Fifteen-degree-of-freedom vehicle dynamics modeling method for steer-by-wire system

Abstract

The invention discloses a 15-degree-of-freedom whole vehicle dynamics modeling method for a steer-by-wire system, and the method comprises the steps: dividing a whole vehicle model into six dynamics modules: a vehicle body model, a nonlinear tire model, a suspension model, a wheel motion model, a motor model, and a steer-by-wire model; in addition, a tire slip angle and slip rate state quantity calculation module and a steer-by-wire control module are adopted, a whole vehicle dynamic model is established and comprises six degrees of freedom of longitudinal, lateral, yawing, pitching, tilting and vertical movement of a vehicle body, two degrees of freedom of rotation and vertical movement of four distributed wheels and one degree of freedom of vehicle steering, and the total number of the degrees of freedom is 15. The method designed by the invention is suitable for common vehicle driving conditions so as to research the control performance, the stability, the safety, the dynamic response and the characteristics of a steer-by-wire system of the vehicle and vehicle performance under different working conditions, and theoretical support is provided for optimization of future vehicle design and control strategies.

Description

Technical Field

[0001] The present invention relates to the technical field of mechanical engineering, and particularly relates to a fifteen-degree-of-freedom vehicle dynamics modeling method for a steer-by-wire system. Background Art

[0002] With the rapid development of modern automotive technologies, the steer-by-wire (SBW) system, as a new type of steering control technology, has gradually become a research and development hotspot. Different from traditional mechanical steering systems, the steer-by-wire system realizes the driver's steering intention through the transmission of electrical signals, cancels the mechanical connection between the steering wheel and the wheels, and has significant advantages such as reducing steering transmission errors, decreasing weight, increasing the vehicle layout freedom, and optimizing the driving experience. In addition, the steer-by-wire system can be more conveniently combined with advanced driver assistance systems and autonomous driving technologies, laying a technical foundation for the realization of future intelligent connected vehicles.

[0003] Mathematical modeling of vehicle dynamics is one of the important methods for studying vehicle dynamics characteristics and the dynamic performance of steering systems. Currently, vehicle dynamics modeling mainly adopts multi-degree-of-freedom system modeling techniques, and common degrees of freedom include two degrees of freedom, seven degrees of freedom, fourteen degrees of freedom, etc. However, existing vehicle dynamics models generally have certain deficiencies in terms of accuracy and applicability. For example, traditional models usually ignore the non-linear characteristics of wheels, the coupling effect between the steering system and the suspension system, and the influence of vehicle body attitude changes on the vehicle's dynamic performance, making it difficult to comprehensively reflect the true operating characteristics of the vehicle under complex working conditions.

[0004] Especially in the research on steering systems, most existing modeling methods still adopt steering systems with traditional mechanical connections, and the research technology is slightly backward; or most of them only stay at the research on a single module of the steer-by-wire system, such as only focusing on the mathematical modeling of the steering actuator or the steering wheel input characteristics, and lacking a comprehensive description of the multi-body dynamics behavior of the whole vehicle. This limitation makes it difficult for existing models to provide accurate theoretical support when studying the complex motion behavior of vehicles, and cannot meet the actual needs of high-precision dynamic analysis and control optimization.

[0005] Therefore, in order to overcome the above deficiencies, the vehicle dynamics modeling method based on multi-body dynamics has become an important development direction for studying vehicle dynamic characteristics. By establishing a fifteen-degree-of-freedom vehicle dynamics model, the dynamic behavior of the vehicle under different working conditions can be more comprehensively described, especially for the dynamic characteristics and control strategy optimization problems of the steer-by-wire system. This modeling method can provide strong support for improving the steering performance of vehicles, optimizing vehicle control, and the development of autonomous driving technologies. Summary of the Invention

[0006] Objective of the present invention: To provide a fifteen-degree-of-freedom vehicle dynamics modeling method for a steer-by-wire system. By establishing dynamic constraints through the mutual coupling of vehicle module models, it is used to study the handling performance, stability, safety, dynamic response of the vehicle, the characteristics of the steer-by-wire system, and the vehicle performance under different working conditions, providing theoretical support for the optimization of future vehicle design and control strategies.

[0007] To achieve the above functions, the present invention designs a fifteen-degree-of-freedom vehicle dynamics modeling method for a steer-by-wire system. For the target vehicle, the following steps S1 - S7 are executed to establish a fifteen-degree-of-freedom vehicle dynamics model, where the fifteen degrees of freedom include six degrees of freedom for the longitudinal, lateral, yaw, pitch, roll, and vertical motions of the vehicle body of the target vehicle, two degrees of freedom for the rotation and vertical motion of four distributed wheels, and one degree of freedom for the steering of the target vehicle:

[0008] Step S1: Regard the target vehicle as a multi-body dynamics system, and establish a global coordinate system, a vehicle body coordinate system, a wheel coordinate system, and a transfer coordinate system respectively.

[0009] Step S2: According to Newton's second law and the principle of moment balance, analyze the acting forces and torques of each degree of freedom, deduce the six-degree-of-freedom dynamic expressions of the vehicle body motion and the two-degree-of-freedom dynamic expressions of each wheel motion, and establish a vehicle body model and a wheel motion model.

[0010] Step S3: Simplify the suspension and tires of the target vehicle into a parallel structure of springs and dampers, superimpose the static suspension force and the dynamic suspension force, and establish a suspension model.

[0011] Step S4: Establish a non-linear tire model using the Magic Formula.

[0012] Step S5: Simplify the structure of the steer-by-wire system of the target vehicle, and establish a steering mechanical part model and a steering execution control model respectively; among them, the steering mechanical part model is divided into two parts: a steering wheel assembly model and a steering execution assembly model. According to the structures of the components of each assembly, establish their dynamic differential equations through the principle of torque balance and the principle of electrical balance.

[0013] Step S6: Based on the relationship of the external characteristic curve of the motor torque, and at the same time simplify the torque response of the motor into a first-order inertial link, and build a motor model.

[0014] Step S7: Neglect the influence of the roll and pitch motions of the vehicle body on the wheel center speeds of each wheel, and only consider the yaw motion effect. Deduce the calculation equations of the tire sideslip angle and the tire slip ratio state variables, and then establish a tire sideslip angle and tire slip ratio calculation module. Use the tire sideslip angle and the tire slip ratio as the inputs of the fifteen-degree-of-freedom vehicle dynamics model, connect and integrate each module, and complete the establishment of the fifteen-degree-of-freedom vehicle dynamics model.

[0015] Beneficial effects: Compared with the prior art, the advantages of the present invention include:

[0016] 1. Based on multi-body dynamics and facing the steer-by-wire system, the present invention includes six degrees of freedom for the vehicle body in longitudinal, lateral, yaw, pitch, roll, and vertical motions, two degrees of freedom for the rotation and vertical motion of each of the four distributed wheels, and one degree of freedom for vehicle steering, totaling a fifteen-degree-of-freedom vehicle dynamics model. Under basic working conditions, it can well calculate the required vehicle state variables, and the accuracy is basically consistent with the output waveform of the built-in model of Carsim.

[0017] 2. The vehicle dynamics model established by the present invention can well cooperate with the steer-by-wire system model, and achieve the expected good steering effect under the action of the steer-by-wire system control model, which is convenient for researchers to continuously optimize the control method of the steer-by-wire system and study the mutual cooperation between the steer-by-wire system and other dynamics models on this basis. For example, it can be jointly simulated and controlled with the upper controller of the four-wheel torque distribution of a four-wheel distributed in-wheel motor electric vehicle. Description of the Drawings

[0018] Figure 1 is a schematic diagram of the fifteen-degree-of-freedom vehicle dynamics model provided by an embodiment of the present invention;

[0019] Figure 2 is a schematic diagram of the steer-by-wire system provided by an embodiment of the present invention;

[0020] Figure 3 is a dynamic block diagram of the fifteen-degree-of-freedom vehicle dynamics model provided by an embodiment of the present invention;

[0021] Figure 4 is a comparison diagram of the steering wheel angle input under the double lane change condition provided by an embodiment of the present invention;

[0022] Figure 5 is a comparison diagram of the yaw rate under the double lane change condition provided by an embodiment of the present invention;

[0023] Figure 6 is a comparison diagram of the sideslip angle of the center of mass under the double lane change condition provided by an embodiment of the present invention;

[0024] Figure 7 is a comparison diagram of the lateral acceleration under the double lane change condition provided by an embodiment of the present invention;

[0025] Figure 8 is a comparison diagram of the steering wheel angle input under the steering wheel angle step input condition provided by an embodiment of the present invention;

[0026] Figure 9It is a comparison diagram of the front wheel angles under the condition of a steering wheel angle step input according to an embodiment of the present invention;

[0027] Figure 10 It is a comparison diagram of the yaw rate under the condition of a steering wheel angle step input according to an embodiment of the present invention;

[0028] Figure 11 It is a comparison diagram of the sideslip angle of the center of mass under the condition of a steering wheel angle step input according to an embodiment of the present invention;

[0029] Figure 12 It is a comparison diagram of the lateral acceleration under the condition of a steering wheel angle step input according to an embodiment of the present invention. Detailed implementation manners

[0030] The present invention will be further described below with reference to the accompanying drawings. The following embodiments are only used to more clearly illustrate the technical solutions of the present invention, and cannot be used to limit the protection scope of the present invention.

[0031] A fifteen-degree-of-freedom vehicle dynamics modeling method for a steer-by-wire system provided by an embodiment of the present invention, referring to Figure 1 For the target vehicle, perform the following steps S1 - step S7 to establish a fifteen-degree-of-freedom vehicle dynamics model, where the fifteen degrees of freedom include six degrees of freedom of the body of the target vehicle in longitudinal, lateral, yaw, pitch, roll, and vertical motions, two degrees of freedom of the rotation and vertical motions of four distributed wheels, and one degree of freedom of the steering of the target vehicle:

[0032] Step S1: Regard the target vehicle as a multi-body dynamics system, which involves the matching and coordination of subsystems such as the body, suspension, and steering. During the driving process of the vehicle, the vehicle operating state is jointly affected by various factors such as the driver's operation, the states of vehicle components, and the driving road conditions. The vehicle components are coupled and interact with each other, and the force conditions are constantly changing. To simplify the analysis of each subsystem, establish a ground coordinate system, a body coordinate system, a wheel coordinate system, and a transfer coordinate system respectively;

[0033] Step S2: According to Newton's second law and the principle of moment balance, analyze the acting forces and torques of each degree of freedom, deduce the six-degree-of-freedom dynamics expressions of the body motion and the two-degree-of-freedom dynamics expressions of the motion of each wheel, and establish a body model and a wheel motion model;

[0034] Perform the following steps S2.1.1 - step S2.1.6. Combine the six-degree-of-freedom dynamics expressions of the body motion with the vehicle coordinate system to establish a differential equation system, and then establish a body model:

[0035] Step S2.1.1: Establish the differential equation of the longitudinal motion of the vehicle body as follows:

[0036]

[0037] Wherein, u is the vehicle driving speed, i.e., the longitudinal speed of the center of mass, v is the lateral speed, w is the vertical speed, r is the yaw angular velocity, q is the pitch angular velocity, p is the roll angular velocity, is the vehicle driving acceleration, i.e., the longitudinal acceleration of the center of mass; F xwij is the longitudinal wheel force, F ywij is the lateral wheel force, where the subscript i = 1, 2 represents the front wheels and rear wheels of the vehicle respectively, and the subscript j = 1, 2 represents the left wheel and right wheel of the vehicle respectively; m is the vehicle mass, α road is the road surface gradient, f is the rolling resistance coefficient, C D is the air resistance coefficient, A is the frontal area, ρ is the air density, δ ij is the steering angle of each wheel, g is the acceleration due to gravity;

[0038] Step S2.1.2: Establish the lateral motion differential equation of the vehicle body as follows:

[0039]

[0040] Wherein, is the lateral acceleration of the center of mass;

[0041] Step S2.1.3: Establish the vertical motion differential equation of the vehicle body as follows:

[0042]

[0043] Wherein, is the vertical acceleration of the center of mass, m b is the sprung mass of the vehicle; p is the roll angular velocity, F zvij is the vertical force of the wheel corresponding to the suspension;

[0044] Step S2.1.4: Establish the yaw motion differential equation of the vehicle body as follows:

[0045]

[0046] Wherein, I xx 、I yy 、I zz are the moments of inertia of the vehicle, B is the wheelbase of the vehicle, L1 is the distance from the first axis to the center of mass, and L2 is the distance from the second axis to the center of mass;

[0047] Step S2.1.5: Establish the pitch motion differential equation of the vehicle body as follows:

[0048]

[0049] Wherein, is the pitch angular acceleration of the center of mass; F zv11, F zv12 , F zv21 , F zv22 are the vertical forces of the suspension corresponding to the left front wheel, right front wheel, left rear wheel, and right rear wheel of the vehicle respectively;

[0050] Step S2.1.6: Establish the differential equation of the vehicle body roll motion as follows:

[0051]

[0052] In the formula, is the angular acceleration of roll at the center of mass.

[0053] Each wheel has two degrees of freedom of motion, namely vertical motion and rotational motion. Perform the following steps S2.2.1 - Step S2.2.2, combine the two - degree - of - freedom dynamic expression of wheel motion with the wheel coordinate system, establish a system of differential equations, and then establish a wheel motion model:

[0054] Step S2.2.1: Establish the wheel rotational motion equation as follows:

[0055]

[0056] In the formula, ω i is the rotational angular velocity of each wheel; T i is the torque provided by the i - th in - wheel motor; T bi is the braking torque; i g is the reduction ratio; I w is the moment of inertia of the wheel; r w is the wheel radius; F zwi is the vertical load of each wheel, and f is the rolling resistance coefficient;

[0057] Step S2.2.2: Establish the wheel vertical motion equation as follows:

[0058]

[0059] In the formula, m ui is the mass of each wheel, w ui is the vertical velocity of the wheel, is the vertical acceleration of the wheel, K ui , C ui are the vertical stiffness and damping of the tire of each wheel respectively; Z ui is the vertical displacement of each wheel; Z roadi is the unevenness of the road surface on which each wheel travels, and w roadi is the change rate of the wheel contact road surface unevenness.

[0060] Step S3: The longitudinal and lateral movements of the vehicle will inevitably cause the transfer of the vertical loads on each axle, and the vertical, roll, and pitch movements will also affect the elastic deformation of the suspension, thereby affecting the driving stability of the vehicle. Therefore, the suspension and tires of the target vehicle are simplified into a parallel structure of a spring and a shock absorber, and the static suspension force and the dynamic suspension force are superimposed to establish a suspension model;

[0061] Execute the following steps S3.1 - S3.3 to establish the suspension model described in step S3:

[0062] Step S3.1: Establish the static suspension force equation as follows:

[0063]

[0064] In the formula, the subscripts i = 1, 2, 3, 4 represent the left front side, right front side, left rear side, and right rear side of the vehicle respectively; F 1zvi is the static suspension force of the suspension corresponding to each wheel, and F 1zwi is the static vertical load of each wheel; m b is the sprung mass of the vehicle, and m ui is the mass of each wheel; L1 is the distance from the first axle to the center of mass, L2 is the distance from the second axle to the center of mass, and g is the acceleration due to gravity;

[0065] Step S3.2: Establish the dynamic suspension force equation as follows:

[0066]

[0067] F 2zwi = K ui (Z roadi - Z ui ) + C ui (w roadi - w ui )

[0068] In the formula, F 2zvi is the dynamic suspension force of the suspension corresponding to each wheel; F 2zwi is the dynamic vertical load of each wheel; K si , C si are the stiffness and damping of each suspension respectively; K ui , C ui are the stiffness and damping of the tires of each wheel respectively; Z s , Z ui are the vertical displacements of the vehicle body and the centers of each wheel respectively; w, w uiare the vertical velocities of the vehicle body and each wheel respectively; φ and θ are the roll angle and pitch angle of the vehicle body respectively; for the sign terms, the roll term of the left suspension is +: the roll term of the right suspension is -: the pitch term of the front axle is -; the pitch term of the rear axle is +; B is the vehicle track width, q is the pitch angular velocity, q is the pitch angular velocity, d i is the longitudinal distance from each wheel to the center of mass;

[0069] Step S3.3: Establish the total suspension force and total wheel vertical load equations as follows:

[0070] F zvi = F 1zvi + F 2zvi

[0071] F zwi = F 1zwi + F 2zwi

[0072] In the formula, F zvi is the total suspension force, and F zwi is the total wheel vertical load.

[0073] Step S4: Establish a non-linear tire model using the Magic Formula;

[0074] In Step S4, a non-linear tire model is established using the PAC2002 Magic Formula, and the unified form of the Magic Formula is:

[0075]

[0076] In the formula, Y is the lateral force, longitudinal force or aligning torque; x is the tire slip ratio or tire side slip angle; y(x) is the longitudinal force under pure longitudinal slip condition or the lateral force under pure side slip condition; X is the side slip angle or longitudinal slip ratio; D is the peak factor; b is the stiffness factor; C is the curve shape factor; E is the curve curvature factor; S h is the horizontal drift of the curve, and S v is the vertical drift of the curve.

[0077] Under the combined condition of turning and braking, there are also:

[0078]

[0079] Among them,

[0080] In the formula, F x is the tire longitudinal force; F y is the tire side force; F x0 is the tire longitudinal force under pure braking single condition; F y0 is the tire side force under pure turning single condition; α is the tire side slip angle; λ is the longitudinal slip ratio; σ xis the braking influence coefficient; σ y is the turning influence coefficient; σ is the combined influence coefficient.

[0081] Step S5: Refer to Figure 2 , simplify the steer-by-wire system structure of the target vehicle, and establish a steering mechanical part model and a steering execution control model respectively; the steering mechanical part model is divided into two parts: a steering wheel assembly model and a steering execution assembly model. According to the structure of each assembly device, establish its dynamic differential equation through the torque balance principle and the electrical balance principle;

[0082] Execute the following steps S5.1 - S5.2 to establish the steering execution assembly model in the steering mechanical part model described in step S5:

[0083] Step S5.1: Establish the steering motor torque balance equation and electrical balance equation as follows:

[0084]

[0085] T fm =K ft I fa

[0086] In the formula, T fm is the electromagnetic torque of the steering motor; θ fm is the steering motor rotation angle; is the angular velocity of the steering motor rotation angle; is the angular acceleration of the steering motor rotation angle; x r is the rack displacement; U fa 、I fa are the steering motor voltage and current; is the differential of the steering motor current with respect to time; B fm is the damping coefficient of the steering execution motor; K fc is the torsional stiffness of the pinion shaft; g fm is the reduction ratio of the steering execution motor reducer; r p is the pitch circle radius of the steering gear pinion; R fa is the resistance at the steering execution motor terminal; L fa is the inductance of the steering execution motor; K fa is the back electromotive force coefficient of the steering execution motor; K ft is the electromagnetic torque coefficient of the steering execution motor;

[0087] Step S5.2: Establish the torque balance equation of the rack and pinion structure as follows:

[0088]

[0089] In the formula, is the rack displacement speed; is the acceleration of the rack displacement; T fl is the aligning torque of the kingpin of the left front wheel; T fr is the aligning torque of the kingpin of the right front wheel; M r is the mass of the rack; B r is the damping coefficient of the rack; L fl 、L fr are the lengths of the steering arms of the left front wheel and the right front wheel.

[0090] Based on the condition that the steady-state yaw rate gain remains unchanged, the ideal transmission ratio is determined. Compared with the traditional steering system with a fixed transmission ratio, the variable transmission ratio of the by-wire system can simplify the driver's steering operation and obtain better handling stability. The steering execution control model described in step S5 is established as follows:

[0091]

[0092] i min ≤i w ≤i max

[0093] In the formula, L is the wheelbase of the vehicle body, L1 is the distance from the first axis to the center of mass, and L2 is the distance from the second axis to the center of mass; u is the vehicle speed; K v is the steady-state yaw rate gain; ξ and γ are both the influence coefficients of acceleration on the steering ratio; β is the influence coefficient of road surface friction on the steering ratio; μ road is the road surface friction coefficient; is the vehicle driving acceleration; i w is the ideal transmission ratio; i min 、i max are the minimum and maximum values of the ideal transmission ratio, and m is the total vehicle mass.

[0094] Step S6: The common motor modeling methods mainly include mathematical methods and empirical methods. Although the mathematical method has high calculation accuracy, it lacks real-time performance and is not suitable for vehicle dynamics control. Therefore, based on the relationship of the motor torque external characteristic curve and simplifying the torque response of the motor to a first-order inertia link, a motor model is built;

[0095] The motor model described in step S6 is as follows:

[0096]

[0097] In the formula, T q (t) is the actual output torque of the motor; n(t) is the actual output speed of the motor; T max is the peak torque of the motor; P e is the rated power of the motor; n N is the speed corresponding to the peak torque under the rated power of the motor, KT and K n are the steady-state gains of torque and rotational speed respectively; τ T and τ n are the time constants of torque and rotational speed responses respectively; n ref is the desired output rotational speed.

[0098] Step S7: The inputs of the fifteen-degree-of-freedom vehicle dynamics model include the tire slip ratio and the tire sideslip angle. These state variables cannot be obtained by sensors and can only be calculated. Therefore, the influence of the vehicle's roll and pitch motions on the wheel center velocities of each wheel is ignored, and only the yaw motion is considered. The calculation equations for the tire sideslip angle and the tire slip ratio state variables are derived, and then a calculation module for the tire sideslip angle and the tire slip ratio is established. The tire sideslip angle and the tire slip ratio are used as the inputs of the fifteen-degree-of-freedom vehicle dynamics model, and each module is connected and integrated to complete the establishment of the fifteen-degree-of-freedom vehicle dynamics model; The dynamics block diagram of the fifteen-degree-of-freedom vehicle dynamics model refers to Figure 3 .

[0099] The specific steps of Step S7 are as follows:

[0100] Step S7.1: Ignore the influence of the vehicle's roll and pitch motions on the wheel center velocities of each wheel, and only consider the yaw motion. Derive the calculation equations for the tire sideslip angle and the tire slip ratio state variables;

[0101] The velocity component equations of the wheel center velocities of each wheel in the horizontal direction are as follows:

[0102]

[0103] where u is the vehicle driving speed; B is the vehicle track width; δ i is the wheel angle, where the subscript i = 1, 2, 3, 4 represents the left front wheel, right front wheel, left rear wheel, and right rear wheel of the vehicle respectively; v whi is the velocity component of the wheel center velocity in the horizontal direction, d i is the longitudinal distance from each wheel to the center of mass; for the yaw term with a variable sign the yaw term of the left wheel is -, and the yaw term of the right wheel is +; for the yaw term ±d i r, the yaw term of the front axle wheels is +, and the yaw term of the rear axle wheels is -; r is the yaw angular velocity, and v is the lateral velocity;

[0104] Step S7.2: Establish the following equation for each tire sideslip angle:

[0105]

[0106] where α i is the tire sideslip angle;

[0107] Step S7.3: Establish the tire slip rate equations as follows:

[0108]

[0109] In the formula, ω i is the rotational angular velocity of each wheel, r w is the wheel radius, and λ i is the tire slip rate.

[0110] To verify the accuracy of the model, a comparison is made with the built-in mechanical steering vehicle model in Carsim, and a real vehicle simulation experiment is carried out. The input quantity of the model is the steering wheel angle.

[0111] The experimental conditions are selected as the steering wheel angle step input condition and the double lane change condition. The present invention verifies the accuracy of the vehicle model under the above conditions through the co-simulation of MATLAB / SIMULINK and Carsim. The following discusses these conditions separately.

[0112] Example 1: Double lane change condition simulation

[0113] The vehicle driving speed is maintained at 80 km / h, Figure 4 as shown in the steering wheel angle input, Figures 5 - 7 and the comparison diagrams of the yaw rate, center of mass side slip angle, and lateral acceleration under this condition are shown respectively.

[0114] Example 2: Angle step condition simulation

[0115] The vehicle driving speed is maintained at 80 km / h, Figure 8 as shown in the steering wheel angle input, Figures 9 - 12 and the comparison diagrams of the front wheel angle, yaw rate, center of mass side slip angle, and lateral acceleration under this condition are shown respectively.

[0116] In summary, the present invention proposes a fifteen-degree-of-freedom vehicle dynamics model based on multi-body dynamics, specifically for the steer-by-wire system. The model has a simple and easy-to-use modeling method, and can effectively study the handling performance, stability, safety, dynamic response of the vehicle, the characteristics of the steer-by-wire system, and its performance under different conditions. In addition, the model is also applicable to the research of control systems for other types of vehicles.

[0117] The above has described the embodiments of the present invention in detail with reference to the accompanying drawings. However, the present invention is not limited to the above embodiments, and various changes can be made without departing from the spirit of the present invention within the scope of knowledge possessed by those of ordinary skill in the art.

Claims

1. A 15-DOF vehicle dynamics modeling method for a steer-by-wire system, characterized in that: For the target vehicle, the following steps S1 to S7 are executed to establish a 15-degree-of-freedom vehicle dynamics model, wherein the 15 degrees of freedom include six degrees of freedom of the target vehicle's body in longitudinal, lateral, yaw, pitch, roll, and vertical motion, two degrees of freedom of rotation and vertical motion of four distributed wheels, and one degree of freedom of steering of the target vehicle: Step S1: The target vehicle is regarded as a multi-body dynamic system, and a geodetic coordinate system, a vehicle body coordinate system, a wheel coordinate system, and a transfer coordinate system are established respectively; Step S2: According to Newton's second law and the principle of moment balance, the forces and moments of each degree of freedom are analyzed, the six-degree-of-freedom dynamic expression of the vehicle body motion and the two-degree-of-freedom dynamic expression of each wheel motion are derived, and the vehicle body model and wheel motion model are established; Step S3: simplifying the suspension and tire of the target vehicle into a spring and shock absorber parallel structure, superimposing the static suspension force and the dynamic suspension force, and establishing a suspension model; Step S4: using the magic formula to establish a nonlinear tire model; Step S5: simplifying the structure of the wire-controlled steering system of the target vehicle, and establishing a steering mechanical part model and a steering execution control model respectively; wherein the steering mechanical part model is divided into two parts: a steering wheel assembly model and a steering execution assembly model; according to the structure of each assembly component, the dynamic differential equation thereof is established through the torque balance principle and the electrical balance principle; Step S6: Based on the relationship of the motor torque external characteristic curve, the motor torque response is simplified to a first-order inertia link, and a motor model is constructed; Step S7: Ignore the influence of the roll and pitch motion of the vehicle body on the wheel center speed of each wheel, only consider the yaw motion effect, derive the tire slip angle and tire slip rate state quantity calculation equation, and then establish the tire slip angle and tire slip rate calculation module, use the tire slip angle and tire slip rate as the input of the 15-DOF vehicle dynamics model, connect and integrate the modules, and complete the establishment of the 15-DOF vehicle dynamics model.

2. A 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: Execute the following steps S2.1.1 to S2.1.6 to establish the vehicle body model described in step S2: Step S2.1.1: Establish the longitudinal motion differential equation of the vehicle body as follows: Where u is the vehicle speed, v is the lateral velocity, w is the vertical velocity, r is the yaw angular velocity, q is the pitch angular velocity, and p is the roll angular velocity. F is the vehicle acceleration; xwij is the wheel longitudinal force, F ywij is the lateral force of the wheel, where the subscript i=1 and 2 represent the front and rear wheels of the car respectively, and the subscript j=1 and 2 represent the left and right wheels of the car respectively; m is the mass of the vehicle, α road is the road slope, f is the rolling resistance coefficient, C D is the air resistance coefficient, A is the windward area, ρ is the air density, δ ij is the steering angle of each wheel, g is the acceleration due to gravity; Step S2.1.2: Establish the differential equation of lateral motion of the vehicle body as follows: In the formula, is the lateral acceleration of the center of mass; Step S2.1.3: Establish the vehicle body vertical motion differential equation as follows: In the formula, is the vertical acceleration of the center of mass, m b is the sprung mass of the vehicle; p is the roll angular velocity, F zvij is the vertical force of the suspension corresponding to the wheel; Step S2.1.4: Establish the differential equation of vehicle body yaw motion as follows: In the formula, I xx ,I yy ,I zz is the vehicle moment of inertia, B is the vehicle wheelbase, L1 is the distance from the first axis to the center of mass, and L2 is the distance from the second axis to the center of mass; Step S2.1.5: Establish the vehicle body pitch motion differential equation as follows: In the formula, is the pitch angular acceleration of the center of mass; F zv11 , F zv12 , F zv21 , F zv22 They are the suspension vertical forces corresponding to the left front wheel, right front wheel, left rear wheel and right rear wheel of the car respectively; Step S2.1.6: Establish the vehicle body roll motion differential equation as follows: In the formula, is the roll angular acceleration of the center of mass.

3. The 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: Execute the following steps S2.2.1 to S2.2.2 to establish the wheel motion model described in step S2: Step S2.2.1: Establish the wheel rotation motion equation as follows: In the formula, ω i is the angular velocity of each wheel; T i is the torque provided by the i-th wheel hub motor; T bi is the braking torque; i g is the reduction ratio; I w is the wheel moment of inertia; r w is the wheel radius; F zwi is the vertical load of each wheel, and f is the rolling resistance coefficient; Step S2.2.2: Establish the wheel vertical motion equation as follows: In the formula, m ui is the mass of each wheel, w ui is the vertical velocity of the wheel, is the vertical acceleration of the wheel, K ui , C ui are the vertical stiffness and damping of each wheel tire; Z ui is the vertical displacement of each wheel; Z roadi is the roughness of the road surface on which each wheel travels, w roadi is the rate of change of the roughness of the road surface the wheel contacts.

4. The 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: Perform the following steps S3.1 to S3.3 to establish the suspension model described in step S3: Step S3.1: Establish the static suspension force equation as follows: Wherein, the subscripts i=1, 2, 3, and 4 represent the left front side, right front side, left rear side, and right rear side of the vehicle respectively; F 1zvi is the static suspension force of each wheel corresponding to the suspension, F 1zwi is the static vertical load of each wheel; m b is the sprung mass of the vehicle, m ui is the mass of each wheel; L1 is the distance from the first axis to the center of mass, L2 is the distance from the second axis to the center of mass, and g is the acceleration due to gravity; Step S3.2: Establish the dynamic suspension force equation as follows: F 2zwi =K ui (Z roadi -Z ui )+C ui (w roadi -w ui ) In the formula, F 2zvi is the dynamic suspension force of the suspension corresponding to each wheel; F 2zwi is the dynamic vertical load of each wheel; K si , C si are the stiffness and damping of each suspension respectively; K ui , C ui are the stiffness and damping of each wheel tire respectively; Z s , Z ui are the vertical displacements of the vehicle body and each wheel center respectively; w, w ui are the vertical velocities of the vehicle body and each wheel respectively; φ and θ are the roll angle and pitch angle of the vehicle body respectively; for the sign terms, the roll term of the left suspension is +: the roll term of the right suspension is -: the pitch term of the front axle is -; the pitch term of the rear axle is +; B is the vehicle wheelbase, q is the pitch angular velocity, q is the pitch angular velocity, d i is the longitudinal distance from each wheel to the center of mass; Step S3.3: Establish the total suspension force and total vertical wheel load equation as follows: F zνi =F 1zνi +F 2zνi F zwi =F 1zwi +F 2zwi In the formula, F zνi is the total suspension force, F zwi is the total vertical load on the wheel.

5. The 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: In step S4, the PAC2002 magic formula is used to establish a nonlinear tire model, wherein the unified form of the magic formula is: Where Y is the lateral force, longitudinal force or self-aligning moment; x is the tire slip rate or tire slip angle; y(x) is the longitudinal force under pure longitudinal slip conditions or the lateral force under pure slip conditions; X is the slip angle or longitudinal slip rate; D is the peak factor; b is the stiffness factor; C is the curve shape factor; E is the curve curvature factor; S h is the horizontal drift of the curve, S v The curve drifts vertically.

6. The 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: Execute the following steps S5.1 to S5.2 to establish the steering execution assembly model in the steering mechanical part model described in step S5: Step S5.1: Establish the steering motor torque balance equation and electrical balance equation as follows: T fm =K ft I fa Where, T fm is the electromagnetic torque of the steering motor; θ fm is the steering motor angle; is the angular velocity of the steering motor; is the angular acceleration of the steering motor; x r is the rack displacement; U fa ,I fa The voltage and current of the steering motor; is the differential of the steering motor current with respect to time; B fm The damping coefficient of the steering actuator motor; K fc is the torsional stiffness of the pinion shaft; g fm The reduction ratio of the motor reducer for steering execution; r p R is the radius of the steering gear pinion pitch circle; fa The motor terminal resistance for steering execution; L fa The inductance of the steering actuator motor; K fa K is the back electromotive force coefficient of the steering actuator motor; ft Electromagnetic torque coefficient of the motor for steering execution; Step S5.2: Establish the gear rack structure torque balance equation as follows: In the formula, is the rack displacement speed; is the rack displacement acceleration; T fl is the left front wheel kingpin aligning torque; T fr is the right front wheel kingpin aligning moment; M r is the rack mass; B r is the rack damping coefficient; L fl , L fr It is the length of the left front wheel and right front wheel steering rocker arm.

7. The 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: The steering execution control model described in step S5 is as follows: i min ≤i w ≤i max Where, L is the wheelbase of the vehicle, L1 is the distance from the first axle to the center of mass, L2 is the distance from the second axle to the center of mass; u is the vehicle speed; K ω is the steady-state yaw rate gain; ξ and γ are the influence coefficients of acceleration on steering ratio; β is the influence coefficient of road friction on steering ratio; μ road is the road friction coefficient; is the vehicle acceleration; i w is the ideal transmission ratio; i min 、i max are the minimum and maximum values ​​of the ideal transmission ratio, and m is the vehicle mass.

8. The 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: The motor model described in step S6 is as follows: Where, T q (t) is the actual output torque of the motor; n(t) is the actual output speed of the motor; T max is the peak torque of the motor; P e is the rated power of the motor; n N K is the speed corresponding to the peak torque at rated power of the motor. T , K n are the steady-state gains of torque and speed respectively; τ T , τ n are the time constants of torque and speed response respectively; n ref is the expected output speed.

9. The 15-DOF vehicle dynamics modeling method for a steer-by-wire system according to claim 1, characterized in that: The specific steps of step S7 are as follows: Step S7.1: Ignore the influence of the vehicle body roll and pitch motion on the wheel center speed of each wheel, only consider the yaw motion effect, and derive the tire slip angle and tire slip rate state quantity calculation equation; The velocity component equation of each wheel center speed in the horizontal direction is as follows: Where, u is the vehicle speed; B is the vehicle wheelbase; δ i is the wheel turning angle, where the subscripts i=1, 2, 3, and 4 represent the left front wheel, right front wheel, left rear wheel, and right rear wheel of the vehicle, respectively; v whi is the velocity component of the wheel center speed in the horizontal direction, d i is the longitudinal distance from each wheel to the center of mass; for The left wheel yaw term is -, the right wheel yaw term is +; for yaw terms with variable signs ±d i r, the yaw term of the front axle wheel is +, and the yaw term of the rear axle wheel is -; r is the yaw angular velocity, and v is the lateral velocity; Step S7.2: Establish the side slip angle equation of each tire as follows: In the formula, α i is the tire slip angle; Step S7.3: Establish the slip rate equation of each tire as follows: In the formula, ω i is the angular velocity of each wheel, r w is the wheel radius, λ i is the tire slip rate.

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