Integrated control system for active rear wheel steering and differential braking coordination of multi-axle special vehicles

Abstract

The application discloses a kind of multi-axle special vehicle active rear wheel steering and differential brake coordination integrated control system, belong to the field of automobile electronic control technology, including signal input module, for the decision module and distribution module corresponding signal input quantity;Decision module, for the collaborative control instruction of each sub-controller in distribution module is decided, and collaborative control instruction is sent to distribution module;Distribution module, for receiving collaborative control instruction that decision module decides, and the active steering angle of active steering axle wheel and the brake torque of each wheel are calculated, then each control quantity calculated is sent to execution module;Execution module, for according to the Trucksim dynamics model of multi-axle special vehicle, each control quantity sent by distribution module is executed, and the motion state of vehicle is fed back, the application can solve how to based on the active safety technology of chassis to improve the driving stability of multi-axle special vehicle technical problem.

Description

Technical Field

[0001] This invention belongs to the field of automotive electronic control technology, specifically relating to an integrated control system for coordinated active rear-wheel steering and differential braking in multi-axle special vehicles. Background Technology

[0002] In recent years, vehicle stability control (VSC) has become a key research area in vehicle chassis safety design. Many scholars both domestically and internationally have proposed advanced control methods, among which the four most widely used are differential braking, active steering, active suspension, and active anti-roll bars. Differential braking and active steering both control the tire forces (lateral forces and braking forces) acting on the vehicle body to control and influence vehicle roll and yaw movements. Active suspension and active anti-roll bars, on the other hand, improve the vehicle's dynamic state by controlling the vehicle's attitude and movement.

[0003] With the development of vehicle electronics technology, single control methods centered around a single actuator can no longer adequately meet the requirements of vehicle stability control. A single actuator is often limited by its capabilities, thus affecting the control effect of the control system. Currently, many scholars are dedicated to integrating multiple control methods to control vehicle stability, utilizing the complementary functions of various control methods to broaden the control range and effectiveness of the control system. For example, Li Shaohua et al. conducted research on the integrated stability control of three-axle heavy vehicles based on active proportional steering control (6WS) and direct yaw moment control (DYC). The results showed that the control effect of 6WS+DYC on vehicle stability was significantly better than that of 6WS control and DYC control alone. Sun Chuanbin et al. used the Takagi-Sugeon (TS) method to establish a 3-DOF yaw and roll model of the vehicle and conducted integrated control research on active front wheel steering (AFS) and direct yaw moment control (DYC). Based on the designed state feedback fuzzy distributed controller (parallel distributed compensation-TS, PDC-TS), they effectively improved the utilization rate of actuator capabilities and the control effect of vehicle stability under strong nonlinear steering. Fu Xiang et al. designed a feedforward and feedback active rear-wheel steering controller with zero center of gravity sideslip angle as the control target, and coordinated it with a designed four-wheel torque controller to effectively improve the vehicle's driving stability and safety. Narjes Ahmadian et al. achieved adaptive coordination of AFS and DYC based on a scheduling mechanism of stability index, which can effectively reduce the vehicle's tire sideslip angle and lateral acceleration, and improve the vehicle's handling stability and path following performance under different maneuvering conditions.

[0004] Modern special-purpose vehicles are mostly characterized by large curb weight, high center of gravity, long body, and narrow wheelbase. In wartime, facing complex driving environments, they are prone to instability such as rollover and tilting. Furthermore, some models of special-purpose vehicles lack active steering capabilities for their rear wheels, and the large number of axles and complex actuators have resulted in insufficient attention being paid to vehicle stability control technologies, especially integrated control technologies. Therefore, research on improving the driving stability of multi-axle special-purpose vehicles based on chassis-based active safety technologies is of great significance. Summary of the Invention

[0005] The purpose of this invention is to overcome the shortcomings of the prior art and provide an integrated control system for coordinated active rear-wheel steering and differential braking of multi-axle special vehicles.

[0006] To achieve the above objectives, the technical solution adopted by the present invention is as follows:

[0007] An integrated control system for coordinated active rear-wheel steering and differential braking in a multi-axle special vehicle, comprising:

[0008] The signal input module is used to provide the corresponding signal input quantities to the decision-making module and the allocation module;

[0009] The decision module is used to determine the collaborative control instructions for each sub-controller in the allocation module and send the collaborative control instructions to the allocation module.

[0010] The allocation module is used to receive the coordinated control instructions decided by the decision module, calculate the active steering angle of the active steering axle wheel and the braking torque of each wheel, and then send the calculated control quantities to the execution module.

[0011] The execution module is used to execute the control quantities sent by the allocation module according to the Trucksim dynamics model of the multi-axle special vehicle, and to provide feedback on the vehicle's motion state, so that the integrated control system forms a closed loop.

[0012] Preferably, the signal input module calculates the desired motion state of the vehicle based on the vehicle's 2-DOF vehicle dynamics model, and uses this as the signal input quantity between the decision module and the allocation module.

[0013] Preferably, the 2-DOF vehicle dynamics model is as follows:

[0014]

[0015] In the formula: The derivative of the centroid sideslip angle. Let J be the derivative of the yaw rate, m be the mass of the vehicle, and J be the weight of the vehicle. zz Let l be the moment of inertia of the entire vehicle about the z-axis of the vehicle body coordinate system. i (i = 1, 2, 3, 4, 5) are variables with positive and negative values, representing the distance from the center of the wheel on the i-th axle to the center of mass. When the axle is in front of the vehicle's center of mass, l i When the axle is positive and located behind the center of gravity, l i If negative, C i (i = 1, 2, 3, 4, 5) represents the equivalent lateral stiffness of the wheel on the i-th axle, which is equal to the sum of the lateral stiffness of the left and right tires, i.e., C. i =C il +C ir v x Let δ be the longitudinal velocity of the vehicle. i (i = 1, 2, 3, 4, 5) represents the steering angle of the wheel on the i-th axle.

[0016] Preferably, the sub-controller of the distribution module includes an active steering controller and a differential braking torque controller. The differential braking torque controller includes a yaw rate sliding film controller and a torque distribution controller. The active steering controller calculates the active steering angle of the vehicle's wheels based on the Ackermann steering model and the PID feedback controller. The yaw rate sliding film controller calculates the additional yaw torque required to maintain the yaw stability of the vehicle. The torque distribution controller calculates the braking torque distributed to each axle wheel and then sends the calculated control quantities to the execution module.

[0017] Preferably, the Ackermann steering model is:

[0018]

[0019] Where: δ i(l，r) Let l be the steering angle of the left and right wheels on the i-th axis. i Let be the distance from the center of the i-th axle to the center of mass.

[0020] The initial steering angles of the four-axle and five-axle wheels are:

[0021]

[0022] Where: δ i(l，r)0 (i = 4, 5) represents the initial active steering angle of the left and right wheels of the four-axle and five-axle vehicles, B is the width of the axle, and δ sw i is the steering wheel angle. sw This is the transmission ratio between the steering wheel and one of the wheels on the axle.

[0023] Preferably, the feedback output of the PID feedback controller is the additional active steering angle of the four-axle and five-axle wheels, and the calculation formula for the additional active steering angle of the four-axle and five-axle wheels is as follows:

[0024]

[0025] In the formula: P1, I1, and D1 are the proportional coefficient, integral coefficient, and derivative coefficient of the PID controller, respectively, e β This represents the deviation of the centroid sideslip angle;

[0026] The total input of the active steering angle of the four-axle and five-axle wheels is:

[0027] δ i(l，r) =δ i(l，r)0 +Δδ i(l，r)

[0028] Where: δ i(l，r)0 Let Δδ be the initial active steering angle of the i-th axle wheel. i(l，r) The additional active steering angle of the i-th axle wheel.

[0029] Preferably, the formula for calculating the additional yaw moment is as follows:

[0030]

[0031] in, C is the derivative of the desired yaw rate. i (i = 1, 2, 3, 4, 5) represents the equivalent lateral stiffness of the wheel on the i-th axle, l i (i = 1, 2, 3, 4, 5) are variables with positive and negative values, representing the distance from the center of the wheel on the i-th axle to the center of mass. When the axle is in front of the vehicle's center of mass, l i When the axle is positive and located behind the center of gravity, l i If negative, J zz Let v be the moment of inertia of the entire vehicle about the z-axis of the vehicle body coordinate system. x v is the longitudinal velocity of the vehicle. y Let ω represent the lateral velocity of the vehicle, and k1, k2, and k3 represent the proportional relationships between the steering angles of wheels on axles 2, 4, and 5 and wheels on axle 1, respectively. z Let δ1 be the yaw rate of the vehicle, δ1 be the rotation angle of one axle wheel, and e be the yaw rate of the vehicle. ω The deviation between the actual and desired yaw rate of the vehicle is denoted by λ, where λ is the sliding surface coefficient. Let be the sliding membrane function. Let be a saturation function of the sliding membrane function. This is the convergence coefficient.

[0032] Preferably, the formula for calculating the braking torque distributed to each axle wheel is as follows:

[0033] T bi(l，r) =F bi(l，r) ·R w

[0034] In the formula: F bi(l，r) R represents the braking force of the left and right wheels on the i-th axle. w The rolling radius of a wheel, T bi(l，r) Let be the braking torque of the left and right wheels of the i-th axle.

[0035] Preferably, the decision module includes a coordination controller, which determines the coordinated control commands of ARS and DBTD based on the vehicle's lateral load transfer rate and yaw rate deviation according to predetermined rules.

[0036] Compared with the prior art, the advantages of this invention are as follows:

[0037] (1) The integrated control system for active rear wheel steering and differential braking coordination of multi-axle special vehicles provided by the present invention reduces the amplitude of the yaw rate deviation and the center of gravity sideslip angle of the vehicle controlled by the integrated controller by 46% and 63% respectively compared with the uncontrolled vehicle, and the amplitude of the rear wheel steering angle of the vehicle is reduced by 1.4 degrees compared with the uncontrolled vehicle. The wheels of each axle can brake proportionally according to the predetermined rules, and the path following performance is also better than that of the uncontrolled vehicle.

[0038] (2) The integrated control system for active rear-wheel steering and differential braking coordination of multi-axle special vehicles provided by this invention reduces the amplitude of the yaw rate and the sideslip angle of the vehicle controlled by the integrated controller in low-adhesion steering driving conditions by 47% and 58%, respectively, compared with the uncontrolled vehicle. The amplitude of the rear wheel steering angle controlled by the controller is increased by 1.2 degrees compared with the uncontrolled vehicle, which increases the steering sensitivity of the vehicle when driving at low speeds and on low-adhesion surfaces, and improves the oversteer situation of the vehicle. Attached Figure Description

[0039] Figure 1 This is a diagram of a linear 2-DOF dynamic model in an embodiment of the present invention;

[0040] Figure 2 This is a structural diagram of the active steering controller in an embodiment of the present invention;

[0041] Figure 3 This is a diagram of the Ackermann steering model in an embodiment of the present invention;

[0042] Figure 4 A braking control strategy diagram of the control system provided in an embodiment of the present invention;

[0043] Figure 5 A speed comparison curve of vehicles with integrated control and uncontrolled vehicles when changing lanes at high speeds and high maneuverability.

[0044] Figure 6 A comparison curve of the driving trajectories of vehicles with integrated control and uncontrolled vehicles when changing lanes at high speeds and with high maneuverability.

[0045] Figure 7 A comparison curve of yaw rate between vehicles with integrated control and uncontrolled vehicles when changing lanes at high speeds and with high maneuverability.

[0046] Figure 8 A curve comparing the sideslip angle of the center of gravity of a vehicle with integrated control and an uncontrolled vehicle when changing lanes at high speeds and with high maneuverability.

[0047] Figure 9A comparison curve of the rear wheel steering angle of a vehicle with integrated control and an uncontrolled vehicle when changing lanes at high speeds and with high maneuverability.

[0048] Figure 10 A diagram showing the braking torque of the vehicle wheels when undergoing lane-changing operations at high speeds and high maneuverability.

[0049] Figure 11 A comparison curve of yaw rate of vehicles with integrated control and uncontrolled vehicles when passing through low-adhesion road conditions;

[0050] Figure 12 A curve comparing the center of gravity sideslip angle of a vehicle with integrated control and an uncontrolled vehicle when passing through a low-adhesion road surface.

[0051] Figure 13 A graph showing the rear wheel steering angles of a vehicle with integrated control and an uncontrolled vehicle when traversing a low-adhesion road surface.

[0052] Figure 14 A comparison curve of vehicle trajectories when a vehicle with integrated control and an uncontrolled vehicle travel on a low-adhesion road surface. Detailed Implementation

[0053] Exemplary embodiments of the present disclosure will now be described in more detail with reference to the accompanying drawings. While exemplary embodiments of the present disclosure are shown in the drawings, it should be understood that the present disclosure may be implemented in various forms and should not be limited to the embodiments set forth herein. Rather, these embodiments are provided so that this disclosure will be thorough and complete, and will fully convey the scope of the disclosure to those skilled in the art.

[0054] An integrated control system for coordinated active rear-wheel steering and differential braking in multi-axle special vehicles, specifically comprising:

[0055] The signal input module is used to provide the corresponding signal input quantities to the decision-making module and the allocation module;

[0056] The decision module is used to determine the collaborative control instructions for each sub-controller in the allocation module and send the collaborative control instructions to the allocation module.

[0057] The allocation module is used to receive the coordinated control instructions decided by the decision module, calculate the active steering angle of the active steering axle wheel and the braking torque of each wheel, and then send the calculated control quantities to the execution module.

[0058] The execution module is used to execute the control quantities sent by the allocation module according to the Trucksim dynamics model of the multi-axle special vehicle, and to provide feedback on the vehicle's motion state, so that the integrated control system forms a closed loop.

[0059] The signal input module calculates the desired motion state of the vehicle based on the vehicle's 2-DOF dynamics model, serving as the signal input for the decision-making and allocation modules. The linear 2-DOF dynamics model is a fundamental model reflecting the vehicle's lateral and yaw dynamics characteristics, and it has important applications in vehicle handling stability research. The model assumes that the tire lateral slip characteristics of the corresponding vehicle are in a linear region, the left and right tires have the same lateral slip characteristics, the steering angle is a small angle input, and the body and suspension are assumed to be a rigid system, with the motion state being ideal. To represent the vehicle's desired motion state, as the basis for vehicle lateral stability research, a linear 2-DOF dynamics model corresponding to the vehicle studied in this invention is established as follows: Figure 1 As shown, Figure 1 In the middle, δ i (i = 1, 2, 3, 4, 5) represent the steering angles of wheels on axles one through five. Axle three is a non-steering axle, so δ3 = 0; α i The slip angle of wheels on axles one through five, l i Let l be the distance from the axles of the first to fifth axles to the center of gravity. For ease of description later, let l be defined. i As variables that can be positive or negative, the distance between the axle located in front of the center of mass and the center of mass is defined as positive, and the distance between the axle located behind the center of mass and the center of mass is defined as negative; F yi v represents the lateral force on wheels one through five; v is the actual velocity at the vehicle's center of gravity; v x With v y These are the x-axis and y-axis components of the actual velocity in the vehicle's coordinate system, respectively. ω z β is the yaw rate of the entire vehicle about its center of mass; β is the sideslip angle around the center of mass; ΔM z To add yaw moment.

[0060] Based on Newton's second law and Euler's second law, Figure 1 The vehicle model shown has the following lateral dynamics equations and moment balance equations about its center of mass:

[0061]

[0062] Where: m is the total vehicle mass, J zz Let l be the moment of inertia of the entire vehicle about the z-axis of the vehicle body coordinate system. i F represents the distance from the axles of the first to fifth axles to the center of gravity. yi The lateral force of wheels on axles one through five; v x With v y Let ω be the x-axis and y-axis components of the actual velocity in the vehicle coordinate system. z Let δ be the yaw rate of the entire vehicle about its center of mass. i (i = 1, 2, 3, 4, 5) represents the steering angles of wheels on axles one through five.

[0063] Based on the small-angle assumption in the simplified model, we have: β≈tanβ=v y / v x cosδ i ≈1, when the vehicle is in steady-state steering, the longitudinal speed is constant, therefore Therefore, equation (1) can be simplified to:

[0064]

[0065] In the formula: Let ω be the derivative of the vehicle's sideslip angle. z The vehicle's yaw rate is... Let v be the derivative of the vehicle's yaw rate, m be the total vehicle mass, and v be the derivative of the vehicle's yaw rate. x J is the longitudinal velocity of the vehicle. zz Let F be the moment of inertia of the entire vehicle about the z-axis of the vehicle body coordinate system. yi Let l be the lateral force of the wheel on the i-th axle. i Let l be a variable with positive and negative values, representing the distance from the center of the i-th axle to the center of mass. When the axle is in front of the center of mass, l i When the axle is positive and located behind the center of gravity, l i It is negative.

[0066] Based on the assumption that the tire's lateral slip characteristics are in the linear region, the lateral forces of each axle wheel are:

[0067]

[0068] In the formula: C i (i = 1, 2, 3, 4, 5) represents the equivalent lateral stiffness of the wheels on axles one through five, which is equal to the sum of the lateral stiffness of the left and right tires, i.e., C. i =C il +C ir α i (i = 1, 2, 3, 4, 5) represents the sideslip angle of the wheel on the i-th axle, v x v is the longitudinal velocity of the vehicle. y l is the lateral velocity of the vehicle. i ω is the distance from the center of the i-th axle to the centroid. z Let δ be the yaw rate of the vehicle. i Let β be the steering angle of the i-th axle wheel, and β be the sideslip angle of the vehicle's center of gravity.

[0069] Substituting equation (3) into equation (2), we obtain the state-space expression for the linear 2-DOF model of the vehicle as follows:

[0070]

[0071] In the formula: Let be the derivative of the vehicle's sideslip angle. Let C be the derivative of the vehicle's yaw rate, m be the total vehicle mass, and C be the derivative of the vehicle's yaw rate. i (i = 1, 2, 3, 4, 5) represents the equivalent lateral stiffness of the wheels on axles one through five, which is equal to the sum of the lateral stiffness of the left and right tires, i.e., C. i =C il +C ir , l i v is the distance from the center of the i-th axle to the center of mass. x J is the longitudinal velocity of the vehicle. zz Let δ be the moment of inertia of the entire vehicle about the z-axis of the vehicle body coordinate system. i (i = 1, 2, 3, 4, 5) represents the steering angle of the wheel on the i-th axle.

[0072] Based on the Ackermann steering principle, the steering angles of the wheels on axles 2, 4, and 5 of the monorail model are related to the steering angle of the wheel on axle 1 as follows:

[0073]

[0074] By still making the assumption from a smaller angle, the above equation can be simplified to:

[0075]

[0076] in:

[0077] Substituting equation (6) into equation (4) yields

[0078]

[0079] By performing a Laplace transform on both sides of equation (7) and setting the initial condition to 0, we obtain the sideslip angle β and the yaw rate ω. z The transfer function for the wheel rotation angle δ1 of one axle is:

[0080]

[0081] From the above equation, the steady-state gains of the yaw rate and the sideslip angle of the center of mass with respect to the rotation angle of the first axle wheel are respectively:

[0082]

[0083]

[0084] In the formula:

[0085]

[0086]

[0087] a=C1+k1C2+k2C4+k3C5

[0088] b = l1C1 + l2k1C2 + l4k2C4 + l5k3C5

[0089] The sub-controllers of the distribution module include an active steering controller and a differential braking torque controller. The differential braking torque controller includes a yaw rate sliding film controller and a torque distribution controller. The active steering controller takes the zero center of gravity sideslip angle as the control target and adaptively calculates the active steering angle of the vehicle's rear wheels based on the Ackermann steering model and the PID feedback controller. The differential braking torque controller, based on the yaw rate deviation of the vehicle, adaptively calculates the additional yaw torque required to maintain the yaw stability of the vehicle based on the yaw rate sliding film controller, and then distributes the braking torque of each axle wheel based on the torque distribution controller.

[0090] The main function of the active steering controller is to maintain the vehicle's roll stability and trajectory tracking capability. When the vehicle's trajectory deviates due to insufficient road adhesion coefficient or saturated tire lateral force, preventing it from following the desired path, or when severe body roll leads to sideslip or even rollover, the controller uses the active steering lever to control the wheels to rotate at a certain angle, ensuring the vehicle maintains good trajectory tracking and roll stability. Good trajectory tracking and roll stability are characterized by a vehicle's center-of-gravity sideslip angle approaching zero; therefore, zero center-of-gravity sideslip angle is taken as the controller's control objective. To prevent perturbations of the center-of-gravity sideslip angle parameter caused by factors such as changes in road surface coefficient, tire stiffness, and crosswind interference, and considering both the controller's control effect and the computational speed of the integrated control algorithm, this invention designs an active steering controller based on the PID control principle to track and control the vehicle's center-of-gravity sideslip angle. The controller's structure is as follows... Figure 2 As shown. Where: β d For the desired centroid sideslip angle, e β e represents the deviation of the centroid sideslip angle. β =β d -β;δ 4(l，r)0 δ 5(l，r)0 These are the initial active steering angles of the left and right wheels of the four-axle and five-axle wheels, respectively, Δδ 4(l，r) ,Δδ 5(l，r) This is the feedback input for the active steering angle of the four-axle and five-axle wheels.

[0091] To solve for the initial active steering angle of the four-axle and five-axle wheels, a system is established as follows: Figure 3 The Ackermann steering model is shown. According to the Ackermann steering principle, when a vehicle is in a stable steering state, the perpendicular lines passing through the rotation centers of each wheel and perpendicular to the tire neutral plane should intersect at the same instantaneous steering center. Since the three axles of this type of special vehicle are non-steering axles, the instantaneous steering center o' is located on the extension line of the centerline of the three axles. Figure 3 In the diagram, B represents the vehicle's track width, and R...d δ is the instantaneous turning radius. i(l，r) This refers to the steering angle of the left and right wheels on each axle.

[0092] Based on geometric relationships, the same-side wheel steering angles of vehicles conforming to the Ackermann steering model satisfy the following relationship:

[0093]

[0094] Where: δ i(l，r) l represents the steering angle of the left and right wheels on each axle. i This refers to the distance from the axle of the first to fifth axle to the center of gravity.

[0095] The steering angles of the left and right wheels on the same axle satisfy the following relationship:

[0096]

[0097] Let the steering wheel angle be δ sw The transmission ratio of the drive train from the steering wheel to the first axle wheel is i. sw Then the turning angle of the left wheel of the axle can be simplified as follows:

[0098]

[0099] From equations (11), (12), and (13), we can obtain

[0100]

[0101] The PID feedback controller is used to track and control the vehicle's sideslip angle, ensuring it always approaches the desired sideslip angle. Its feedback output is the additional active steering angle of the four-axle and five-axle wheels. The formula for calculating the additional active steering angle of the four-axle and five-axle wheels is as follows:

[0102]

[0103] In the formula: P1, I1, and D1 are the proportional coefficient, integral coefficient, and derivative coefficient of the PID controller, respectively.

[0104] Based on the above analysis, the total input of the active steering angle of the four-axle and five-axle wheels is:

[0105] δ i(l，r) =δ i(l，r)0 +Δδ i(l，r) (16)

[0106] Where: δ i(l，r) Δδ is the total input of the active steering angle of the four-axle and five-axle wheels. i(l，r) This is the additional active steering angle for four-axle and five-axle wheels.

[0107] Since the power source of the vehicle studied in this invention is still the traditional single drive form that relies on the engine to provide total power, which is then transmitted to each axle and wheel by the power transmission system, it is difficult to achieve the longitudinal force difference between the left and right wheels by distributing the driving torque on each wheel. Therefore, this invention uses differential braking to generate a longitudinal force difference between the left and right wheels, thereby forming an additional yaw moment to control the yaw stability of the vehicle. Differential braking control not only has a simple and feasible actuator, but also, in some dangerous situations such as sideslip and lateral deviation caused by driver error, can effectively alleviate the driver's panic by reducing the vehicle speed, avoiding serious consequences due to failure to correct the driving error in time. The differential braking torque controller consists of a yaw rate controller and a torque distributor, which are designed as follows:

[0108] A yaw rate diaphragm controller is used to track and control the yaw rate of a vehicle, maintaining the actual yaw rate close to the ideal yaw rate, thereby ensuring the vehicle's yaw stability. The desired yaw rate of the vehicle is equal to the steady-state gain of the yaw rate. The product of the yaw rate and the rotation angle δ1 of one axle wheel is used to calculate the vehicle's expected yaw rate, as follows:

[0109]

[0110] In the formula: The steady-state gain of the yaw rate is given by δ1, where δ1 is the rotation angle of one axle wheel, and v is the yaw rate. x L represents the longitudinal velocity of the vehicle. e K is the equivalent wheelbase, and K is the stability factor.

[0111] The ideal yaw rate of a vehicle, derived from the constraint of the road adhesion coefficient on the vehicle's lateral acceleration, should satisfy the following constraints.

[0112]

[0113] In the formula: μ is the road adhesion coefficient, g is the acceleration due to gravity, and v x Let ω be the longitudinal velocity of the vehicle. zd Let be the vehicle's desired yaw rate.

[0114] Therefore, the vehicle's desired yaw rate is:

[0115]

[0116] Let the additional yaw moment exerted on the vehicle body by the control system be ΔM. z Under the action of the additional yaw moment, the new moment equilibrium equation of the vehicle about its center of mass is:

[0117]

[0118] In the formula: ξz These are uncertainties caused by model simplification, parameter perturbations, etc. The meanings of other parameters are the same as described above.

[0119] Combining the small angle assumption and equations (7) and (20), we can obtain the reciprocal of the vehicle's yaw rate at this time as follows:

[0120]

[0121] In the formula:

[0122] The tracking error of yaw rate is defined as:

[0123] e ω =ω z -ω zd (twenty two)

[0124] Take the sliding diaphragm function as the sum of the tracking error and the integral of the tracking error, where the yaw rate is the yaw function.

[0125]

[0126] In the formula, Let λ be the deviation between the actual yaw rate and the expected yaw rate of the vehicle, and let λ be the coefficient of the slick surface. Combining equation (21), we can obtain the derivative of the above equation.

[0127]

[0128] In the formula, v y Let ω be the lateral velocity of the vehicle. z Let ΔM be the yaw rate of the vehicle. z To add yaw moment, ξ z For the disturbance term of the additional yaw moment, The derivative of the desired yaw rate, q represents the deviation between the actual yaw rate and the desired yaw rate of the vehicle. i The meaning of (i = 1, 2, 3, 4) is the same as described above.

[0129] The chattering problem of the sliding function is eliminated by using a saturation function whose rate of convergence is exponential, i.e., taking:

[0130]

[0131] In the formula: Let be the sliding membrane function. Let be a saturation function of the sliding membrane function. This is the convergence coefficient.

[0132] Substituting equation (25) into equation (24), the additional yaw torque output of the sliding membrane controller is:

[0133]

[0134] In the formula: δ1 is the rotation angle of one axle wheel, and the meanings of the other parameters are the same as those described above.

[0135] During vehicle steering, the braking force of the inner rear wheel generates the most effective inward yaw moment, while the braking force of the outer front wheel generates the most effective outward yaw moment. Based on this, a rule-based braking torque distribution strategy was designed, specifically: when the required yaw moment is towards the inward side of the steering, the inner rear wheel is braked first; when the required yaw moment is towards the outward side of the steering, the outer front wheel is braked first.

[0136] Referring to this allocation rule, the braking control strategy for the vehicle of this invention is formulated as follows: First, the vehicle's steering state is determined based on the yaw rate; then, the direction of the required yaw moment is determined; finally, the yaw moment is achieved according to the braking force distribution rule. The vehicle has four steering states. The yaw moment required for each steering state is obtained by braking the wheel that produces the greatest benefit as the primary braking wheel, with the remaining wheels on the same side as secondary braking wheels. The secondary braking wheels are used to share the braking pressure of the primary braking wheels, preventing the yaw moment requirement from being insufficient due to the primary braking wheel's output saturation. From primary to secondary, the braking force of each wheel is distributed in a ratio of 50%:25%:15%:7%:3%, as follows: Figure 4 As shown.

[0137] The counterclockwise direction is defined as the positive direction for both yaw rate and yaw moment. The specific braking control rules are shown in Table 1 below.

[0138] Table 1 Braking Control Rules

[0139]

[0140] The braking control rules are explained as follows:

[0141] 1) When the vehicle's desired yaw rate ω zd Greater than 0, and the absolute value of the vehicle's actual yaw rate |ω z |less than the absolute value of the desired yaw rate|ω zd At this time, the vehicle is in a state of left turn and understeering. The direction of the required additional yaw moment is counterclockwise. To obtain this additional yaw moment, the braking strategy is to brake the left wheels, with the left rear wheel as the main brake. Figure 4 As shown in (a).

[0142] 2) When the vehicle's desired yaw rate is greater than 0, and the absolute value of the vehicle's actual yaw rate is greater than the absolute value of the desired yaw rate, the vehicle is in a left-turning and oversteering motion state. At this time, the direction of the additional yaw moment required by the vehicle is clockwise, and the braking strategy involves braking the right-side wheels, with the right front wheel as the main braking wheel. Figure 4 As shown in (b).

[0143] 3) When the vehicle's desired yaw rate is less than 0, and the absolute value of the vehicle's actual yaw rate is less than the absolute value of the desired yaw rate, the vehicle is in a right-turning and understeering motion state. At this time, the direction of the additional yaw moment required by the vehicle is clockwise, and the braking strategy involves braking the right-side wheels, with the right rear wheel as the main braking wheel. Figure 4 As shown in (c).

[0144] 4) When the vehicle's desired yaw rate is less than 0, and the absolute value of the vehicle's actual yaw rate is greater than the absolute value of the desired yaw rate, the vehicle is in a right turn and oversteering motion state. At this time, the direction of the additional yaw moment required by the vehicle is counterclockwise, and the braking strategy involves braking the left wheels, with the main braking wheel being the left front wheel. Figure 4 As shown in (d).

[0145] Let the braking force of each wheel be F. bi(l，r) The resultant force F of braking force b The braking force of each wheel is calculated based on the additional yaw moment output by the diaphragm controller as follows:

[0146] For vehicles in state (a):

[0147]

[0148] For vehicles in state (b)

[0149]

[0150] For vehicles in state (c)

[0151]

[0152] For vehicles in state (d)

[0153]

[0154] Where: δ is the width of the B axle. i(l，r) The rotation angle of the wheels on the left and right sides of the i-axis.

[0155] Let the rolling radius of the wheel be R. w Based on the calculated braking forces of each wheel, the braking torque acting on each wheel can be obtained as follows:

[0156] T bi(l，r) =F bi(l，r) ·R w (31)

[0157] In the formula: F bi(l，r) T represents the braking force of the left and right wheels of the i-th axle. bi(l，r) Let be the braking torque of the left and right wheels of the i-th axle.

[0158] Two issues must be considered when integrating multiple control methods in a vehicle stability integrated control algorithm: First, when the vehicle is at risk of instability, how to determine which control method(s) to use for stability control based on the vehicle's motion state. Second, when multiple control methods are used simultaneously to control vehicle stability, how to allocate the control range and control ratio of each sub-control system.

[0159] An excellent integrated control system should not only maximize the control effect of each sub-control system, but also ensure that they do not interfere with each other when working collaboratively, compensating for their respective shortcomings and limitations, and expanding the control range of the overall system. Therefore, it is necessary to analyze and study the timing of operation of various control methods and the coordinated allocation of these methods.

[0160] The stability integrated control system of this invention controls vehicle stability by integrating active steering and differential braking. The active steering controller is mainly used to control vehicle roll stability, while the differential braking controller is mainly used to control vehicle yaw stability, compensating for insufficient tire lateral force and limitations of active steering control capabilities under high-speed conditions. Vehicle roll stability can be characterized by changes in the lateral load transfer rate. Therefore, the decision module of this invention includes a coordination controller, which determines the coordinated control commands for the active steering controller and the differential braking controller based on the vehicle's lateral load transfer rate and yaw rate deviation according to predetermined rules.

[0161] This invention uses the lateral load transfer ratio (LTR) as the criterion for vehicle roll stability control, and the vehicle's yaw rate and its difference from the ideal yaw rate as the criterion for vehicle yaw stability control. LTR is defined as follows. d To determine the lateral load transfer rate threshold at which a vehicle is at risk of rollover, a coordinated control strategy as shown in Table 2 can be formulated based on rules.

[0162] Table 2 Coordination and Control Strategies

[0163]

[0164] In the table, LTR represents the lateral load transfer rate of the vehicle. d To determine the threshold for lateral load transfer rate when a vehicle is at risk of rollover, |eω | The absolute value of the yaw rate difference, where e1 is a positive threshold for the yaw rate difference.

[0165] The specific details of the coordination and control strategy are as follows:

[0166] 1) When LTR < LTR d ,|e ω When | < e1, the vehicle is determined to be in a stable state, and the control system does not perform stability control on the vehicle.

[0167] 2) When LTR < LTR d ,|e ω When |≥e1, it is determined that the vehicle is at risk of yaw instability, and the control system only uses differential braking to control the vehicle's stability.

[0168] 3) When LTR ≥ LTR d ,|e ω When | < e1, it is determined that the vehicle is at risk of tilting and instability, and the control system only performs stability control on the vehicle through active steering.

[0169] 4) When LTR ≥ LTR d ,|e ω |≥e1, and sgn(a y ) = sgn(e ω When it is determined that the vehicle is at risk of both roll instability and yaw instability, and the direction of the yaw moment required for both is the same, the control system performs stability control on the vehicle based on active steering and differential braking.

[0170] (5) When LTR ≥ LTR d ,|e ω |≥e1, and sgn(a y )≠sgn(e ω When a vehicle is deemed to have both roll and yaw risks, but the required yaw moments are in opposite directions, the control system first uses active steering, then differential braking, to control vehicle stability. This is because rollover accidents often result in more severe damage than yaw accidents.

[0171] When two control methods work together, in order to eliminate control system chatter caused by frequent switching between the two control methods, the output weights of the two control methods can be defined as follows by defining a positive threshold e2 of the yaw rate difference:

[0172] Define the distribution ratio function of the rear wheel active steering angle with respect to the total control amount as follows:

[0173]

[0174] In the formula: ρ(|e w |) is about |e w | is a monotonically decreasing function.

[0175] The distribution ratio function of the additional yaw moment with respect to the total control amount is:

[0176]

[0177] At this time, the control quantities output by each sub-control system are:

[0178]

[0179] I. Analysis and Verification

[0180] 1.1 Working Condition Design

[0181] To verify the control effect of the integrated controller on vehicle stability, a Simulink model of the integrated controller was built based on Matlab, and a co-simulation platform was built with the Trucksim vehicle dynamics model.

[0182] 1.2 Dual-line shifting condition test

[0183] To verify the effectiveness of the integrated controller in controlling the stability of vehicles during high-maneuverability lane changes on high-adhesion road surfaces, a simulation test was conducted on a double-lane road surface with a road adhesion coefficient μ = 0.85. The vehicle speed was set to 100 km / h. The operating states of the vehicle with integrated control and the vehicle without control were compared when passing through this condition. The comparison curves are shown below. Figures 5-10 As shown. By Figure 5 and Figure 6 It can be seen that the overshoot of the lateral displacement of the vehicle with integrated control when passing the last curve of the double lane change road is 0.126m smaller than that of the uncontrolled vehicle. The time for the lateral displacement of the vehicle with integrated control to reach steady state is 7.6s, while that of the uncontrolled vehicle is 9.5s.

[0184] Figure 7 and with Figure 8 The figures show a comparison of the vehicle's yaw rate and sideslip angle. The data indicates that the amplitudes of the yaw rate deviation and sideslip angle of the vehicle with integrated control are reduced by 46% and 63%, respectively, compared to the uncontrolled vehicle. This demonstrates that the designed integrated control system significantly improves the vehicle's driving stability at high speeds and on surfaces with high adhesion coefficients.

[0185] Figure 9The graph shows a comparison of the steering angles (rear wheel angles) of the four-axle and five-axle vehicles. It can be seen from the graph that the extreme values ​​of the rear wheel angles of the vehicle with integrated control are very small in both positive and negative directions, and the peak values ​​are also more uniform when turning left and right. This indicates that the vehicle runs more smoothly under integrated control, which corresponds to the fact that the amplitude of the yaw rate deviation and the center of gravity sideslip angle of the vehicle is smaller than that of the uncontrolled vehicle.

[0186] Figure 10 A diagram showing the braking torque for integrated control of vehicle wheels. (From...) Figure 10 It can be seen that when the yaw rate of the vehicle exceeds the set threshold, the brakes of the wheels on the left and right sides of the vehicle can brake the wheels proportionally according to the established rules.

[0187] pass Figures 5 to 10 The results show that, for high-attachment, high-speed lane-changing driving conditions, the integrated control system can effectively improve the driving stability of multi-axle special vehicles.

[0188] 1.3 Low-adhesion pavement test

[0189] Simulation tests were conducted on an S-shaped road surface with an adhesion coefficient μ = 0.4, and the simulated vehicle speed was set to 60 km / h. The comparison of vehicle motion states was as follows: Figures 11-14 As shown.

[0190] Figure 11 This is a comparison chart of yaw angular velocities. Figure 12 A comparison diagram of the centroid sideslip angle, through Figure 11 and Figure 12 Data shows that the integrated control system provided in this embodiment of the invention reduces the yaw rate amplitude and the center-of-gravity sideslip angle amplitude of the vehicle by 47% and 58%, respectively, compared to the uncontrolled vehicle.

[0191] Figure 13 This is a comparison chart of the rear wheel steering angles of the vehicles. Figure 13 Data shows that the amplitude of the wheel steering angle of the four-axle and five-axle vehicles with integrated control is increased by deg and deg, respectively, compared with the uncontrolled vehicles. The increased rear wheel steering angle widens the available margin of tire lateral force.

[0192] Figure 14 A comparison chart of vehicle driving trajectories, by Figure 14 As can be seen from the enlarged view, both the integrated control vehicle and the uncontrolled vehicle exhibit slight oversteering, but the oversteering of the integrated control vehicle is less than that of the uncontrolled vehicle.

[0193] pass Figures 11 to 14 The results show that the integrated control system can improve the steering sensitivity of the vehicle at low speeds and on low-friction surfaces.

[0194] In summary, the stability integrated control system of active rear wheel steering control and differential braking control provided in this embodiment of the invention can effectively improve the driving stability of multi-axle special vehicles and improve the steering sensitivity of vehicles on low-speed, low-friction surfaces.

[0195] Although embodiments of the invention have been shown and described, those skilled in the art will understand that various changes, modifications, substitutions and alterations can be made to these embodiments without departing from the principles and spirit of the invention, the scope of which is defined by the claims and their equivalents.

Claims

1. An integrated control system for coordinated active rear-wheel steering and differential braking in a multi-axle special vehicle, characterized in that, include: The signal input module is used to provide the corresponding signal input quantities to the decision-making module and the allocation module; The decision module is used to determine the collaborative control instructions for each sub-controller in the allocation module and send the collaborative control instructions to the allocation module. The allocation module is used to receive the coordinated control command decided by the decision module, calculate the active steering angle of the active steering axle wheel and the braking torque of each wheel, and then send the calculated control quantities to the execution module. The sub-controller of the distribution module includes an active steering controller and a differential braking torque controller. The differential braking torque controller includes a yaw rate sliding film controller and a torque distribution controller. The active steering controller calculates the active steering angle of the vehicle's wheels based on the Ackermann steering model and the PID feedback controller. The yaw rate sliding film controller calculates the additional yaw torque required to maintain the yaw stability of the vehicle. The torque distribution controller calculates the braking torque distributed to each axle wheel and then sends the calculated control quantities to the execution module. The feedback output of the PID feedback controller is the additional active steering angle of the four-axle and five-axle wheels. The calculation formula for the additional active steering angle of the four-axle and five-axle wheels is as follows: ； In the formula: These are the proportional coefficient, integral coefficient, and derivative coefficient of the PID controller. This represents the deviation of the centroid sideslip angle; The total input of the active steering angle of the four-axle and five-axle wheels is: ； In the formula: Let be the initial active steering angle of the i-th axle wheel. For the additional active steering angle of the i-th axle wheel; The formula for calculating the additional yaw moment is as follows: ； in, , , , , , , The derivative of the desired yaw rate, For the first The equivalent lateral stiffness of the axle wheel. Let be a variable with positive and negative signs, representing the th . The distance from the center of the wheel to the center of gravity of the axle, when the axle is in front of the vehicle's center of gravity. When the axle is positive and located behind the center of gravity, Negative, For the whole vehicle around the body coordinate system Moment of inertia of the shaft, Let be the longitudinal speed of the vehicle. Let be the lateral speed of the vehicle. , , The figures show the proportional relationships between the steering angles of wheels on axles 2, 4, and 5 and wheels on axle 1, respectively. Let yaw rate be the vehicle's angular velocity. The turning angle of a wheel on one axle. The deviation between the vehicle's actual yaw rate and the desired yaw rate. The sliding surface coefficient, Let be the sliding membrane function. Let be a saturation function of the sliding membrane function. , The approximation coefficient; The decision module includes a coordination controller, which determines the coordinated control commands of the active steering controller and the differential braking torque controller based on the vehicle's lateral load transfer rate and yaw rate deviation according to predetermined rules. The execution module is used to execute the control quantities sent by the allocation module according to the Trucksim dynamics model of the multi-axle special vehicle, and to provide feedback on the vehicle's motion state, so that the integrated control system forms a closed loop.

2. The integrated control system for coordinated active rear-wheel steering and differential braking of multi-axle special vehicles as described in claim 1, characterized in that, The signal input module calculates the vehicle's desired motion state based on the vehicle's 2-DOF vehicle dynamics model, and uses this as the signal input quantity for the decision module and the allocation module.

3. The integrated control system for coordinated active rear-wheel steering and differential braking of multi-axle special vehicles as described in claim 2, characterized in that, The 2-DOF vehicle dynamics model is as follows: ； In the formula: The derivative of the centroid sideslip angle. The derivative of the yaw rate. For the overall vehicle quality, For the whole vehicle around the body coordinate system Moment of inertia of the shaft, Let be a variable with positive and negative signs, representing the th . The distance from the center of the wheel to the center of gravity of the axle, when the axle is in front of the vehicle's center of gravity. When the axle is positive and located behind the center of gravity, Negative, For the first The equivalent lateral stiffness of the wheel axle is equal to the sum of the lateral stiffness of the left and right tires, i.e. ; Let be the longitudinal speed of the vehicle. For the first The steering angle of the axle wheel.

4. The integrated control system for coordinated active rear-wheel steering and differential braking of multi-axle special vehicles as described in claim 1, characterized in that, The Ackermann steering model is as follows: ； In the formula: Let be the steering angle of the left and right wheels on the i-th axis. Let be the distance from the center of the i-th axle to the center of mass; The initial steering angles of the four-axle and five-axle wheels are: ； In the formula: ( ( ) represents the initial active steering angle of the left and right wheels of the four-axle and five-axle systems. The width of the axle. For steering wheel angle, This is the transmission ratio between the steering wheel and one of the wheels on the axle.

5. The integrated control system for coordinated active rear-wheel steering and differential braking of multi-axle special vehicles as described in claim 1, characterized in that, The formula for calculating the braking torque distributed to each axle wheel is as follows: ； In the formula: This represents the braking force of the left and right wheels on the i-th axis. The rolling radius of the wheel, Let be the braking torque of the left and right wheels of the i-th axle.

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