relatively simple structure, fewer components, light weight, and small volume; 2) Compared with traditional fuel vehicles, during normal driving of electric vehicles, there are significant differences in the vibration excitation generated by their powertrains. The vibration excitation is simpler, more stable and concentrated, with a high vibration frequency, a small vibration amplitude, and a relatively small impact on the whole vehicle.

[0055] The design method of the rubber mounting system for the powertrain of an electric vehicle according to the present invention. The rubber mounts for the powertrain of the electric vehicle include a front mount cushion assembly, a left mount cushion assembly, and a right mount cushion assembly (hereinafter referred to as mounting elements). Among them, the mounting rubber main springs of the front mount cushion assembly, the left mount cushion assembly, and the right mount cushion assembly have the same structure and material. The mounting rubber main spring includes an X-type bushing main spring, a metal outer sleeve, and a metal inner core. The method includes the steps of: jointly optimizing the powertrain mounting system using Matlab and Isight software, optimizing the static stiffness of each mounting element using the NSGA-II algorithm in Isight software, reasonably adjusting the static stiffness parameters, and performing a simulation analysis on the optimized mounting system to verify the rationality and feasibility of the optimized mounting system; performing a non-linear design on the stiffness curves of each mounting element, and using Adams / View software to calculate the six-degree-of-freedom displacements at the center of mass of the powertrain according to the general 28 working conditions, verifying the rationality of the non-linear stiffness curves of each mounting element in the U, V, and W directions, and calculating the loads of each mounting element.

[0056] Among them, (1) Multi-objective optimization design: Combining the modal energy decoupling rate and the NSGA-II algorithm, perform multi-objective optimization on the stiffness of the mounting elements to ensure that the decoupling rate of each degree of freedom is ≥90%, the natural frequency interval is ≥0.8 Hz, and avoid the human sensitive frequencies (10 - 50 Hz);

[0057] (2) Joint simulation verification: Jointly model through Matlab and Adams software to verify the optimization results in real time to ensure that the error between the natural frequency and the decoupling rate is ≤3%;

[0058] (3) Non-linear stiffness segmented design: Use a three-segment or five-segment linear curve to approximate the non-linear stiffness of the rubber mount. Combine the general 28 working condition simulations to ensure that the displacements of the powertrain (±11 mm / ±5 mm / ±11 mm in the X / Y / Z directions) and the angles of rotation (±0.5° / ±3° / ±1° in the RX / RY / RZ directions) meet the requirements of the extreme working conditions;

[0059] (4) Structural consistency optimization: Adjust the stiffness ratio of the mounting components to reduce the mold development cost and achieve structural consistency of the front, left, and right mounting components.

[0060] To facilitate the analysis of the powertrain mounting system of an electric vehicle, the powertrain is simplified into a rigid body, and each mounting component is regarded as a massless spring body with mutually perpendicular principal axis directions. The mounting system is simplified into a six-degree-of-freedom model, as Figure 1 shown.

[0061] In establishing the kinematic differential equation of the simplified model of the powertrain, the Lagrangian method is usually adopted. This method uses generalized coordinates, generalized forces, and energy to describe the motion of the powertrain mounting system. The expression of the Lagrangian equation is:

[0062]

[0063] In the formula: Qi is the generalized coordinate of the mounting system;

[0064] T is the kinetic energy of the mounting system during vibration;

[0065] U is the potential energy of the mounting system during vibration;

[0066] D is the dissipated energy of the mounting system during vibration;

[0067] F i Excitation force;

[0068] (1) The kinetic energy T of the mounting system during vibration

[0069] When the powertrain mounting system vibrates, its kinetic energy is mainly composed of the translational kinetic energy and rotational kinetic energy of the system as a whole. The generalized displacement column vector is:

[0070] Q = (x, y, z, θ x , θ y , θ z ) T #(3-2)

[0071] In the formula: x, y, z are the translational displacements relative to the centroid coordinate system of the powertrain;

[0072] θ x , θ y , θ z are the rotation angles relative to the centroid coordinate system of the powertrain;

[0073] Then, according to Equation (3-2), the kinetic energy of the powertrain mounting system during vibration can be expressed as:

[0074]

[0075] Where: I x , I y , I z is the moment of inertia of the powertrain;

[0076] I xy , I yz , I zx are the products of inertia of the powertrain;

[0077] Express the kinetic energy during the vibration of the powertrain mounting system in matrix form:

[0078]

[0079] In Equation (3-4), the mass matrix

[0080] (2) The potential energy U during the vibration of the mounting system

[0081] When the powertrain mounting system vibrates, its potential energy is mainly composed of the gravitational potential energy of the powertrain and the mounting system, as well as the potential energy when the mounting elements deform. When the powertrain is operating normally, the change in the gravitational potential energy of the powertrain and the mounting elements is relatively small. Therefore, when analyzing the powertrain mounting system, it is usually only necessary to consider the potential energy when the mounting elements deform.

[0082] Suppose there are n mounting elements in the powertrain mounting system, and the angles between the elastic principal axis directions of the mounting elements and the vehicle coordinate system are shown in Table 3-1.

[0083] Table 3-1 Angles between the elastic principal axes of the mounting elements and the vehicle coordinate system

[0084]

[0085]

[0086] For each mounting element in the powertrain mounting system, assume the coordinates of the i-th mounting element are x i , y i , z i , then the displacement vector of its elastic center point in the generalized coordinate system can be expressed as:

[0087] Δr i = Δx i i + Δy i k + Δz i k#(3-5)

[0088] Where: Δx i = x - y i θ z + z i θ y ;

[0089] Δy i = y - z i θ x + x i θ z ;

[0090] Δz i = z - x i θ y + y i θ x ;

[0091] Let i, j, k be the unit vectors of the elastic main axes u, v, w of the suspension element, then the displacement vector Δr of the elastic center point of the i-th suspension element i = Δx i i + Δy i k + Δz i k along its elastic main axes u i , v i , w i The components in 3 directions are:

[0092]

[0093] Expressed in matrix form as:

[0094] Δq i = T i E i Q#(3 - 7)

[0095] Where: Δq i = (Δu i Δv i Δw i ) T

[0096] Direction transfer matrix

[0097] Position transfer matrix

[0098] Let the elastic main axis stiffnesses of the i-th suspension element be k ui , k vi , k wi According to Equation (3 - 6), the potential energy of the powertrain suspension system is:

[0099]

[0100] The stiffness matrix K of the powertrain suspension system can be expressed as:

[0101]

[0102] Where: k idenotes the stiffness matrix of the \(i\)-th mounting element,

[0103] According to Eqs. (3-8) and (3-9), the powertrain mounting system can be expressed in matrix form as:

[0104]

[0105] (3) Dissipated energy \(D\) of the mounting system during vibration

[0106] For the non-conservative powertrain mounting system under study, there are non-damping forces (such as frictional forces, resistance, etc.), resulting in energy dissipation and damping phenomena. According to the potential energy derivation, the dissipated energy of the system can be obtained. Assuming that the damping is a function of the generalized velocity, the dissipated energy can be expressed by the following formula:

[0107]

[0108] The damping matrix \(C\) of the powertrain mounting system can be expressed as:

[0109]

[0110] where: \(c\) i denotes the damping matrix of the \(i\)-th mounting element,

[0111] According to Eq. (3-7), it can be known that:

[0112] \(\Delta q\) i \( = T\) i \(C\) i \(Q\) #(3-13)

[0113] Expressed in matrix form as:

[0114]

[0115] Based on the derivation of \(T\), \(U\), and \(D\) during the vibration of the powertrain mounting system and combined with Eq. (3-1), the vibration equation of the powertrain mounting system can be obtained as:

[0116]

[0117]

Establishment of Powertrain Mounting System Model

[0118] 1. Modal analysis of the powertrain mounting system

[0119] When analyzing the inherent characteristics of the powertrain mounting system, the damping of the system is usually not considered. Then the free vibration equation of the powertrain mounting system can be simplified as:

[0120]

[0121] Where: M is the mass matrix;

[0122] K is the stiffness matrix;

[0123] is the acceleration vector;

[0124] X is the displacement vector;

[0125] Solving the matrix equation gives:

[0126] (K - ω 2 M)X = 0 #(3 - 17)

[0127] Where: ω is the natural circular frequency;

[0128] From Equation (3 - 17), the linear equations of the vibration modes of the six degrees of freedom of the powertrain mounting system can be obtained:

[0129]

[0130] From Equation (3 - 19), the six - order natural circular frequencies ω 1 to ω 6 and the vibration modes φ i corresponding to the natural circular frequency ω i .

[0131] After obtaining the natural circular frequency of the i - th order of the mounting system, according to Equation (3 - 20), the natural frequency f i of the i - th order of the mounting system can be obtained.

[0132]

[0133] 2. Decoupling analysis of the powertrain mounting system

[0134] The degree of decoupling is a parameter that describes the degree of coupling between different vibration modes in the system. The higher the degree of decoupling, the more independent the vibration modes in the system are and the less likely they are to affect or couple with each other. Therefore, increasing the decoupling rate of the mounting system can improve the vibration isolation performance of the mounting system.

[0135] The modal energy decoupling rate is used to describe the degree of decoupling of each degree of freedom. When the modal decoupling rate reaches 100%, it indicates that the mounting system is completely decoupled in this degree of freedom. Assuming that the kinetic energy of the mounting system represents the energy of the mounting system, when the powertrain mounting system performs the main vibration of the i - th order natural circular frequency, the maximum kinetic energy of the system is:

[0136]

[0137] Where: φ i is the i - th order vibration mode;

[0138] (φi ) k is the k-th element of φ i ;

[0139] (φ i ) k is the l-th element of φ i ;

[0140] m kl is the element at the k-th row and l-th column of the M mass matrix;

[0141] From Equation (3-21), the kinetic energy occupied by the k-th generalized coordinate of the powertrain mounting system can be obtained as:

[0142]

[0143] From the above analysis, the modal energy decoupling rate of the k-th coordinate system is:

[0144]

[0145] For example, the mass of the powertrain of the present invention is 57.1 kg. The centroid coordinates of the powertrain and the elastic center point coordinates of each mounting element are shown in Table 3-2, and the inertial parameters of the powertrain are shown in Table 3-3.

[0146] Table 3-2 Coordinates of Powertrain and Mounting Elements

[0147]

[0148] Table 3-3 Inertial Parameter Table of Powertrain

[0149]

[0150] The installation angles and initial stiffnesses of each mounting element are shown in Tables 3-4 and 3-5, and the dynamic-to-static stiffness ratio of each mounting element is taken as 1.4.

[0151] Table 3-4 Installation Angles of Mounting Elements

[0152]

[0153] Table 3-5 Initial Stiffness Parameters of Mounting Elements

[0154]

[0155] According to the derived relevant formulas of modal analysis and energy decoupling rate, a calculation program is written in Matlab software to preliminarily calculate the modes and energy decoupling rate of the powertrain mounting system, providing relevant references for the subsequent optimization of the system. The relevant parameters of the powertrain and mounting elements are input into this decoupling program, and the calculation results are shown in Table 3-6.

[0156] Table 3-6 Calculation Results of the Powertrain Mounting System Based on Matlab

[0157]

[0158] Use Adams simulation software to build a model of the powertrain mounting system for simulation analysis to verify the correctness of the decoupling program of the powertrain mounting system. In Adams software, create a simplified rigid body model of the powertrain according to the relevant parameters of the powertrain, as Figure 2 shown. Set bushings at the elastic centers of each mounting element to simulate the mounting elements, and input the dynamic stiffness parameters of each mounting element in the X, Y, and Z directions.

[0159] As Figure 3 shown, when using Matlab and Isight jointly to optimize the powertrain mounting system, it is necessary to determine the optimization objective, optimization parameters, and constraint conditions, and select a suitable optimization algorithm in Isight software to optimize the mounting system. Finally, conduct relevant verification on the optimization results to determine whether the natural characteristics of the optimized mounting system meet the design requirements.

[0160] Based on the energy decoupling rate requirements of each degree of freedom of the electric vehicle powertrain mounting system, the optimization objective function of the mounting system is:[[]]

[0161]

[0162] In the formula: w i is the weight coefficient, and the value is 1;

[0163] T i is the energy decoupling rate of the i-th order mode in a certain degree of freedom;

[0164] In this embodiment, the optimization target range of the energy decoupling rate of each degree of freedom of the electric vehicle powertrain mounting system is shown in Table 3-7.

[0165] Table 3-7 Optimization Target Range of the Mounting System

[0166]

[0167] The modal frequencies of each order of the electric vehicle mounting system are in the range of 10 - 50 Hz, and the interval between the modal frequencies of different orders is greater than 0.8 Hz. The energy decoupling rate in each excitation direction of the powertrain needs to reach more than 90%. Therefore, taking the natural frequency of the mounting system and the interval between the natural frequencies of adjacent order modes as the constraint conditions, the optimization model constraint function is:[[]]

[0168]

[0169] In the formula: f iis the natural frequency of the i-th order suspension system, where i = 1 to 6.

[0170]

Optimization and Result Analysis of Suspension System Based on Matlab and Isight

[0171] Build the co-simulation framework of Isight and Matlab software as shown in Figure 4 Open the Matlab module, and establish input variables and output variables respectively. The input variables are the stiffnesses of each suspension element, and the output variables are the modal frequencies of each order of the powertrain suspension system, the interval between adjacent modal frequencies, and the energy decoupling rate in each excitation direction. Finally, complete the setting of the Matlab module by importing the written Matlab file and setting the Matlab executable file to start.

[0172] Open the Optimization module, set the range and optimization objectives of the input and output variables, select the NSGA-II optimization algorithm, and set the parameters of the optimization algorithm. The parameter settings of the optimization algorithm are shown in Table 3-8.

[0173] Table 3-8 NSGA-II Optimization Algorithm Parameters

[0174]

[0175] Run the built co-simulation framework, and the optimization process diagrams of each optimization parameter can be obtained, as shown in Figure 5 shown.

[0176] After optimization, the optimal solutions of the parameters of the suspension system optimization model output by the Isight software are shown in Table 3-9.

[0177] Table 3-9 Static Stiffness Parameters of Suspension Elements after Optimization

[0178]

[0179] When designing the static stiffness of the suspension element, not only the natural frequency of the powertrain suspension system should be considered, but also the manufacturing cost of the powertrain suspension system. To reduce the mold development and later experimental costs of the main springs of each suspension element, the structure and raw materials of the main springs of each suspension element should be kept as consistent as possible, that is, the three-way stiffness ratio is consistent. Therefore, to ensure the consistency of each suspension structure, according to the optimized parameters in Table 3-11, each suspension is arranged in the form as shown in Figure 6 shown.

[0180] To ensure the consistency of the structure of each suspension element, slightly adjust the optimized parameters. Finally, the three-way stiffnesses of each suspension element after adjustment are shown in Table 3-10.

[0181] Table 3-10 Final Static Stiffness Parameters of Suspension Elements

[0182]

[0183] Substitute the static stiffness of each optimized mounting element into the above-mentioned mounting decoupling program, conduct decoupling analysis and calculation on the powertrain mounting system, and compare the natural characteristics of the mounting system before and after optimization, as shown in Table 3-11.

[0184] Table 3-11 Comparison of Natural Characteristics of Mounting System Before and After Optimization

[0185]

[0186] As can be seen from Table 3-11, the natural frequencies of each order mode of the optimized powertrain mounting system are all between 10 - 50 Hz, meeting the design requirements. Among them, the minimum frequency interval is 0.91, and the frequency intervals of the remaining adjacent modes are all greater than 0.8 Hz, meeting the design requirements; the energy decoupling rates of X, Y, Z, RX, RY, and RZ of the optimized powertrain mounting system are 94.72%, 90.45%, 99.63%, 99.61%, 95.37%, and 91.39% respectively, all greater than the design requirement of 90%, meeting the design requirements. The energy decoupling rates of each excitation direction of the optimized powertrain mounting system are all improved compared with those before optimization. At the same time, the natural frequency range and the frequency interval of adjacent modes are relatively reasonable, and the vibration isolation performance of the powertrain mounting system is improved.

[0187]

Nonlinear Stiffness Design of Mounting System

[0188] The function of the powertrain mounting is not only to provide good vibration isolation performance, but also to prevent the powertrain from interfering with other components in the engine compartment. Therefore, it is necessary to add limits to each mounting element to make the stiffness curve of the mounting element non-linear. The spatial motion control requirements of the powertrain of electric vehicles are shown in Table 3-12.

[0189] Table 3-12 Spatial Motion Control Requirements of Powertrain

[0190]

[0191] Generally, the stiffness curves (force F vs. displacement S curves) of rubber mountings in the U, V, and W directions are continuous non-linear curves, as Figure 7 shown. Generally, in actual engineering applications, when conducting non-linear stiffness design on mounting elements, multi-segment linear segments of three segments (bc - cd - de) or five segments (ab - bc - cd - de - ef) are used to replace the non-linear stiffness curve, as Figure 8 shown. According to k 1 、k 2 、k 3The stiffness difference is used to judge the design method of the non-linear stiffness curve. If k 1 , k 2 , k 3 have a small stiffness difference, then a three-segment line is used to design the non-linear stiffness curve of the mounting element; otherwise, an additional transition segment is added and a five-segment line is used to design the non-linear stiffness curve of the mounting element.

[0192] The design requirements for the non-linear stiffness of the electric vehicle powertrain mounting system are as follows:

[0193] (1) The non-linear stiffness curves of each mounting element should be as gentle as possible;

[0194] (2) When the mounting element limits the powertrain, it should be ensured that at least two mounting elements are involved to avoid excessive load on a single mounting element;

[0195] (3) Transition segment: The maximum torque of the forward and reverse gears of the electric vehicle powertrain + the forward and backward acceleration of 0.5g, 1g for lateral and vertical directions;

[0196] (4) Hard limit segment: Under the extreme working conditions of the electric vehicle, the force and displacement of the powertrain are the largest. To prevent the powertrain from having kinematic interference with other components in the engine compartment, according to the general 28 working conditions of the vehicle, hard limits should be set for the extreme working conditions of the mounting element.

[0197] Since the structures of the main springs of the three mounting elements are the same and the front mounting cushion assembly in the electric vehicle mounting system bears the largest force, only the non-linear stiffness curve of the front mounting element needs to be designed, and it is considered that the non-linear stiffness curves of the left, right and front mounting cushion assemblies are the same. Referring to the general 28 working conditions, the mounting system is simulated using Admas software to obtain relevant simulation data. According to the design requirements of the non-linear stiffness of the electric vehicle and the requirements for the spatial motion control of the powertrain, the non-linear stiffness curves of the front mounting cushion assembly in the U, V, and W directions are respectively as Figure 9 , Figure 10 and Figure 11 shown.

[0198]

Simulation Analysis of the 28 Working Conditions of the Powertrain Mounting System

[0199] After the non-linear stiffness curves of each mounting element are designed, it is necessary to perform a simulation analysis on the powertrain mounting system according to the general 28 working conditions to check the linear stiffness curves of the designed mounting elements. Since the stiffness curves of each mounting element are non-linear, and the stiffness of the bushing model (each mounting element model) in the powertrain mounting system model established previously is linear stiffness, it is necessary to establish a three-force model at the elastic center point of each mounting element to replace the bushing model. Add the non-linear stiffness curves of each direction of the mounting elements designed in Section 3.5 to the Adams / View software respectively, and import the non-linear stiffness curves into the U, V, and W directions of the mounting elements respectively. After completing the setting of the non-linear stiffness curves of each mounting element, set the torque around the Y-axis at the centroid of the powertrain. The simulation analysis model of the mounting system working conditions is as Figure 12 shown. After the simulation analysis model of the mounting system working conditions is established, it is also necessary to set the measurement of the displacement and rotation angle at the centroid of the powertrain.

[0200] After the simulation analysis model of the mounting system working conditions is established, apply the acceleration under the corresponding working conditions to the mounting system according to the general 28 working conditions. For some working conditions, it is also necessary to apply a driving load at the centroid of the powertrain to obtain the simulation results of the mounting system under different working conditions. The general 28 working conditions are shown in Table 3-13.

[0201] Table 3-13 General 28 Working Condition Calculation Specifications

[0202]

[0203] After obtaining the results of the simulation analysis on the powertrain mounting system according to the general 28 working conditions, view and extract the simulation results in the Adams / PostProcessor software. The displacement of the powertrain centroid is shown in Table 3-14.

[0204] Table 3-14 Displacement of the Powertrain Centroid

[0205]

[0206] As can be seen from Table 3-14, the maximum displacement of the powertrain centroid in the X direction is 8.993 mm, the maximum displacement in the Y direction is 4.289 mm, and the maximum displacement in the Z direction is 3.130 mm, meeting the requirements for the spatial motion control of the powertrain centroid; the maximum rotation angles of the powertrain centroid in the RX, RY, and RZ directions are 0.215°, 2.109°, and 0.615° respectively, meeting the requirements for the spatial motion control of the powertrain centroid. Therefore, the non-linear stiffness curves of each mounting element in the U, V, and W directions meet the design requirements.

[0207] While calculating the centroid displacement and rotation angle of the powertrain, calculate the loads of the front mount soft pad assembly, the left mount soft pad assembly, and the right mount soft pad. The elastic loads of each mount component are shown in Table 3-15.

[0208] Table 3-15 Mount Loads

[0209]

[0210]

[0211]

[0212] Specifically, through multi-disciplinary joint optimization and non-linear stiffness design, the decoupling rate of the mount system of the present invention is increased by more than 30%, the resonance risk is reduced by 60%, and the manufacturing cost is reduced by 15%, significantly improving the NVH performance and driving safety of electric vehicles.

[0213] The present invention and its embodiments have been schematically described above. The description is not restrictive, and only one of the embodiments of the present invention is shown in the drawings. The actual structure is not limited thereto. Therefore, if those of ordinary skill in the art are inspired by it and design similar structural modes and embodiments without creative efforts without departing from the gist of the present invention, they shall fall within the protection scope of the present invention.

Claims

1. A design method for an electric vehicle powertrain rubber suspension system based on multi-objective optimization, characterized in that: The method comprises the following steps: optimizing the powertrain suspension system by using Matlab and Isight software, optimizing the static stiffness of each suspension component by using the NSGA-II algorithm in the Isight software, reasonably adjusting the static stiffness parameters, simulating and analyzing the optimized suspension system, and verifying the rationality and feasibility of the optimized suspension system; performing nonlinear design on the stiffness curve of each suspension component, and calculating the displacement of six degrees of freedom at the center of mass of the powertrain according to the general 28 working conditions by using Adams / View software, verifying the rationality of the nonlinear stiffness curve of each suspension component in the U, V and W directions, and calculating the load of each suspension component.

2. The design method of the electric vehicle powertrain rubber suspension system based on multi-objective optimization according to claim 1 is characterized in that: The specific steps include: a. Establishment of mathematical model of powertrain suspension system: b. Establishment of powertrain suspension system model: including modal analysis of powertrain suspension system and decoupling analysis of powertrain suspension system; c. Powertrain mounting system simulation and optimization: Use Matlab and Isight software to optimize the powertrain mounting system and improve the vibration isolation performance of the system; d. Nonlinear stiffness design of suspension system: e. Simulation analysis of 28 working conditions of powertrain suspension system.

3. The design method of the electric vehicle powertrain rubber suspension system based on multi-objective optimization according to claim 2 is characterized in that: In step (a), the powertrain is simplified into a rigid body, each suspension element is regarded as a massless spring body with main axis directions perpendicular to each other, the suspension system is simplified into a six-degree-of-freedom model, and the motion differential equation of the simplified powertrain model is established; Where: Qi is the generalized coordinate of the suspension system; T is the kinetic energy of the suspension system when it vibrates; U is the potential energy of the suspension system when it vibrates; D is the dissipated energy when the suspension system vibrates; F i Exciting force.

4. The design method of the electric vehicle powertrain rubber suspension system based on multi-objective optimization according to claim 2 is characterized in that: In step (c), a joint simulation framework of Isight and Matlab software is built, the Matlab module is opened, and input variables and output variables are established respectively, wherein the input variable is the stiffness of each suspension component, and the output variable is each order modal frequency of the powertrain suspension system, the interval between adjacent order modal frequencies, and the energy decoupling rate of each excitation direction. The setting of the Matlab module is completed by importing the written Matlab file and setting and starting the Matlab executable file; Open the Optimization module, set the range of input and output variables and the optimization target, select the NSGA-II optimization algorithm, and set the parameters of the optimization algorithm; run the constructed joint simulation framework to obtain the optimization process diagram of each optimization parameter; After optimization, the Isight software outputs the optimal solution for the parameters of the suspension system optimization model.

5. The design method of the electric vehicle powertrain rubber suspension system based on multi-objective optimization according to claim 2 is characterized in that: In step (d), when performing nonlinear stiffness design on the suspension element, a multi-segment linear line segment of three segments (bc—cd—de) or five segments (ab—bc—cd—de—ef) is used to replace the nonlinear stiffness curve, and the design method of the nonlinear stiffness curve is determined according to the stiffness difference between adjacent segments. If the stiffness difference is small, a three-segment line is used to design the nonlinear stiffness curve of the suspension element; On the contrary, a transition section is added and a five-segment line is used to design the nonlinear stiffness curve of the suspension element.

6. The design method of the electric vehicle powertrain rubber suspension system based on multi-objective optimization according to claim 5 is characterized in that: In step (d), the nonlinear stiffness curves of the suspension elements designed above in each direction are added to the Adams / View software respectively, and the nonlinear stiffness curves are imported into the U, V, and W directions of the suspension elements respectively. After the setting of the nonlinear stiffness curves of each suspension element is completed, the torque around the Y-axis is set at the center of mass of the powertrain, and a suspension system working condition simulation analysis model is established. After the establishment is completed, the displacement and rotation angle at the center of mass of the powertrain are measured.

7. The design method of the electric vehicle powertrain rubber suspension system based on multi-objective optimization according to claim 2 is characterized in that: In step (e), after the suspension system working condition simulation analysis model is established, the acceleration under the corresponding working condition is applied to the suspension system according to the general 28 working conditions. Some working conditions require applying a driving load at the center of mass of the powertrain to obtain simulation results of the suspension components under different working conditions; The powertrain center of mass displacement simulation results are then viewed and extracted in Adams / PostProcessor software.

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