Chassis coordination control method and system based on tire lateral deviation state evaluation

Abstract

The invention discloses a chassis coordination control method and system based on tire side deviation state evaluation, and relates to the technical field of four-wheel steering four-wheel drive vehicle dynamics control. The method comprises the steps that an expected front wheel rotation angle is determined based on an expected steering wheel rotation angle input by a driver, and a front wheel rotation angle is determined; further determining a feed-forward front wheel steering angle, a feed-forward rear wheel steering angle and an additional yaw moment based on feed-forward control and feedback control, determining a first tire slip angle and a second tire slip angle based on a road adhesion coefficient, and further determining a wheel torque control weight, a feedback front wheel steering angle control weight and a feedback rear wheel steering angle control weight; and further determining a driving yaw moment, a feedback front wheel steering angle and a feedback rear wheel steering angle, determining a total longitudinal force based on an expected longitudinal acceleration input by a driver, and performing four-wheel torque distribution based on the total longitudinal force and the driving yaw moment to obtain a four-wheel driving torque. The driving stability of the vehicle under the limiting working condition can be improved.

Description

Technical Field

[0001] The present application relates to the technical field of four-wheel steering and four-wheel drive vehicle dynamics control, and in particular to a chassis coordination control method and system based on tire side slip state assessment. Background Art

[0002] With the continuous development of intelligent and electrified new energy vehicles, chassis-by-wire control has become an inevitable trend in future vehicles. Electric vehicles equipped with four-wheel steering and four-wheel drive can effectively improve vehicle safety, stability, and maneuverability through coordinated control of four-wheel torque and angle, attracting widespread attention from academia and industry. Achieving safe vehicle operation requires a thorough understanding of the relationship between driver input and vehicle motion, while utilizing the chassis actuation system to control vehicle motion in accordance with the driver's desired behavior. Therefore, desired motion state analysis and coordinated chassis control are key to ensuring safe and efficient vehicle operation.

[0003] At present, when performing chassis coordinated control, driving stability under extreme working conditions is still poor. Summary of the Invention

[0004] The purpose of this application is to provide a chassis coordination control method and system based on tire side slip state evaluation, which can improve the driving stability of the vehicle under extreme working conditions.

[0005] To achieve the above objectives, this application provides the following solutions:

[0006] In a first aspect, the present application provides a chassis coordinated control method based on tire cornering state evaluation, the chassis coordinated control method based on tire cornering state evaluation comprising:

[0007] determining a desired front wheel angle based on a desired steering wheel angle input by the driver;

[0008] Taking the desired front wheel steering angle as input, determining a feedforward front wheel steering angle, a feedforward rear wheel steering angle, and an additional yaw moment based on feedforward control and feedback control;

[0009] determining a first tire slip angle and a second tire slip angle based on a road adhesion coefficient, and determining a wheel torque control weight, a feedback front wheel steering angle control weight, and a feedback rear wheel steering angle control weight based on the first tire slip angle and the second tire slip angle; the first tire slip angle being a critical tire slip angle between a linear region and a nonlinear region in a tire slip curve, and the second tire slip angle being a critical tire slip angle between a nonlinear region and a saturated region in the tire slip curve;

[0010] determining a driving yaw moment, a feedback front wheel angle, and a feedback rear wheel angle based on the additional yaw moment, the wheel torque control weight, the feedback front wheel angle control weight, and the feedback rear wheel angle control weight;

[0011] Based on the expected longitudinal acceleration input by the driver, a total longitudinal force is determined, and four-wheel torque distribution is performed based on the total longitudinal force and the driving yaw moment to obtain a four-wheel driving torque.

[0012] In a second aspect, the present application provides a computer system comprising: a memory, a processor, and a computer program stored in the memory and executable on the processor, wherein the processor executes the computer program to implement the above-mentioned chassis coordination control method based on tire side deviation state evaluation.

[0013] According to the specific embodiments provided in this application, this application has the following technical effects:

[0014] The present application provides a chassis coordination control method and system based on tire slip state evaluation. In order to solve the problem that driving stability is still poor under extreme working conditions, tire slip state evaluation is introduced. Specifically, after determining the feedforward front wheel angle, feedforward rear wheel angle and additional yaw moment based on feedforward control and feedback control, the first tire slip angle and the second tire slip angle are determined based on the road adhesion coefficient to perform tire slip state evaluation. The first tire slip angle is the critical tire slip angle between the linear region and the nonlinear region in the tire slip curve, and the second tire slip angle is the critical tire slip angle between the nonlinear region and the saturation region in the tire slip curve. The wheel torque control weight, the feedback front wheel angle control weight and the feedback angle are further determined based on the first tire slip angle and the second tire slip angle. The rear wheel angle control weight is fed back, so that the weight distribution process can be completed based on the tire side slip state. The driving yaw torque, feedback front wheel angle and feedback rear wheel angle are further determined based on the additional yaw moment, wheel torque control weight, feedback front wheel angle control weight and feedback rear wheel angle control weight. Based on the expected longitudinal acceleration input by the driver, the total longitudinal force is determined, and the four-wheel torque is distributed based on the total longitudinal force and the driving yaw moment to obtain the four-wheel drive torque. Finally, the feedforward front wheel angle, feedforward rear wheel angle, feedback front wheel angle, feedback rear wheel angle and four-wheel drive torque are output, thereby completing the four-wheel angle and torque distribution. Since the tire side slip state evaluation is introduced in the four-wheel angle and torque distribution process, the driving stability of the vehicle under extreme working conditions can be improved. BRIEF DESCRIPTION OF THE DRAWINGS

[0015] In order to more clearly illustrate the embodiments of the present application or the technical solutions in the prior art, the following briefly introduces the drawings required for use in the embodiments. Obviously, the drawings described below are only some embodiments of the present application. For ordinary technicians in this field, other drawings can be obtained based on these drawings without creative work.

[0016] Figure 1 This is an application environment diagram of a chassis coordination control method based on tire side slip state assessment provided in Example 1 of the present application.

[0017] Figure 2 A flowchart of a chassis coordination control method based on tire side slip state assessment provided in Example 1 of the present application.

[0018] Figure 3 Schematic diagram of the chassis coordination control architecture of a four-wheel steering four-wheel drive vehicle provided in Example 1 of the present application.

[0019] Figure 4 Schematic diagram of the two-degree-of-freedom vehicle model provided in Example 1 of the present application.

[0020] Figure 5 A schematic diagram of the vehicle steering motion state provided in Example 1 of the present application.

[0021] Figure 6 Schematic diagram of the tire side slip curve as the road adhesion coefficient changes, provided in Example 1 of the present application.

[0022] Figure 7 Schematic diagram of the coordinated control process of four-wheel angle and four-wheel torque provided in Example 1 of the present application. DETAILED DESCRIPTION

[0023] The following will be combined with the drawings in the embodiments of this application to clearly and completely describe the technical solutions in the embodiments of this application. Obviously, the embodiments described are only part of the embodiments of this application, not all of the embodiments. Based on the embodiments in this application, all other embodiments obtained by ordinary technicians in this field without making creative efforts are within the scope of protection of this application.

[0024] Example 1

[0025] The chassis coordination control method based on tire side slip state evaluation provided in the embodiment of the present application can be applied to Figure 1In the application environment shown, the terminal communicates with the server via a network. The data storage system can store the data that the server needs to process. The data storage system can be set up separately, integrated with the server, or placed in the cloud or on other servers. The terminal can send the pending control request (including the expected steering wheel angle and expected longitudinal acceleration input by the driver) to the server. After the server receives the pending control request, for the pending control request, the server determines the expected front wheel angle based on the expected steering wheel angle input by the driver; uses the expected front wheel angle as input, and determines the feedforward front wheel angle, feedforward rear wheel angle and additional yaw moment based on feedforward control and feedback control; determines the first tire slip angle and the second tire slip angle based on the road adhesion coefficient, and determines the wheel torque control weight, the feedback front wheel angle control weight and the feedback rear wheel angle control weight based on the first tire slip angle and the second tire slip angle; determines the driving yaw torque, the feedback front wheel angle and the feedback rear wheel angle based on the additional yaw moment, the wheel torque control weight, the feedback front wheel angle control weight and the feedback rear wheel angle control weight; determines the total longitudinal force based on the expected longitudinal acceleration input by the driver, and distributes the four-wheel torque based on the total longitudinal force and the driving yaw torque to obtain the four-wheel drive torque. The server may feed back the obtained control result (including feedforward front wheel angle, feedforward rear wheel angle, feedback front wheel angle, feedback rear wheel angle and four-wheel drive torque) for the control request to the terminal.

[0026] In addition, in some embodiments, the chassis coordination control method based on tire side deviation state evaluation can also be implemented independently by a server or a terminal. For example, the terminal can directly process the control request to be processed, or the server can obtain the control request to be processed from the data storage system and process the control request to be processed.

[0027] In an exemplary embodiment, Figure 2 As shown, a chassis coordination control method based on tire side slip state evaluation is provided. The method is executed by a computer device, specifically a computer device such as a terminal or a server, or a terminal and a server. In the embodiment of the present application, the method is applied to Figure 1 The following steps are used as an example to illustrate the server.

[0028] Step S1: determining a desired front wheel angle based on a desired steering wheel angle input by a driver.

[0029] Step S2 : Taking the desired front wheel steering angle as input, determining the feedforward front wheel steering angle, the feedforward rear wheel steering angle and the additional yaw moment based on feedforward control and feedback control.

[0030] Step S3, determining a first tire slip angle and a second tire slip angle based on the road adhesion coefficient, and determining a wheel torque control weight, a feedback front wheel steering angle control weight, and a feedback rear wheel steering angle control weight based on the first tire slip angle and the second tire slip angle; the first tire slip angle is a critical tire slip angle between a linear region and a nonlinear region in a tire slip curve, and the second tire slip angle is a critical tire slip angle between a nonlinear region and a saturated region in the tire slip curve.

[0031] Step S4: determining a driving yaw torque, a feedback front wheel angle, and a feedback rear wheel angle based on the additional yaw torque, the wheel torque control weight, the feedback front wheel angle control weight, and the feedback rear wheel angle control weight.

[0032] Step S5: determining the total longitudinal force based on the desired longitudinal acceleration input by the driver, and performing four-wheel torque distribution based on the total longitudinal force and the driving yaw moment to obtain a four-wheel drive torque.

[0033] By implementing the above-mentioned steps S1 to S5, this embodiment determines the first tire slip angle and the second tire slip angle based on the road adhesion coefficient when performing chassis coordinated control, and determines the wheel torque control weight, the feedback front wheel angle control weight, and the feedback rear wheel angle control weight based on the first tire slip angle and the second tire slip angle. It further determines the driving yaw moment, the feedback front wheel angle, and the feedback rear wheel angle, and then distributes the four-wheel torque to obtain the four-wheel drive torque. Therefore, by introducing the first tire slip angle and the second tire slip angle, the tire slip state evaluation is introduced when calculating the weight, thereby achieving efficient coordination of the four-wheel steering and four-wheel drive vehicles under different driving conditions, thereby improving the driving stability of the vehicle under extreme conditions.

[0034] This embodiment proposes a chassis coordination control architecture for a four-wheel steering four-wheel drive vehicle based on tire side slip state assessment. Specifically, the desired steering wheel angle and desired longitudinal acceleration input by the driver are used as inputs, and the front and rear wheel angles (including feedforward front wheel angle, feedforward rear wheel angle, feedback front wheel angle, and feedback rear wheel angle) and four-wheel drive torque are used as outputs. A global chassis coordination control architecture based on tire side slip state assessment is developed, such as Figure 3As shown in the figure, first, considering the differences in control requirements of the vehicle under different working conditions, a hybrid feedforward control method integrating steady-state control, dynamic compensation, and oblique compensation is developed; secondly, a kinematic model is used to establish an analytical method for the desired motion state covering multiple steering modes, and the desired motion state is output; then, state error feedback control is used to track the desired motion state and obtain the required additional yaw moment; on this basis, combined with the calculation results of the tire lateral force, a coordinated control strategy of the four-wheel steering angle and the four-wheel torque is constructed; finally, a four-wheel torque distribution method considering the tire slip characteristics is proposed to achieve the optimal distribution of the four-wheel drive torque, which specifically includes the following steps.

[0035] Step 1: Design a variable transmission ratio according to the desired steering wheel angle δ input by the driver. w and variable transmission ratio i w , get the desired front wheel turning angle δ f_exp =δ w / i w .

[0036] For a four-wheel steering four-wheel drive vehicle, the variable transmission ratio used in this embodiment is:

[0037]

[0038] In the above formula, i w is the variable transmission ratio; v x is the longitudinal velocity of the vehicle; G ωr is the yaw rate gain, which can be set to 0.24s -1 ; m is the vehicle mass; b is the distance from the vehicle center of mass to the rear axle; a is the distance from the vehicle center of mass to the front axle; k f is the front axle cornering stiffness; χ is the intermediate parameter; i fixed The fixed transmission ratio is 21.5 according to the data of CarSim (a vehicle dynamics simulation software). max is the maximum transmission ratio in the transition zone, which can be set to 21.5; i min is the minimum transmission ratio of the transition zone, which can be set to 18.55; v2 is the maximum vehicle speed corresponding to the endpoint of the transition zone, which can be set to 90km / h; v1 is the minimum vehicle speed corresponding to the endpoint of the transition zone, which can be set to 40km / h. The vehicle speed corresponding to the filter zone is [40km / h, 90km / h].

[0039] This embodiment uses the variable transmission ratio and the desired steering wheel angle as inputs and the desired front wheel angle as output. The calculation formula for the desired front wheel angle is:

[0040] δ f_exp =δ w / i w (3)

[0041] In formula (3), δ f_exp is the desired front wheel turning angle; δ w The desired steering wheel angle input by the driver; i w For variable transmission ratio.

[0042] At this time, in step S1, the expected front wheel angle is determined based on the expected steering wheel angle input by the driver. Specifically, the expected front wheel angle is calculated using formula (3).

[0043] Step 2: Based on the desired front wheel steering angle and desired front wheel steering angular velocity obtained in step 1, two-degree-of-freedom vehicle dynamics and vehicle kinematics modeling is performed, and steady-state control is performed based on the established vehicle dynamics model to obtain steady-state control parameters (including the steady-state control front wheel steering angle δ f_s and steady-state control rear wheel steering angle δ r_s ), dynamic compensation is performed based on the established vehicle dynamics model and vehicle kinematics model to obtain dynamic compensation control parameters (including dynamic compensation front wheel angle δ f_dc and the dynamic compensation rear wheel steering angle δ r_dc ), perform oblique compensation according to the established vehicle kinematic model to obtain oblique compensation control parameters (including oblique compensation front wheel angle δ f_oc and the rear wheel steering angle δ r_oc ), and perform hybrid feedforward control according to the steady-state control parameters, dynamic compensation control parameters and oblique compensation control parameters to obtain the final feedforward wheel angle (including the feedforward front wheel angle δ ff and the feedforward rear wheel steering angle δ fr ).

[0044] In order to give full play to the role of each feedforward control module, this section combines steady-state control, dynamic compensation and oblique compensation, and develops a hybrid feedforward control method that meets different control requirements. In the feedforward control, the desired front wheel angle and the desired front wheel angular velocity are used as input, and the feedforward front wheel angle and feedforward rear wheel angle are used as output.

[0045] (1) Vehicle dynamics model

[0046] This part is about theoretical model construction, which can provide a model basis for steady-state control, dynamic compensation and state error feedback control. In this embodiment, a two-degree-of-freedom vehicle model is used to describe the lateral and yaw motions of the vehicle, such as Figure 4 As shown, Figure 4 In, V y is the lateral velocity of the vehicle, and the specific formula is:

[0047]

[0048] In formula (4), is the sideslip angular velocity of the center of mass; k f is the front axle cornering stiffness; k r is the rear axle lateral stiffness; m is the vehicle mass; v x is the longitudinal velocity of the vehicle; β is the sideslip angle of the center of mass; a is the distance from the center of mass of the vehicle to the front axle; b is the distance from the center of mass of the vehicle to the rear axle; ω r is the yaw rate of the vehicle; δ f is the actual front wheel turning angle; δ r is the actual rear wheel turning angle; is the vehicle's yaw angular acceleration; I z is the yaw moment of inertia of the vehicle.

[0049] Then transform Equation (4) into the state space form to meet the controller design requirements. The vehicle dynamics model is:

[0050]

[0051] In formula (5), the derivative of the first matrix is Second matrix First matrix The third matrix The Fourth Matrix

[0052] (2) Vehicle kinematic model

[0053] This section builds a theoretical model, providing a model foundation for dynamic compensation, oblique driving compensation, and desired motion state analysis. Traditional driver desired motion state analysis primarily relies on vehicle dynamics models to obtain the driver's driving intention (i.e., desired motion state), but rarely analyzes the driver's driving intention from the driver's perspective, resulting in a deviation between the driver's driving intention and the vehicle's ideal motion state. To more accurately analyze the driver's driving intention, this embodiment uses a kinematic model to analyze the mapping relationship between the front and rear wheel angles and the vehicle's turning radius. Based on this, the driver's desired yaw rate is analyzed.

[0054] like Figure 5 As shown, for triangles OCA and OCB, the sine theorem gives:

[0055]

[0056] In the above formula, R is the turning radius of the vehicle.

[0057] Simplifying equations (6) and (7), we can obtain:

[0058] (tanδ f -tanδ r )cosβ=(a+b) / R(8)

[0059] The turning radius of the vehicle is:

[0060]

[0061] Then, the driver's desired yaw rate is:

[0062]

[0063] In formula (10), ω rd is the desired yaw rate.

[0064] Formula (10) is the vehicle kinematic model of this embodiment. It can be seen from Formula (10) that obtaining the driver's desired yaw rate requires determining the front wheel angle, rear wheel angle, center of mass sideslip angle, and longitudinal speed of the vehicle.

[0065] (3) Steady-state control

[0066] This part takes the desired front wheel angle as input and outputs the steady-state controlled front and rear wheel angles (including the steady-state controlled front wheel angle and the steady-state controlled rear wheel angle).

[0067] According to the vehicle dynamics model, β=0, Then the steady-state dynamic equation of a two-degree-of-freedom vehicle is:

[0068] 0=A0x+B0u s (11)

[0069] In formula (11), the fifth matrix u s =[δ f_s δ r_s ] T , δ f_s To control the front wheel angle in a steady state, δ r_s To control the rear wheel angle in a steady state.

[0070] The front and rear wheel steering angles under steady-state control can be calculated by formula (11), where the calculation formula for the front wheel steering angle under steady-state control is:

[0071] δ f_s =δ f_exp (12)

[0072] In formula (12), δ f_s To control the front wheel angle in a steady state; δ f_exp is the desired front wheel angle.

[0073] The calculation formula for the steady-state control rear wheel angle is:

[0074]

[0075] In formula (13), δ r_sis the rear wheel steering angle for steady-state control; a is the distance from the vehicle's center of mass to the front axle; b is the distance from the vehicle's center of mass to the rear axle; k f is the front axle cornering stiffness; k r is the rear axle lateral stiffness; m is the vehicle mass; v x is the longitudinal velocity of the vehicle.

[0076] (4) Dynamic compensation

[0077] This part takes the desired front wheel angle and the desired front wheel angular velocity as input, and outputs the dynamically compensated front and rear wheel angles (including the dynamically compensated front wheel angle and the dynamically compensated rear wheel angle).

[0078] Combining the vehicle kinematic model and the vehicle dynamics model, and considering the dynamic characteristics of the driver's control input, the expected front wheel angular velocity is used as input to obtain the expected yaw angular acceleration:

[0079]

[0080] In formula (14), is the desired yaw angular acceleration; is the desired front wheel angular velocity; is the desired rear wheel angular velocity, δ r_exp is the desired rear wheel turning angle,

[0081] The sixth matrix u when dynamic compensation control is introduced dc =[δ f_dc ,δ r_dc ] T , δ f_dc To dynamically compensate the front wheel angle, δ r_dc To dynamically compensate the rear wheel angle, it satisfies:

[0082]

[0083] Substituting equation (14) into equation (15), the dynamic compensation front and rear wheel steering angles of the vehicle are calculated according to the expected yaw angular acceleration. The calculation formula for the dynamic compensation front wheel steering angle is:

[0084]

[0085] In formula (16), δ f_dc To dynamically compensate for the front wheel angle; I z is the yaw moment of inertia of the vehicle; is the desired yaw angular acceleration.

[0086] The calculation formula for the dynamic compensation rear wheel angle is:

[0087]

[0088] In formula (17), δ r_dc To dynamically compensate for rear wheel angle.

[0089] (5) Slant compensation

[0090] This part takes the desired front wheel steering angle and road adhesion coefficient as input, and outputs the front and rear wheel steering angles for slant compensation (including the front wheel steering angle for slant compensation and the rear wheel steering angle for slant compensation).

[0091] Considering the limitation of road adhesion coefficient μ, the maximum yaw rate of the vehicle is:

[0092]

[0093] In formula (18), ω r_max is the maximum yaw angular velocity; g is the acceleration due to gravity.

[0094] At this time, the maximum front wheel turning angle allowed by the road surface is:

[0095]

[0096] In formula (19), δ f_max is the maximum front wheel turning angle.

[0097] When the desired front wheel steering angle exceeds the maximum front wheel steering angle allowed by the road adhesion coefficient, the slant compensation control will be used to obtain the slant compensation front and rear wheel steering angles. The calculation formulas for the slant compensation front and rear wheel steering angles are as follows:

[0098] δ f_oc =δ r_oc =δ f_exp -δ f_max (20)

[0099] In formula (20), δ f_oc Compensate the front wheel angle for oblique driving; δ r_oc Compensates rear wheel angle for diagonal driving.

[0100] (6) Hybrid feedforward control

[0101] This part takes the steady-state control front and rear wheel angles, the dynamic compensation front and rear wheel angles, the oblique compensation front and rear wheel angles, the desired yaw rate, and the road adhesion coefficient as inputs. By calculating the control weights of dynamic compensation and oblique compensation, it outputs the feedforward wheel angles (including the feedforward front wheel angle and the feedforward rear wheel angle) to the feedback control and vehicle steering system.

[0102] In feedforward control, steady-state control primarily responds to driver inputs and remains active. Dynamic compensation control compensates for dynamic changes in the vehicle's desired motion and primarily operates within the vehicle's stable region. Slope compensation control aims to improve vehicle stability and primarily activates when the vehicle's desired motion exceeds the road adhesion limit. To maximize the effectiveness of each feedforward control module, this embodiment designs a hybrid feedforward control method that meets diverse control requirements. Its key principle is to design weighting coefficients for dynamic compensation and sloping compensation based on the vehicle's desired yaw rate. The feedforward wheel angle is then output based on the hybrid weighting coefficients.

[0103] When the driver's input steering angle exceeds the maximum steering angle allowed by road adhesion, the vehicle's yaw motion approaches the stability boundary. At this point, the vehicle's control requirement changes from dynamic response control to stability control. To meet the vehicle stability control requirement, dynamic compensation switches to oblique compensation. To achieve smooth switching of compensation modes, a smooth switching function is designed as follows:

[0104]

[0105] In formula (21), η1 is the first weight coefficient; ω rd is the desired yaw rate; ω r_tra is the yaw rate control boundary value; ρ1 is the correction coefficient of control weight distribution.

[0106] In the slant compensation control mode, the vehicle's yaw rate will drop rapidly, reducing the vehicle's steering performance and requiring the driver to apply a larger steering wheel angle, increasing the driver's driving burden. Therefore, to ensure good steering performance, the slant compensation control mode cannot be engaged for too long. This embodiment uses the vehicle's yaw rate information to limit the duration of the slant compensation control. Specifically, the slant compensation control mode is exited when the vehicle's actual yaw rate falls below 90% of its peak value. To avoid sudden changes in wheel angle during the exit process, the inverse tangent function is used to smooth the exit process:

[0107]

[0108] In formula (22), η2 is the second weight coefficient; ω r is the actual yaw angular velocity; ρ2 is the smoothing weight coefficient of the exit process of the oblique driving condition, which can be set to 50 in this embodiment.

[0109] In summary, the calculation formula for the feedforward front wheel angle is:

[0110] δ ff =η1δ f_dc +η2δ f_oc +δf_s (twenty three)

[0111] In formula (23), δ ff is the feedforward front wheel angle; δ f_dc is the dynamic compensation front wheel angle; δ f_oc Compensate the front wheel angle for oblique driving; δ f_s To control the front wheel angle in a steady state.

[0112] The calculation formula of the feedforward rear wheel angle is:

[0113] δ fr =η1δ r_dc +η2δ r_oc +δ r_s (twenty four)

[0114] In formula (24), δ fr is the feedforward rear wheel angle; δ r_dc is the dynamic compensation rear wheel angle; δ r_oc Compensate the rear wheel angle for oblique driving; δ r_s To control the rear wheel angle in a steady state.

[0115] Step 3: Based on the feedforward wheel angle obtained in step 2, the motion state is analyzed to obtain the driver's desired motion state (including the reference center of mass side slip angle β ref and the reference yaw rate ω r_ref ), and calculate the vehicle state error (including the center of mass sideslip angle tracking error β) based on the expected motion state and the actual motion state e and yaw rate tracking error ω re ), and according to the vehicle state error, the state error feedback control is performed to obtain the additional yaw moment ΔM z .

[0116] This part takes the feedforward wheel angle, actual yaw rate, and actual sideslip angle as inputs, and uses the additional yaw moment as output. This part is divided into two steps. The first step is to analyze the driver's desired motion state to provide the vehicle's desired motion state for state error feedback control. To further improve vehicle driving performance, eliminate system steady-state errors, and enhance robustness to external disturbances and model parameter uncertainties, a closed-loop system state error feedback controller is established using a sliding mode control algorithm based on feedforward control.

[0117] (1) Analysis of driver's expected motion state

[0118] This section takes the feedforward wheel angle and road adhesion coefficient as input, and the reference yaw rate and reference center of mass sideslip angle as output. In the feedforward control, in order to meet the different control requirements of the driver, this embodiment adopts a feedforward control that integrates steady-state control, dynamic compensation and oblique compensation, wherein dynamic compensation is used to compensate for the dynamic changes of the driver, so there is no need to consider its impact on the ideal motion state of the vehicle. Therefore, in the analysis of the driver's expected motion state, only the impact of conventional steering and oblique compensation steering on the ideal motion state of the vehicle is considered. The driver's expected motion state under conventional steering conditions can be determined by formula (10). When the vehicle has oblique compensation, since the front and rear wheels have the same angle steering compensation, the ideal center of mass sideslip angle of the vehicle will no longer be zero. Substituting formulas (12), (13) and (20) into the vehicle dynamics model shown in formula (5), it is simplified to:

[0119]

[0120] In formula (25), the seventh matrix u oc =[δ f_oc ,δ r_oc ] T .

[0121] In formula (25), Then the expected operating state under the oblique compensation steering condition is obtained as follows:

[0122]

[0123] In formula (26), β d is the desired sideslip angle of the center of mass.

[0124] In addition, to ensure that the vehicle can generate sufficient yaw moment, considering the road adhesion limit, the reference state of the vehicle is corrected to:

[0125]

[0126] In formula (27), β ref is the reference center of mass sideslip angle; ω r_ref is the reference yaw rate. ref and ω r_ref The switching point between the two smaller terms in the calculation formula is when δ f_exp =δ f_max Established.

[0127] In summary, the driver's expected motion state during vehicle turning can be analyzed as follows:

[0128] 1) The calculation formula of the reference center of mass sideslip angle is:

[0129]

[0130] In formula (28), β ref is the reference center of mass sideslip angle; δ f_exp is the desired front wheel turning angle; δ f_max is the maximum front wheel turning angle; μ is the road adhesion coefficient; g is the acceleration due to gravity.

[0131] 2) The calculation formula for the reference yaw rate is:

[0132]

[0133] In formula (29), ω r_ref is the reference yaw angular velocity; ω rd is the desired yaw rate; ω r_max is the maximum yaw angular velocity.

[0134] (2) State error feedback control

[0135] This section takes the reference center of mass sideslip angle β ref , reference yaw angular velocity ω r_ref , actual center of mass sideslip angle β, actual yaw rate ω r The system tracking error is defined as e = [β e ,ω re ]=[β-β ref ,ω r -ω r_ref ] T , where e is the system tracking error, β e is the center of mass sideslip angle tracking error, ω re For the yaw rate tracking error, the discrete state error equation of the closed-loop system is constructed:

[0136]

[0137] In formula (30), k is the sampling time point; the matrix T s is the sampling time; the matrix Parameter u e =ΔM z , ΔM z is the additional yaw moment.

[0138] According to the yaw rate tracking error and the center of mass sideslip angle tracking error, the sliding mode surface of the sliding mode controller is constructed:

[0139]

[0140] In formula (31), s is the sliding surface; c is the yaw rate tracking error weight; d is the center of mass sideslip angle tracking error weight.

[0141] Substituting the discrete state error equation shown in formula (30) into the sliding surface shown in formula (31), we get:

[0142]

[0143] The reaching rate of the designed sliding mode controller is:

[0144] s(k+1)=-εT s sat(s(k))+(1-λT s )s(k)(33)

[0145] In formula (33), ε is the first sliding mode reaching rate parameter; λ is the second sliding mode reaching rate parameter.

[0146] To prevent controller jitter, a saturation function is used to achieve smooth switching of the controller near the sliding surface:

[0147]

[0148] In formula (34), Δ is the transition zone of the sliding surface and can be set to 0.01.

[0149] In summary, the sliding mode control law is:

[0150]

[0151] At this time, in step S2 , the desired front wheel steering angle is used as input, and the feedforward front wheel steering angle, the feedforward rear wheel steering angle, and the additional yaw moment are determined based on the feedforward control and feedback control.

[0152] The desired front wheel angle is used as input, and the feedforward front wheel angle, feedforward rear wheel angle, and additional yaw moment are determined based on feedforward control and feedback control, specifically including:

[0153] (1) Taking the desired front wheel steering angle as input, steady-state control is performed based on the vehicle dynamics model to obtain the steady-state controlled front wheel steering angle and the steady-state controlled rear wheel steering angle. The calculation formulas are (12) and (13).

[0154] (2) Taking the desired front wheel angle as input, dynamic compensation is performed based on the vehicle dynamics model and the vehicle kinematics model to obtain the dynamic compensation front wheel angle and the dynamic compensation rear wheel angle. The calculation formulas are (16) and (17).

[0155] (3) Taking the desired front wheel steering angle and road adhesion coefficient as input, the oblique running compensation is performed based on the vehicle kinematic model to obtain the oblique running compensation front wheel steering angle and the oblique running compensation rear wheel steering angle. The calculation formula is (20).

[0156] (4) Based on the steady-state control front wheel angle, the dynamic compensation front wheel angle and the oblique compensation front wheel angle, the feedforward front wheel angle is determined by hybrid feedforward control, and the calculation formula is formula (23). Based on the steady-state control rear wheel angle, the dynamic compensation rear wheel angle and the oblique compensation rear wheel angle, the feedforward rear wheel angle is determined by hybrid feedforward control, and the calculation formula is formula (24), thereby completing the feedforward control.

[0157] (5) Based on the feedforward front wheel angle and the feedforward rear wheel angle, the expected motion state is analyzed to obtain the expected motion state. The expected motion state includes the reference center of mass sideslip angle and the reference yaw angular velocity. The calculation formulas are (28) and (29).

[0158] (6) With the actual motion state and the desired motion state as input, the state error feedback control is performed based on the sliding mode controller to obtain the additional yaw moment. The calculation formula is (35), and the feedback control is completed. The actual motion state includes the actual center of mass sideslip angle and the actual yaw angular velocity.

[0159] Step 4: The additional yaw moment ΔM obtained in step 3 z , calculate the tire lateral force, and according to the calculation results, formulate the coordinated control strategy of four-wheel steering and four-wheel drive to obtain the feedback front wheel angle δ FWS , feedback rear wheel steering angle δ RWS and driving yaw moment ΔM 4WD , and according to the longitudinal force control process, the total longitudinal force F is obtained x , according to the driving yaw torque ΔM 4WD and the total longitudinal force F x , four-wheel torque distribution is achieved through the optimization model of four-wheel torque distribution.

[0160] This part takes the driver's desired longitudinal acceleration, additional yaw moment and the vehicle's key motion states (vehicle longitudinal acceleration, lateral acceleration, yaw angular velocity, four-wheel slip rate) as input, and the feedback of front and rear wheel angles (including feedback front wheel angle and feedback rear wheel angle) and four-wheel drive torque as output.

[0161] This section is divided into five steps, among which the tire lateral force calculation and tire model construction provide tire status information for the formulation of coordinated control strategy, the formulation of coordinated control strategy and longitudinal force control provide distribution targets for four-wheel torque distribution, and the four-wheel torque distribution is used to generate four-wheel drive torque.

[0162] The front and rear wheel angles, as well as the four-wheel torque, can all track the driver's desired motion state, but the control effects of each actuator system vary with the vehicle's state. Therefore, to fully utilize the control performance of each actuator system and improve vehicle stability under extreme conditions, this embodiment develops a four-wheel angle and four-wheel drive coordination strategy based on tire side slip state assessment. The three actuator systems work together to generate the required additional yaw torque to ensure vehicle stability. The relationship between each actuator system and the additional yaw torque is:

[0163]

[0164] In formula (36), ΔM z is the additional yaw moment; δ FWS is the feedback front wheel angle; FWS controller stands for front wheel angle controller; δ RWS is the feedback rear wheel steering angle; RWS controller stands for rear wheel steering angle controller; ΔM 4WD is the additional yaw moment generated by the four-wheel drive torque; 4WD controller stands for Four Wheel Torque Controller.

[0165] (1) Tire lateral force calculation

[0166] This section uses longitudinal acceleration, lateral acceleration, and yaw angular acceleration as inputs and four-wheel lateral force as output, providing data support for the development of coordinated control strategies. The vehicle's motion can be decomposed into two parts: the circular motion of the center of mass around the steering center and the rotational motion of the vehicle around the center of mass. Therefore, considering the vehicle's yaw motion characteristics, the vehicle's single-axis lateral acceleration is:

[0167]

[0168] In formula (37), a yf is the lateral acceleration of the front axle; a y is the lateral acceleration of the vehicle; is the yaw angular acceleration; a yr is the lateral acceleration of the rear axle.

[0169] Considering the influence of the vehicle's longitudinal and lateral acceleration, the calculation formula for the four-wheel vertical load is:

[0170]

[0171] In formula (38), F zfl is the vertical load on the left front wheel; g is the acceleration due to gravity; a x is the longitudinal acceleration of the vehicle; h g is the height of the center of mass; l w F is the wheelbase of the left and right wheels; zfr is the vertical load on the right front wheel; F zrlis the vertical load on the left rear wheel; F zrr is the vertical load on the right rear wheel.

[0172] The four-wheel equivalent load mass is obtained based on the four-wheel vertical load:

[0173]

[0174] In formula (39), m ij F is the mass of the j-th wheel on the i-th axle; zij is the vertical load of the j-th wheel on the i-th axle, including F zfl 、F zfr 、F zrl 、F zrr ; i represents the front and rear axle wheels, f represents the front axle, r represents the rear axle, j represents the left and right wheels, l represents the front wheel, and r represents the rear wheel.

[0175] According to Newton's second law, the lateral force of each wheel tire is:

[0176]

[0177] In formula (40), F yfl is the tire lateral force of the left front wheel; m fl is the mass of the left front wheel; a yf is the lateral acceleration of the front axle; F yfr is the tire lateral force of the right front wheel; m fr is the mass of the right front wheel; F yrl is the tire lateral force of the left rear wheel; m rl is the mass of the left rear wheel; a yr is the lateral acceleration of the rear axle; F yrr is the tire lateral force of the right rear wheel; m rr is the mass of the right rear wheel.

[0178] (2) Construction of tire model

[0179] This part takes tire slip rate and road adhesion coefficient as input and tire lateral force at the intersection of tire side slip state interval as output, providing evaluation threshold of tire side slip state for the formulation of coordinated control strategy.

[0180] In order to evaluate the side slip status of each tire, the magic tire formula model is introduced:

[0181]

[0182] In formula (41), F y (α) is the tire lateral force under the tire slip angle, α is the tire slip angle; s tire is the tire slip rate; D y B is the peak adhesion coefficient of the tire;y ,C y ,E y are all parameters, which are less affected by the vertical load and road adhesion coefficient. In this embodiment, they can be set as constants. Specifically, B y =7.098,C y =2.216,E y =1.009.

[0183] Tire peak adhesion coefficient D y It is greatly affected by the vertical load and can be expressed as:

[0184]

[0185] In formula (42), a0 and a1 are the fitting parameters of the vertical load, which are obtained by fitting the data of the tire lateral force under different vertical loads. Specifically, they can be a0 = -9.742 and a1 = 1056.

[0186] Considering the influence of the road adhesion coefficient, the tire lateral force expression under different road adhesion coefficients is:

[0187] F y =μF y (α / μ)(43)

[0188] In formula (43), μ is the road adhesion coefficient; F y (α / μ) is the tire lateral force calculated by replacing α in formula (41) with α / μ.

[0189] This embodiment is based on the tire test module in Carsim to obtain the tire sideways curve under different road adhesion coefficients. Figure 6 As can be seen from the figure, when the slip angle is small, the tire lateral force is linearly correlated with the slip angle, indicating that the tire has sufficient lateral force to fulfill the driver's intended steering. As the slip angle increases, the tire lateral force becomes nonlinearly correlated with the slip angle, and the lateral force gradually saturates. At this point, relying solely on the tire lateral force is insufficient to accurately track the driver's intended steering. Therefore, this embodiment divides the tire's slip state into three regions: a linear region, a nonlinear region, and a saturation region. Based on the tire's slip characteristics, a slip angle less than 1 degree is defined as a strictly linear region. When the slope of the lateral force versus slip angle decreases by less than 20% compared to the strictly linear region, it remains in the linear region. When the slope decreases by more than 20%, it is defined as a nonlinear region. When the slip angle exceeds the slip angle corresponding to the maximum lateral force, the tire is considered to be in the saturation region. The intersection points of the tire slip state intervals are primarily affected by the road adhesion coefficient. The slip angles corresponding to the intersection points of each tire slip state interval are shown in Table 1.

[0190] Table 1 Intersection points of tire sideways state intervals

[0191] Road adhesion coefficient μ 0.2 0.4 0.6 0.8 1.0 Linear region / nonlinear region intersection point (deg) 0.45 0.80 1.15 1.50 1.90 Nonlinear region / saturation region intersection point (deg) 1.85 3.70 5.60 7.45 9.30

[0192] (3) Development of coordinated control strategies

[0193] This part takes the tire lateral force and the tire lateral force at the intersection of the tire sideways state interval as input, and the feedback of the front and rear wheel angles and the driving yaw moment as output. This embodiment designs a coordinated control strategy for four-wheel steering and four-wheel drive by evaluating the sideways state of the tire. The flow chart is as follows Figure 7 The specific calculation process is as follows.

[0194] 1) Calculation of tire lateral force at the intersection of tire cornering state intervals: Determine the tire lateral force at the intersection of different tire cornering state intervals based on tire slip rate, vertical load, and road adhesion coefficient.

[0195] 2) Calculation of tire lateral force: Calculate the tire lateral force of the four wheels separately according to formula (40).

[0196] 3) Single wheel control weight design: Design the single wheel drive torque control weight η based on the difference between each tire lateral force and the tire lateral force at the intersection of different tire sideways state intervals 4WD_ij :

[0197]

[0198] In formula (44), F y (α) is the tire lateral force of the j-th wheel on the i-th axis, which is equal to the product of the mass of the j-th wheel on the i-th axis and the lateral acceleration of the i-th axis, and is calculated using formula (40). α is the tire slip angle; F y (α1) is the tire lateral force at the first tire slip angle, which is calculated using the magic tire formula model with the first tire slip angle and the road adhesion coefficient as input. Specifically, it is calculated using formula (43), where α1 is the first tire slip angle, which is selected according to Table 1; F y (α2) is the tire lateral force at the second tire slip angle, which is calculated using the magic tire formula model with the second tire slip angle and the road adhesion coefficient as input. Specifically, it is calculated using formula (43). α2 is the second tire slip angle, which is selected according to Table 1.

[0199] 4) Torque control weight calculation: Compare the single-wheel drive torque control weights of the four wheels and select the minimum value as the final initial wheel torque control weight:

[0200] η 4WD1 =min(η 4WD_ij )(45)

[0201] In formula (45), η 4WD1 is the initial wheel torque control weight; η 4WD_ij is the single-wheel drive torque control weight of the j-th side wheel on the i-th axis, where the i-th axis includes the front axle and the rear axle, and the j-th side includes the left and right sides.

[0202] It should be noted that according to the tire adhesion ellipse theory, when the road adhesion coefficient is low, the tire force margin is small, and the additional yaw moment generated by the tire longitudinal force is small. In this case, the direction of the wheel force can be changed by adjusting the wheel angle, thereby generating a larger additional yaw moment. Therefore, this embodiment uses the road adhesion coefficient as the limit for the angle and torque distribution weight. In this case, the wheel torque control weight is:

[0203] η 4WD =min(η 4WD1 ,μ)(46)

[0204] In formula (46), η 4WD is the wheel torque control weight; μ is the road adhesion coefficient.

[0205] 5) Calculation of steering angle control weights: Based on this, it is necessary to further determine the control weights of the front and rear wheel steering angles. Considering the vehicle's steering characteristics, the yaw rate tracking error and the steering wheel angle are used as inputs to design the control weights of the front and rear wheel steering angles. The calculation formula for the feedback front wheel steering angle control weight is:

[0206] η FWS =(1-η 4WD )(tanh(e ωr *δ sw / 2)+1) / 2(47)

[0207] In formula (47), η FWS Feedback front wheel angle control weight; e ωr is the yaw rate tracking error; δ sw is the steering wheel angle.

[0208] The calculation formula for the feedback rear wheel steering angle control weight is:

[0209] η RWS =1-η 4WD -η FWS (48)

[0210] In formula (48), η RWS Feedback rear wheel angle control weight.

[0211] 6) Calculation of rotation angle and torque:

[0212] The calculation formula of the driving yaw torque is:

[0213] ΔM4WD =ΔM z η 4WD (49)

[0214] In formula (49), ΔM 4WD is the driving yaw torque; ΔM z is the additional yaw moment; η 4WD is the wheel torque control weight.

[0215] The calculation formula for the feedback front wheel angle is:

[0216]

[0217] In formula (50), δ FWS is the feedback front wheel angle; η FWS is the feedback front wheel steering angle control weight; a is the distance from the vehicle center of mass to the front axle; k f is the front axle cornering stiffness.

[0218] The calculation formula for the feedback rear wheel angle is:

[0219]

[0220] In formula (51), δ RWS is the feedback rear wheel steering angle; η RWS is the feedback rear wheel steering angle control weight; b is the distance from the vehicle center of mass to the rear axle.

[0221] (4) Longitudinal force control

[0222] This section uses the driver's desired longitudinal acceleration as input and the total longitudinal force as output. To ensure that the vehicle can effectively track the driver's desired speed, this embodiment will directly obtain the driver's desired longitudinal acceleration input from Carsim and calculate the total longitudinal force required by the vehicle based on it:

[0223] F x =ma x_req (52)

[0224] In formula (52), F x is the total longitudinal force; a x_req The desired longitudinal acceleration entered by the driver.

[0225] (5) Four-wheel torque distribution

[0226] This section uses total longitudinal force, driving yaw moment, and four-wheel slip as inputs, and four-wheel drive torque as output. In the lower-level torque distribution, the total longitudinal force and additional yaw moment output from the upper level are used as tracking targets. At the same time, to prevent wheel slip / spin, the four-wheel tire slip rate is used as one of the optimization objective functions. Based on this, considering hard constraints such as the maximum torque of the motor and road adhesion limits, an optimization model for four-wheel torque distribution is established:

[0227] minJ=W1(v-Bu)+W2(s tire )u(53)

[0228]

[0229] In the above formula, J is the objective function; W1 is the control weight of the tracking target, W1 = [0.001, 0.001]; v is the tracking target, v = [F x ΔM 4WD ] T , including the total longitudinal force F x and driving yaw moment ΔM 4WD ; B is the control matrix,

[0230] δ f is the actual front wheel turning angle, l w is the wheelbase of the left and right wheels, δ r is the actual rear wheel turning angle; u is the control variable, including the four-wheel longitudinal force, u=[F xfl F xfr F xrl F xrr ] T , F xfl 、F xfr 、F xrl 、F xrr They are the tire longitudinal force of the left front wheel, the tire longitudinal force of the right front wheel, the tire longitudinal force of the left rear wheel, and the tire longitudinal force of the right rear wheel respectively; W2 is the control weight of the control variable, including the control weight of the four-wheel longitudinal force, and the control weight of the four-wheel longitudinal force includes the control weight of the single-wheel longitudinal force of each wheel in the four wheels, W2=[W s (s tire_fl ),W s (s tire_fr ),W s (s tire_rl ),W s (s tire_rr )] T , W s (s tire_fl ), W s (s tire_fr ), W s (stire_rl ), W s (s tire_rr ) are the control weights of the single-wheel longitudinal force of the left front wheel, the control weight of the single-wheel longitudinal force of the right front wheel, the control weight of the single-wheel longitudinal force of the left rear wheel, and the control weight of the single-wheel longitudinal force of the right rear wheel, respectively. tire_fl 、s tire_fr 、s tire_rl 、s tire_rr They are the tire slip rate of the left front wheel, the tire slip rate of the right front wheel, the tire slip rate of the left rear wheel and the tire slip rate of the right rear wheel respectively; s tire is the tire slip rate; T max is the maximum torque that the drive motor can output; R tire is the tire radius; F xij is the longitudinal force of the tire on the j-th side of the i-th axis, where the i-th axis includes the front and rear axles, and the j-th side includes the left and right sides; μ is the road adhesion coefficient; F zij is the vertical load of the wheel on the j-th side of the i-th axle, where the i-th axle includes the front axle and the rear axle, and the j-th side includes the left and right sides.

[0231] The calculation formula for the control weight of the single wheel longitudinal force is:

[0232]

[0233] In formula (55), W s (s tire ) is the control weight of the single wheel longitudinal force.

[0234] Finally, the above optimization model is solved using the effective set method to obtain the four-wheel drive torque:

[0235] T ij =F xij R tire (56)

[0236] In formula (56), T ij is the driving torque of the j-th wheel on the i-th axis; F xij is the longitudinal force of the tire on the j-th side of the i-th axis.

[0237] At this time, in step S3, the first tire slip angle and the second tire slip angle are determined based on the road adhesion coefficient. The first tire slip angle is the critical tire slip angle between the linear region and the nonlinear region in the tire slip curve, and the second tire slip angle is the critical tire slip angle between the nonlinear region and the saturated region in the tire slip curve. They are specifically determined based on Table 1. For example, when the road adhesion coefficient is 0.2, the first tire slip angle is 0.45 and the second tire slip angle is 1.85. The wheel torque control weight, the feedback front wheel steering angle control weight, and the feedback rear wheel steering angle control weight are determined based on the first tire slip angle and the second tire slip angle, and are specifically calculated using Formulas (46), (47), and (48).

[0238] In step S4, the driving yaw torque, the feedback front wheel angle, and the feedback rear wheel angle are determined based on the additional yaw torque, the wheel torque control weight, the feedback front wheel angle control weight, and the feedback rear wheel angle control weight, and are specifically calculated using equations (49), (50), and (51).

[0239] In step S5, the total longitudinal force is determined based on the expected longitudinal acceleration input by the driver, specifically calculated using formula (52), and the four-wheel torque is distributed based on the total longitudinal force and the driving yaw moment to obtain the four-wheel drive torque.

[0240] Among them, the four-wheel torque distribution is performed based on the total longitudinal force and the driving yaw moment to obtain the four-wheel drive torque, which specifically includes: establishing an optimization model for the four-wheel torque distribution, namely, equations (53) and (54); using the total longitudinal force and the driving yaw moment as input, solving the optimization model for the four-wheel torque distribution to perform four-wheel torque distribution and obtain the four-wheel longitudinal force; calculating the product of the four-wheel longitudinal force and the tire radius to obtain the four-wheel drive torque, which is specifically calculated by equation (56).

[0241] This embodiment relates to a coordinated control method of four-wheel steering and four-wheel drive, which mainly solves the problems of expected motion state analysis and chassis coordinated control of four-wheel steering and four-wheel drive vehicles, and provides technical support for chassis control of high-performance four-wheel steering and four-wheel drive vehicles.

[0242] Related solution one is: a four-wheel steering and four-wheel drive intelligent chassis stability control method, including: ① obtaining the front wheel angle and requested vehicle speed information input by the driver; determining the ideal center of mass sideslip angle and ideal yaw rate of the chassis based on the front wheel angle and longitudinal vehicle speed; ② according to the ideal and actual values ​​of the center of mass sideslip angle and yaw rate of the chassis, through adaptive terminal sliding mode control, the rear wheel angle and additional yaw moment of the chassis required to maintain vehicle stability are obtained; ③ performing vehicle speed tracking control, according to the vehicle speed and longitudinal vehicle speed requested by the driver, the longitudinal torque required to maintain the requested speed is determined; ④ performing driving force distribution control, using a sequential quadratic programming algorithm to obtain the four-wheel hub motor torque that meets the longitudinal torque and additional yaw moment requirements. However, the first related solution uses adaptive terminal sliding mode control to achieve integrated control of the rear wheel steering angle and additional yaw moment. However, it does not fully consider the kinematic advantages of special steering modes (such as oblique steering) of four-wheel steering vehicles under extreme conditions. It also ignores the difference between the driver's expected motion state of four-wheel steering vehicles and front-wheel steering vehicles, resulting in poor vehicle stability under extreme conditions. This embodiment redesigns the feedforward control and feedback control, simultaneously considering steady-state control, oblique steering compensation, and dynamic compensation in the feedforward control, and introducing a kinematic model in the feedback control to analyze the expected motion state, thereby solving the above problems, improving stability, and reducing the driver's burden.

[0243] Related solution 2 is a phase-plane-based vehicle stability determination method, chassis coordinated control method, and system, including: ① obtaining vehicle intrinsic parameters and motion parameters; ② introducing a nonlinear tire model and combining it with a linear two-degree-of-freedom model to obtain a nonlinear two-degree-of-freedom model; ③ setting the domain of the front wheel steering angle, road adhesion coefficient, and phase-plane stability boundary function; ④ using the front wheel steering angle and road adhesion coefficient as inputs and deriving the stability boundary function for the current operating condition through fuzzy reasoning; ⑤ dividing the phase plane into three control domains based on the stability boundary, using the sideslip angle and its derivative obtained by the observer to determine the current vehicle state; and ⑥ calculating the total additional yaw moment based on the current vehicle stability state and distributing it to the subsystems. However, Related Solution 2 employs a phase-plane-based vehicle stability determination and chassis coordinated control method, utilizing the phase plane to determine the vehicle's stable motion state and adopting different coordinated control strategies based on the vehicle's stability state. This method requires extensive calibration and is difficult to adapt to complex and changing traffic scenarios. This embodiment only requires calibration of the road adhesion coefficient, reducing calibration effort and making it adaptable to complex and changing traffic scenarios.

[0244] This embodiment addresses the challenge of efficient coordination for four-wheel-steer (4WD) vehicles by proposing a global coordinated control architecture based on tire side slip state assessment. First, combining driver dynamic input with the vehicle kinematic model, a hybrid feedforward control method encompassing steady-state control, dynamic compensation, and lean compensation is proposed. Second, a desired motion state analysis method integrating conventional steering and lean steering is established using the kinematic model to obtain the driver's desired motion state. A sliding mode control algorithm is then employed to obtain the additional yaw torque required to track the desired motion state. Based on this, a weighted distribution strategy for front-wheel steering, rear-wheel steering, and the additional yaw torque is developed based on vehicle state information such as lateral acceleration and yaw angular acceleration. Finally, to prevent tire slip / slip, a four-wheel torque optimization distribution method is established to reduce tire slip rate and tracking error of the upper-level control variable. This achieves efficient coordination for 4WD vehicles under different driving conditions, improving vehicle handling stability.

[0245] In order to solve the problem of efficient coordinated control of the chassis of a four-wheel steering four-wheel drive vehicle, this embodiment proposes a global coordinated control architecture based on tire side slip state evaluation, which has the following advantages: (1) A coordinated control method of four-wheel steering and four-wheel drive that combines feedforward and feedback is proposed, which improves the vehicle's tracking accuracy of driving intentions; (2) A method for analyzing the expected motion state under steering modes such as conventional steering and oblique steering is proposed, which reduces the driver's driving burden and improves the vehicle's stability under extreme working conditions; (3) A four-wheel steering and four-wheel drive coordination strategy based on tire side slip state evaluation is proposed, which realizes efficient coordination of four-wheel steering and four-wheel drive vehicles under different driving conditions and improves the vehicle's driving stability under extreme working conditions.

[0246] Electric vehicles equipped with four-wheel steering and four-wheel drive can effectively improve vehicle safety, stability, and maneuverability through coordinated control of four-wheel torque and steering angle. However, existing research methods fail to fully consider the unique steering mode advantages of four-wheel steering vehicles under extreme conditions, limiting the vehicle's driving stability under these conditions. Furthermore, existing methods require extensive calibration and are difficult to adapt to complex and changing traffic scenarios. To address these issues, this embodiment proposes a global coordinated control architecture based on tire side slip state assessment. First, based on the driver's dynamic input, a hybrid feedforward control method is proposed, which includes steady-state control, dynamic compensation, and oblique steering compensation. Secondly, a driving intention analysis method that integrates conventional steering and oblique steering is established using a kinematic model to obtain the driver's desired motion state. Then, a sliding mode control algorithm is used to obtain the additional yaw torque required to track the desired motion state. On this basis, a weight distribution strategy for front-wheel steering, rear-wheel steering, and driving yaw torque is developed based on vehicle state information such as lateral acceleration and yaw angular acceleration. Finally, to avoid tire slip / spin, a four-wheel torque optimization distribution method is established with the goal of reducing tire slip rate and upper-level control quantity tracking error, effectively improving the vehicle's driving stability during extreme obstacle avoidance.

[0247] The present application also provides an application scenario, which applies the above-mentioned chassis coordination control method based on tire side slip state evaluation. Specifically, the chassis coordination control method based on tire side slip state evaluation provided in this embodiment can be applied in a vehicle driving scenario. The vehicle driving scenario includes a parameter determination link and a control link. The parameter determination link is used to output the feedforward front wheel angle, the feedforward rear wheel angle, the feedback front wheel angle, the feedback rear wheel angle and the four-wheel drive torque. The control link is used to control the vehicle driving process based on the feedforward front wheel angle, the feedforward rear wheel angle, the feedback front wheel angle, the feedback rear wheel angle and the four-wheel drive torque. The chassis coordination control method based on tire side slip state evaluation provided in this embodiment belongs to the parameter determination link.

[0248] Example 2

[0249] In an exemplary embodiment, a computer system is provided, including a memory and a processor. The memory stores a computer program, and the processor implements the chassis coordinated control method based on tire side slip state evaluation in Example 1 when executing the computer program.

[0250] Example 3

[0251] In an exemplary embodiment, a computer-readable storage medium is provided, storing a computer program. When the computer program is executed by a processor, the method for chassis coordinated control based on tire cornering state evaluation in embodiment 1 is implemented.

[0252] Example 4

[0253] In an exemplary embodiment, a computer program product is provided, including a computer program. When the computer program is executed by a processor, the method for chassis coordinated control based on tire cornering state evaluation in embodiment 1 is implemented.

[0254] It should be noted that the user information (including but not limited to user device information, user personal information, etc.) and data (including but not limited to data used for analysis, stored data, displayed data, etc.) involved in this application are all information and data authorized by the user or fully authorized by all parties, and the collection, use and processing of relevant data must comply with relevant regulations.

[0255] The technical features of the above embodiments can be combined arbitrarily. To make the description concise, not all possible combinations of the technical features in the above embodiments are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0256] This document uses specific examples to illustrate the principles and implementation methods of this application. The description of the above examples is only intended to help understand the method and core concept of this application. At the same time, for those skilled in the art, based on the concept of this application, there may be changes in the specific implementation methods and application scope. In summary, the content of this specification should not be understood as limiting this application.

Claims

1. A chassis coordinated control method based on tire side slip state evaluation, characterized in that: The chassis coordinated control method based on tire side slip state evaluation includes: determining a desired front wheel angle based on a desired steering wheel angle input by the driver; Taking the desired front wheel steering angle as input, determining a feedforward front wheel steering angle, a feedforward rear wheel steering angle, and an additional yaw moment based on feedforward control and feedback control; determining a first tire slip angle and a second tire slip angle based on a road adhesion coefficient, and determining a wheel torque control weight, a feedback front wheel steering angle control weight, and a feedback rear wheel steering angle control weight based on the first tire slip angle and the second tire slip angle; the first tire slip angle being a critical tire slip angle between a linear region and a nonlinear region in a tire slip curve, and the second tire slip angle being a critical tire slip angle between a nonlinear region and a saturated region in the tire slip curve; determining a driving yaw moment, a feedback front wheel angle, and a feedback rear wheel angle based on the additional yaw moment, the wheel torque control weight, the feedback front wheel angle control weight, and the feedback rear wheel angle control weight; Based on the expected longitudinal acceleration input by the driver, a total longitudinal force is determined, and four-wheel torque distribution is performed based on the total longitudinal force and the driving yaw moment to obtain a four-wheel driving torque.

2. The chassis coordinated control method based on tire side slip state evaluation according to claim 1, characterized in that: Taking the desired front wheel steering angle as input, determining the feedforward front wheel steering angle, the feedforward rear wheel steering angle, and the additional yaw moment based on feedforward control and feedback control specifically includes: Taking the desired front wheel steering angle as input, performing steady-state control based on a vehicle dynamics model to obtain a steady-state controlled front wheel steering angle and a steady-state controlled rear wheel steering angle; Taking the desired front wheel steering angle as input, dynamic compensation is performed based on a vehicle dynamics model and a vehicle kinematics model to obtain a dynamic compensated front wheel steering angle and a dynamic compensated rear wheel steering angle; Taking the desired front wheel steering angle and the road adhesion coefficient as input, performing slant compensation based on a vehicle kinematic model to obtain a slant-compensated front wheel steering angle and a slant-compensated rear wheel steering angle; Based on the steady-state controlled front wheel angle, the dynamic compensation front wheel angle, and the slant compensation front wheel angle, a feedforward front wheel angle is determined by hybrid feedforward control; based on the steady-state controlled rear wheel angle, the dynamic compensation rear wheel angle, and the slant compensation rear wheel angle, a feedforward rear wheel angle is determined by hybrid feedforward control to complete feedforward control; performing a desired motion state analysis based on the feedforward front wheel angle and the feedforward rear wheel angle to obtain a desired motion state; the desired motion state includes a reference center of mass sideslip angle and a reference yaw rate; The actual motion state and the desired motion state are used as inputs, and state error feedback control is performed based on a sliding mode controller to obtain an additional yaw moment and complete feedback control; the actual motion state includes an actual center of mass sideslip angle and an actual yaw angular velocity.

3. The chassis coordinated control method based on tire side slip state evaluation according to claim 2, characterized in that: The calculation formula for the steady-state control front wheel angle is: d f_s =d f_exp ; Among them, δ f_s To control the front wheel angle in a steady state; δ f_exp is the desired front wheel turning angle; The calculation formula for the steady-state control rear wheel angle is: Among them, δ r_s is the rear wheel steering angle for steady-state control; a is the distance from the vehicle's center of mass to the front axle; b is the distance from the vehicle's center of mass to the rear axle; k f is the front axle cornering stiffness; k r is the rear axle lateral stiffness; m is the vehicle mass; v x is the longitudinal velocity of the vehicle; The calculation formula for the dynamic compensation front wheel angle is: Among them, δ f_dc To dynamically compensate for the front wheel angle; I z is the yaw moment of inertia of the vehicle; is the desired yaw angular acceleration; in, is the desired front wheel angular velocity; is the desired rear wheel angular velocity; δ r_exp is the desired rear wheel turning angle; The calculation formula for the dynamic compensation rear wheel angle is: Among them, δ r_dc To dynamically compensate for rear wheel angle; The calculation formulas for the oblique driving compensation front wheel steering angle and the oblique driving compensation rear wheel steering angle are as follows: d f_oc =d r_oc =d f_exp -d f_max ; Among them, δ f_oc Compensate the front wheel angle for oblique driving; δ r_oc Compensate the rear wheel angle for oblique driving; δ f_max is the maximum front wheel turning angle; Among them, ω r_max is the maximum yaw angular velocity.

4. The chassis coordinated control method based on tire side slip state evaluation according to claim 2, characterized in that: The calculation formula of the feedforward front wheel angle is: d ff =η1δ f_dc +η2δ f_oc +d f_s ; Among them, δ ff is the feedforward front wheel angle; η1 is the first weight coefficient; δ f_dc is the dynamic compensation front wheel angle; η2 is the second weight coefficient; δ f_oc Compensate the front wheel angle for oblique driving; δ f_s To control the front wheel angle in a steady state; The calculation formula of the feedforward rear wheel angle is: d fr =η1δ r_dc +η2δ r_oc +d r_s ; Among them, δ fr is the feedforward rear wheel angle; δ r_dc is the dynamic compensation rear wheel angle; δ r_oc Compensate the rear wheel angle for oblique driving; δ r_s To control the rear wheel angle in a steady state; Among them, ω rd is the desired yaw rate; ω r_tra is the yaw rate control boundary value; ρ1 is the correction coefficient of control weight distribution; Among them, ω r is the actual yaw angular velocity; ρ2 is the smoothing weight coefficient of the exit process of the oblique working condition.

5. The chassis coordinated control method based on tire side slip state evaluation according to claim 2, characterized in that: The calculation formula of the reference center of mass sideslip angle is: Among them, β ref is the reference center of mass sideslip angle; δ f_exp is the desired front wheel turning angle; δ f_max is the maximum front wheel turning angle; μ is the road adhesion coefficient; g is the acceleration due to gravity; The calculation formula of the reference yaw rate is: Among them, ω r_ref is the reference yaw angular velocity; ω rd is the desired yaw rate; ω r_max is the maximum yaw angular velocity.

6. The chassis coordinated control method based on tire side slip state evaluation according to claim 1, characterized in that: The calculation formula of wheel torque control weight is: or 4WD =min(η 4WD1 ,m); Among them, η 4WD is the wheel torque control weight; η 4WD1 is the initial wheel torque control weight; μ is the road adhesion coefficient; or 4WD1 =min(η 4WD_ij ); Among them, η 4WD_ij is the single-wheel drive torque control weight of the j-th side wheel on the i-th axis, where the i-th axis includes the front and rear axles, and the j-th side includes the left and right sides; Among them, F y (α) is the lateral force of the tire on the j-th side of the i-th axis, which is equal to the product of the mass of the j-th side of the i-th axis and the lateral acceleration of the i-th axis. α is the tire slip angle; F y (α1) is the tire lateral force at the first tire slip angle, which is calculated using the Magic Tire Formula model with the first tire slip angle and the road adhesion coefficient as input, where α1 is the first tire slip angle; F y (α2) is the tire lateral force at the second tire slip angle, calculated using the Magic Tire Formula model with the second tire slip angle and the road adhesion coefficient as input, where α2 is the second tire slip angle; The calculation formula for the feedback front wheel steering angle control weight is: or FWS =(1-th 4WD (tanh(e) ωr *d sw / 2)+1) / 2; Among them, η FWS Feedback front wheel angle control weight; e ωr is the yaw rate tracking error; δ sw is the steering wheel angle; The calculation formula for the feedback rear wheel steering angle control weight is: or RWS =1st 4WD -or FWS ; Among them, η RWS Feedback rear wheel angle control weight.

7. The chassis coordinated control method based on tire side slip state evaluation according to claim 1, characterized in that: The calculation formula of the driving yaw torque is: DM 4WD =ΔM z or 4WD ; Among them, ΔM 4WD is the driving yaw moment; ΔM z is the additional yaw moment; η 4WD is the wheel torque control weight; The calculation formula for the feedback front wheel angle is: Among them, δ FWS is the feedback front wheel angle; η FWS is the feedback front wheel steering angle control weight; a is the distance from the vehicle center of mass to the front axle; k f is the front axle lateral stiffness; The calculation formula for the feedback rear wheel angle is: Among them, δ RWS is the feedback rear wheel steering angle; η RWS is the feedback rear wheel steering angle control weight; b is the distance from the vehicle center of mass to the rear axle.

8. The chassis coordinated control method based on tire side slip state evaluation according to claim 1, characterized in that: The four-wheel torque distribution is performed based on the total longitudinal force and the driving yaw moment to obtain the four-wheel driving torque, specifically including: Establish an optimization model for four-wheel torque distribution; Solving a four-wheel torque distribution optimization model using the total longitudinal force and the driving yaw moment as input to perform four-wheel torque distribution and obtain four-wheel longitudinal forces; Calculating the product of the four-wheel longitudinal force and the tire radius to obtain the four-wheel drive torque; Among them, the optimization model of four-wheel torque distribution is: minJ=W1(v-Bu)+W2(s tire )u; Where J is the objective function; W1 is the control weight of the tracking target; v is the tracking target, including the total longitudinal force and the driving yaw moment; B is the control matrix; u is the control variable, including the four-wheel longitudinal force; W2 is the control weight of the control variable; s tire is the tire slip rate; T max is the maximum torque that the drive motor can output; R tire is the tire radius; F xij is the longitudinal force of the tire on the j-th side of the i-th axis, where the i-th axis includes the front and rear axles, and the j-th side includes the left and right sides; μ is the road adhesion coefficient; F zij is the vertical load on the j-th wheel on the i-th axle.

9. The chassis coordinated control method based on tire side slip state evaluation according to claim 8, characterized in that: The control weights of the control variables include the control weights of the four-wheel longitudinal forces. The control weights of the four-wheel longitudinal forces include the control weights of the single-wheel longitudinal forces of each of the four wheels. The calculation formula for the control weights of the single-wheel longitudinal forces is: Among them, W s (s tire ) is the control weight of the single wheel longitudinal force.

10. A computer system comprising: A memory, a processor, and a computer program stored in the memory and executable on the processor, wherein the processor executes the computer program to implement the chassis coordination control method based on tire side slip state evaluation according to any one of claims 1 to 9.

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