Multifunctional dual-motor coupling active steering system of commercial vehicle and control method of multifunctional dual-motor coupling active steering system

Abstract

The invention discloses a multifunctional dual-motor coupling active steering system of a commercial vehicle and a control method of the multifunctional dual-motor coupling active steering system. The control method comprises the steps of collecting a turning angle signal and a torque signal of a vehicle steering wheel; the stability margin of the vehicle is calculated according to the collected signals, and the state of the electromagnetic actuator is determined; calculating the transmission ratio required under the limit working condition; calculating a rotation angle value required by the auxiliary motor, and sending a control signal to the auxiliary motor; a power-assisted torque value required by the main motor is calculated, and a control signal is sent to the main motor; calculating torque distribution values of the auxiliary motor and the main motor under a common working condition, and respectively sending the torque distribution values to the auxiliary motor and the main motor; and outputting a corresponding torque or corner to finish vehicle steering. Under a common working condition, through dual-motor power-assisted coupling, the steering portability at a low speed can be further improved; under the limiting working condition, power is independently assisted through the main motor, the auxiliary motor achieves active compensation on the steering angle of a driver through the double-row planetary gear module, the stepless variable transmission ratio can be achieved, and active rollover prevention control is achieved.

Description

Technical Field

[0001] The present invention belongs to the technical field of vehicle steering, and specifically refers to a multi-functional dual-motor coupled active steering system for commercial vehicles and its control method. Background Art

[0002] The steering system is one of the key components of an automobile. It not only ensures that the automobile can steer according to the driver's will, but also is closely related to the comfort, handling stability and active safety of the automobile. At present, the electro-hydraulic power steering system (EHPS) and the electric power steering system (EPS) have been widely applied to actual vehicles. EHPS uses a hydraulic system to provide power, and traditional hydraulic components are indispensable in its power steering process, so the structure is not simple enough. While EPS completely abandons hydraulics, the components and pipelines eliminate the risk of hydraulic oil leakage, the structure is simpler, the maintenance and installation are more convenient, and at the same time, without the corrosive hydraulic liquid, it is more environmentally friendly. EPS is driven and controlled by an electronic control unit (ECU) to drive the actuator motor to generate steering torque, which has the advantages of no hydraulic delay, faster response speed, no need to change hydraulic oil, and lower cost at the same time. EPS is regarded by many as a superior alternative to EHPS, making it an important direction for the future development of steering systems.

[0003] However, there are problems with the steering systems of commercial vehicles in terms of lightness and safety. On the one hand, commercial vehicles require greater torque, and a single motor cannot fully meet the torque requirements. Currently, there are certain solutions to this problem. For example, a dual-motor steering system for a driving simulator in intelligent vehicle human-machine co-driving testing, disclosed in Chinese Patent Application No. CN202410918416.5, provides greater assist torque through dual motors. On the other hand, commercial vehicles have higher requirements for safety. Traditional power steering systems can only apply variable steering assistance by means of a hydraulic power system or an electric power system according to the driver's steering operation to reduce the driver's steering burden, and cannot perform variable transmission ratio control for various situations encountered by the vehicle during driving, thus failing to play an important role in the vehicle's handling stability. Currently, there are also certain solutions to this problem. For example, an adaptive active front-wheel steering transmission system for commercial vehicles, disclosed in Chinese Patent Application No. CN201911282013.1, controls the rotation angle of the common planet carrier through a worm and worm gear mechanism to achieve variable transmission ratio. Currently, there is also an active steering system with three motors. For example, a multi-mode dual-redundancy active steering system assembly and control method, disclosed in Chinese Patent Application No. CN202210530849.4, realizes power steering and improves system safety through the coordinated operation of three motors. A simple dual-motor steering system can meet the steering power requirements, but it cannot achieve variable transmission ratio and has insufficient operating stability. Although the steering motor uses a worm and worm gear mechanism to control the rotation angle of the common planet carrier, it can achieve variable transmission ratio, but the steering motor only plays a role in angle control and cannot provide power steering, lacking lightness. The coordinated operation of three motors realizes power steering. Although it can achieve greater power steering and variable transmission ratio, the three-motor steering system is relatively redundant, increasing safety hazards and requiring more complex control methods.

[0004] In summary, with the development of vehicle electrification and intelligence, the electric power steering system has gradually become the mainstream development trend of future vehicle steering systems. However, traditional commercial vehicle power steering systems have problems with lightness and safety. A single motor cannot fully meet the torque requirements, and traditional dual-motor steering systems cannot perform variable transmission ratio control for various situations encountered by the vehicle during driving, thus failing to play an important role in the vehicle's handling stability. Although there is currently an active steering system with three motors, adding motors will bring safety hazards and increase the complexity of the control method. Therefore, there is an urgent need for a multi-functional dual-motor coupled active steering system and its control strategy that can meet power requirements, be safe and reliable, and have simpler control. Summary of the Invention

[0005] Aiming at the deficiencies of the above-mentioned existing technologies, the purpose of the present invention is to provide a multi-functional dual-motor coupled active steering system for commercial vehicles and its control method, so as to solve the problems of insufficient power, lack of safety and reliability, and complex control system in the existing active steering systems. The present invention deeply integrates the double-row planetary gear and the electromagnetic actuator to realize a multi-functional dual-motor coupled active steering system. That is, in normal working conditions, the electromagnetic actuator is locked, and the auxiliary motor superimposes the torque on the steering wheel torque through the double-row planetary gear, realizing the coupling of the torque outputs of the main and auxiliary motors in normal working conditions. In extreme working conditions, the electromagnetic actuator is disengaged, and the auxiliary motor and the main motor respectively control the steering angle and torque of the system. That is, the auxiliary motor realizes the active compensation for the driver's steering angle through the double-row planetary gear mechanism, and cooperates with the assist control of the main motor to realize the active rollover prevention control in extreme working conditions.

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

[0007] A multi-functional dual-motor coupled active steering system for commercial vehicles of the present invention includes: a steering wheel module, a double-row planetary gear module, a steering gear module and a control module;

[0008] The steering wheel module includes: a steering wheel, a steering column and a steering column housing. The steering wheel is fixedly connected to the input end of the steering column, and the steering column is arranged inside the steering column housing;

[0009] The double-row planetary gear module includes: an upper-row planetary gear, a lower-row planetary gear, a common planetary carrier, an auxiliary motor and an electromagnetic actuator;

[0010] The input shaft of the upper-row planetary gear is fixedly connected to the output end of the steering column;

[0011] The upper-row planetary gear includes: an upper-row tooth ring, upper-row planetary gears and an upper-row sun gear. The upper-row planetary gears are installed on the common planetary carrier and are arranged between the upper-row tooth ring and the upper-row sun gear; the input shaft of the upper-row planetary gear is fixedly connected to the upper-row sun gear;

[0012] The lower-row planetary gear includes: a lower-row tooth ring, lower-row planetary gears and a lower-row sun gear. The lower-row planetary gears are installed on the common planetary carrier and are arranged between the lower-row tooth ring and the lower-row sun gear; the output shaft of the lower-row planetary gear is fixedly connected to the lower-row sun gear;

[0013] The auxiliary motor includes: a force / displacement auxiliary motor, an upper-row gear of the auxiliary motor and a lower-row gear of the auxiliary motor; the upper-row gear of the auxiliary motor and the lower-row gear of the auxiliary motor are installed on the output shaft of the force / displacement auxiliary motor; the upper-row gear of the auxiliary motor meshes with the outer side of the upper-row tooth ring; the lower-row gear of the auxiliary motor meshes with the outer side of the lower-row tooth ring;

[0014] The electromagnetic actuator includes a coil and a fixed magnetic core. The coil is disposed outside the fixed magnetic core. The current passing through the coil generates a magnetic field that interacts with the fixed magnetic core, and an electromagnetic force is generated on the fixed magnetic core to achieve disengagement and locking with the common planet carrier.

[0015] The steering gear module includes: a planetary roller screw steering gear, a main motor, a steering arm, a steering drag link, a trapezoidal arm, a steering cross rod, a steering knuckle, a wheel hub and a wheel.

[0016] The planetary roller screw steering gear includes: a screw, planetary rollers, a cage, an internal gear ring, a nut and a tooth sector. The upper end of the screw is fixedly connected to the output shaft of the lower row of planetary gears. The planetary rollers are mounted on the cage and are disposed between the screw and the nut. The internal gear ring is mounted at the upper and lower ends of the planetary rollers and is disposed between the cage and the nut. The nut meshes with the tooth sector. The tooth sector is fixedly connected to the steering arm.

[0017] The steering arm is hinged to the steering drag link.

[0018] The steering drag link is hinged to the trapezoidal arm.

[0019] The trapezoidal arm is hinged to the steering cross rod.

[0020] The steering cross rod is hinged to the steering knuckle.

[0021] The steering knuckle is fixedly connected to the wheel hub.

[0022] The wheel is fixedly connected to the wheel hub.

[0023] The control module includes: an electronic control unit, a steering angle sensor and a torque sensor. The steering angle sensor and the torque sensor are installed between the steering column and the steering column housing and are electrically connected to the electronic control unit. The electronic control unit is also electrically connected to the electromagnetic actuator, the auxiliary motor and the main motor.

[0024] The present invention also provides a control method for a multi-functional dual-motor coupled active steering system of a commercial vehicle. Based on the above system, it includes the following steps:

[0025] 1) Collect the steering angle signal and torque signal of the vehicle steering wheel and transmit the signals to the electronic control unit.

[0026] 2) The electronic control unit calculates the vehicle stability margin according to the collected steering angle signal and torque signal, and determines the state of the electromagnetic actuator according to the vehicle stability margin. If the state of the electromagnetic actuator is disengaged, go to step 3); if the state of the electromagnetic actuator is locked, go to step 4).

[0027] 3) The electronic control unit calculates the required transmission ratio under extreme operating conditions; calculates the required rotation angle value of the auxiliary motor, and sends a control signal to the auxiliary motor; calculates the required assist torque value of the main motor, and sends a control signal to the main motor, then enters step 5);

[0028] 4) The electronic control unit calculates the torque distribution values of the auxiliary motor and the main motor under normal operating conditions, and sends them to the auxiliary motor and the main motor respectively, then enters step 5);

[0029] 5) According to the control signal in step 3) or step 4), output the corresponding torque or rotation angle, and transmit it to the wheel to complete vehicle steering.

[0030] Further, the locking or disengaging of the electromagnetic actuator according to the vehicle stability margin in step 2) is as follows:

[0031] 21) Calculate the vehicle stability coefficient as follows:

[0032]

[0033] In the formula, K is the vehicle stability coefficient; m is the vehicle mass; l f is the distance from the vehicle front axle to the center of mass; l r is the distance from the vehicle rear axle to the center of mass; C f is the cornering stiffness of the front axle; C r is the cornering stiffness of the rear axle;

[0034] 22) Calculate the yaw rate corresponding to the vehicle steady-state equilibrium point as follows:

[0035]

[0036] In the formula, is the yaw rate corresponding to the vehicle steady-state equilibrium point; v x is the vehicle longitudinal speed; δ f is the front wheel steering angle;

[0037] 23) Calculate the maximum value of the vehicle yaw rate as follows:

[0038] ω r-max = μg / v x

[0039] In the formula, ω r-max is the maximum value of the vehicle yaw rate; μ is the road adhesion coefficient; g is the acceleration due to gravity;

[0040] 24) Calculate the vehicle stability margin as follows:

[0041]

[0042] In the formula, is the vehicle stability margin; I z is the moment of inertia of the vehicle about the z-axis;

[0043] 25) Calculate the longitudinal energy and lateral energy of the vehicle. The longitudinal energy and lateral energy of the vehicle are expressed as:

[0044]

[0045] In the formula, E x and E y are the longitudinal energy and lateral energy of the vehicle respectively; v y is the lateral speed of the vehicle; ω r is the yaw rate of the vehicle;

[0046] 26) Calculate the lateral energy ratio of the vehicle. The lateral energy ratio of the vehicle is expressed as:

[0047]

[0048] In the formula, S y-x is the lateral energy ratio of the vehicle;

[0049] 27) Compare the lateral energy ratio of the vehicle with the vehicle stability margin. If the lateral energy ratio of the vehicle is less than the vehicle stability margin, the steering system is in the normal condition and the electromagnetic actuator is in the locked state; if the lateral energy ratio of the vehicle is greater than the vehicle stability margin, the steering system is in the limit condition and the electromagnetic actuator is in the disengaged state.

[0050] Furthermore, the assistance characteristic of the main motor in step 3) is linear.

[0051] Furthermore, step 3) specifically includes:

[0052] 31) Calculate the ideal angular transmission ratio of the steering system under the limit condition as follows:

[0053]

[0054] In the formula, i is the ideal angular transmission ratio; L is the wheelbase of the vehicle; K s is the steering sensitivity coefficient, and its value range is 0.12 - 0.417;

[0055] 32) Calculate the required transmission ratio of the double-row planetary gear module as follows:

[0056]

[0057] In the formula, i eg is the transmission ratio of the double-row planetary gear module; i g is the transmission ratio of the planetary roller screw steering gear;

[0058] 33) Calculate the relationship between the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle and the required transmission ratio of the double-row planetary gear module as follows:

[0059] The rotation angle of the auxiliary motor and the rotation angle of the steering wheel satisfy the following relationship:

[0060]

[0061] In the formula, is the rotation angle of the auxiliary motor, k is the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle, is the rotation angle of the steering wheel;

[0062] The rotation angle of the auxiliary motor and the rotation angle of the upper row ring gear satisfy the following relationship:

[0063]

[0064] In the formula, i u is the transmission ratio of the upper row gear of the auxiliary motor and the upper row ring gear, is the rotation angle of the upper row ring gear;

[0065] The rotation angle of the auxiliary motor and the rotation angle of the lower row ring gear satisfy the following relationship:

[0066]

[0067] In the formula, i d is the transmission ratio of the lower row gear of the auxiliary motor and the lower row ring gear, is the rotation angle of the lower row ring gear;

[0068] The upper row planetary gear satisfies the following relationship:

[0069]

[0070] In the formula, is the rotation angle of the upper row sun gear, is the rotation angle of the common planetary carrier, α u is the tooth number ratio of the upper row ring gear and the upper row sun gear;

[0071] The lower row planetary gear satisfies the following relationship:

[0072]

[0073] In the formula, is the rotation angle of the lower row sun gear, α d is the tooth number ratio of the lower row ring gear and the lower row sun gear;

[0074] The rotation angle of the steering wheel is the rotation angle of the upper row sun gear, as follows:

[0075]

[0076] The relationship between the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle and the required transmission ratio of the double-row planetary gear module is as follows:

[0077]

[0078] 34) Calculate the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle as follows:

[0079]

[0080] 35) Calculate the required rotation angle of the auxiliary motor through the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle as follows:

[0081]

[0082] 36) Calculate the steering resistance torque and the steering wheel output torque as follows:

[0083]

[0084] In the formula, T r is the steering resistance torque, f is the sliding friction coefficient between the vehicle tires and the road surface, G1 is the front axle load of the vehicle, and p is the tire pressure of the vehicle;

[0085]

[0086] In the formula, T h is the steering wheel output torque, and η + is the transmission efficiency;

[0087] 37) Calculate the required assist torque value of the main motor, and output the main motor assist torque according to the main motor assist curve, specifically including:

[0088] Calculate the maximum assist torque of the main motor as follows:

[0089] T amax = T h - T i

[0090] In the formula, T amax is the maximum assist torque of the main motor, and T i is the average torque of the steering wheel;

[0091] Calculate the vehicle speed induction coefficient as follows:

[0092]

[0093] In the formula, K V (V) is the vehicle speed induction coefficient, T dmax is the steering wheel torque corresponding to the maximum assist torque of the main motor, and T d0is the dead zone torque of the steering wheel;

[0094] The assist torque characteristic of the main motor is:

[0095]

[0096] In the formula, T a is the assist torque of the main motor, and T d is the steering wheel torque.

[0097] Furthermore, the assist characteristic of the auxiliary motor and the main motor in step 4) is a broken line type.

[0098] Furthermore, step 4) specifically includes:

[0099] 41) Calculate the transmission ratios of the upper planetary gear and the lower planetary gear as follows:

[0100] The upper planetary gear and the lower planetary gear satisfy the following:

[0101]

[0102] The transmission ratio of the upper sun gear and the upper ring gear is:

[0103]

[0104] In the formula, i ueg is the transmission ratio of the upper sun gear and the upper ring gear;

[0105] The transmission ratio of the lower ring gear and the lower sun gear is:

[0106]

[0107] In the formula, i deg is the transmission ratio of the lower ring gear and the lower sun gear;

[0108] 42) Calculate the transmission ratio from the steering wheel to the wheel, the transmission ratio from the auxiliary motor to the wheel, and the transmission ratio from the main motor to the wheel as follows:

[0109] The transmission ratio from the steering wheel to the wheel is as follows:

[0110]

[0111] In the formula, i sw is the transmission ratio from the steering wheel to the wheel;

[0112] The transmission ratio from the auxiliary motor to the wheel is as follows:

[0113] i f = i d i deg i g

[0114] where i f is the transmission ratio from the auxiliary motor to the wheel;

[0115] The transmission ratio from the main motor to the wheel is as follows:

[0116] i z = i g

[0117] where i z is the transmission ratio from the main motor to the wheel;

[0118] 43) Distribute the auxiliary motor and main motor assist torques according to the transmission ratio from the auxiliary motor to the wheel and the transmission ratio from the main motor to the wheel, as follows:

[0119] Calculate the steering resistance torque and the steering wheel output torque, as follows:

[0120]

[0121] Calculate the sum of the maximum assist torques of the auxiliary motor and the main motor, as follows:

[0122] T smax = T h - T i

[0123] where T smax is the sum of the maximum assist torques of the auxiliary motor and the main motor;

[0124] Calculate the vehicle speed induction coefficient, as follows:

[0125]

[0126] where K sV (V) is the total vehicle speed induction coefficient of the sum of the assist torques of the auxiliary motor and the main motor;

[0127] The sum of the assist torques of the auxiliary motor and the main motor is:

[0128]

[0129] where T s is the sum of the assist torques of the auxiliary motor and the main motor;

[0130] Calculate the vehicle speed induction coefficient of the auxiliary motor and the vehicle speed induction coefficient of the main motor respectively, as follows:

[0131]

[0132] where K fV1 (V), K fV2 (V) are both the vehicle speed induction coefficients of the auxiliary motor; K zV1(V), K zV2 (V) are the main motor vehicle speed sensitivity coefficients;

[0133] The auxiliary motor assist torque distribution value is:

[0134]

[0135] In the formula, T f is the auxiliary motor assist torque distribution value; T d1 is the steering wheel torque corresponding to the turning point of the auxiliary motor and main motor assist torque distribution value;

[0136] The main motor assist torque distribution value is:

[0137]

[0138] In the formula, T z is the main motor assist torque distribution value.

[0139] Furthermore, the specific steps of step 5) include:

[0140] Taking the error value e in the auxiliary motor and the main motor as the input of the closed-loop control, the output expression of the controller is as follows:

[0141]

[0142] In the formula, u(t) is the output control quantity of the auxiliary motor and the main motor; e(t) is the input of the closed-loop control; K is the proportional coefficient; T is the integral constant; T d is the differential constant; u0 is the initial control output.

[0143] The beneficial effects of the present invention:

[0144] The present invention integrates the design of a double-row planetary gear module and an electromagnetic actuator, and realizes a multi-functional dual-motor coupled active steering system through the switching of the working states of the electromagnetic actuator and the auxiliary motor; respectively realizes the coupling of the torque outputs of the main motor and the auxiliary motor under normal conditions, and can meet the steering demand power; under extreme conditions, the auxiliary motor and the main motor respectively control the corner and torque of the system, realize variable transmission ratio and active rollover prevention control under extreme conditions, make the system safer and more reliable, and greatly reduce the control complexity of the active steering system;

[0145] The present invention can improve the steering lightness at low speeds more than single-motor electric power steering through dual-motor assist coupling under normal conditions, and through the separate assistance of the main motor under extreme conditions, the auxiliary motor realizes the active compensation of the driver's steering angle through the double-row planetary gear module, can realize stepless variable transmission ratio, and realizes active rollover prevention control.

[0146] The present invention meets the power requirements of the commercial vehicle steering system, and its stepless variable transmission ratio function makes commercial vehicles safer and more reliable. BRIEF DESCRIPTION OF THE DRAWINGS

[0147] Figure 1 It is a schematic structural diagram of the system of the present invention;

[0148] Figure 2 It is a schematic diagram of the double-row planetary gear module of the present invention;

[0149] Figure 3 It is a schematic diagram of the main motor assistance curve under extreme conditions of the present invention;

[0150] Figure 4 It is a schematic diagram of the auxiliary motor assistance curve under normal conditions of the present invention;

[0151] Figure 5 It is a schematic diagram of the main motor assistance curve under normal conditions of the present invention;

[0152] Figure 6 It is a flowchart of the method of the present invention;

[0153] In the figure, 1 - electronic control unit, 2 - double-row planetary gear module, 3 - steering column housing, 4 - steering wheel, 5 - steering column, 6 - angle sensor, 7 - torque sensor, 8 - planetary roller screw steering gear, 9 - tooth sector, 10 - main motor, 11 - steering rocker arm, 12 - wheel, 13 - trapezoidal arm, 14 - steering drag link, 15 - steering tie rod, 16 - upper row ring gear, 17 - upper row planetary gear, 18 - upper row sun gear, 19 - common planetary carrier, 20 - electromagnetic actuator, 21 - upper row gear of auxiliary motor, 22 - lower row gear of auxiliary motor, 23 - force / displacement auxiliary motor, 24 - lower row planetary gear, 25 - lower row sun gear, 26 - lower row ring gear. DETAILED DESCRIPTION OF THE INVENTION

[0154] For the convenience of understanding by those skilled in the art, the present invention will be further described below in conjunction with embodiments and the drawings. The content mentioned in the embodiments does not limit the present invention.

[0155] Referring to Figure 1 、 Figure 2 As shown, a multi-functional dual-motor coupled active steering system for commercial vehicles of the present invention includes: a steering wheel module, a double-row planetary gear module, a steering gear module, and a control module;

[0156] The steering wheel module includes: a steering wheel 4, a steering column 5, and a steering column housing 3. The steering wheel 4 is fixedly connected to the input end of the steering column 5, and the steering column 5 is arranged inside the steering column housing 3;

[0157] The double-row planetary gear module 2 includes: an upper row of planetary gears, a lower row of planetary gears, a common planet carrier 19, an auxiliary motor and an electromagnetic actuator 20;

[0158] The input shaft of the upper row of planetary gears is fixedly connected to the output end of the steering column 6;

[0159] The upper row of planetary gears includes: an upper row of ring gears 16, an upper row of planetary gears 17 and an upper row of sun gears 18. The upper row of planetary gears 17 are mounted on a common planet carrier 19 and are arranged between the upper row of ring gears 16 and the upper row of sun gears 18. The input shaft of the upper row of planetary gears is fixedly connected to the upper row of sun gears 18.

[0160] The lower row of planetary gears includes: a lower row of ring gears 26, a lower row of planetary gears 24 and a lower row of sun gears 25. The lower row of planetary gears 24 are mounted on the common planet carrier 19 and are arranged between the lower row of ring gears 26 and the lower row of sun gears 25. The output shaft of the lower row of planetary gears is fixedly connected to the lower row of sun gears 25.

[0161] The auxiliary motor includes: a force / displacement auxiliary motor 23, an upper gear 21 of the auxiliary motor, and a lower gear 22 of the auxiliary motor; the upper gear 21 of the auxiliary motor and the lower gear 22 of the auxiliary motor are installed on the output shaft of the force / displacement auxiliary motor 23; the upper gear 21 of the auxiliary motor is meshed with the outer side of the upper gear ring 16; the lower gear 22 of the auxiliary motor is meshed with the outer side of the lower gear ring 26;

[0162] The electromagnetic actuator 20 includes a coil and a fixed magnetic core. The coil is arranged outside the fixed magnetic core. The current in the coil generates a magnetic field that interacts with the fixed magnetic core. The fixed magnetic core generates an electromagnetic force to achieve disengagement and locking with the common planet carrier 19.

[0163] The steering gear module includes: a planetary roller screw steering gear 8, a main motor 10, a steering rocker arm 11, a steering straight tie rod 14, a trapezoidal arm 13, a steering tie rod 15, a steering knuckle, a wheel hub and a wheel 12;

[0164] The planetary roller screw steering gear 8 comprises: a screw, a planetary roller, a cage, an inner gear ring, a nut and a gear sector 9; the upper end of the screw is fixedly connected to the output shaft of the lower row of planetary gears; the planetary roller is mounted on the cage and arranged between the screw and the nut; the inner gear ring is mounted on the upper and lower ends of the planetary roller and arranged between the cage and the nut; the nut is meshed with the gear sector 9; the gear sector 9 is fixedly connected to the steering rocker arm 11;

[0165] The steering rocker arm 11 is hinged to the steering straight tie rod 14;

[0166] The steering straight tie rod 14 is hinged to the trapezoidal arm 13;

[0167] The trapezoidal arm 13 is hinged to the steering tie rod 15;

[0168] The tie rod 15 is articulated with the steering knuckle;

[0169] The steering knuckle is fixedly connected to the wheel hub;

[0170] The wheel 12 is fixedly connected to the wheel hub;

[0171] The control module includes: an electronic control unit 1, a steering angle sensor 6 and a torque sensor 7. The steering angle sensor 6 and the torque sensor 7 are installed between the steering column 5 and the steering column housing 3 and are electrically connected to the electronic control unit 1; the electronic control unit is also electrically connected to the electromagnetic actuator 20, the auxiliary motor and the main motor.

[0172] Refer to Figure 6 As shown, the present invention also provides a control method for a multi-functional dual-motor coupled active steering system of a commercial vehicle. Based on the above system, it includes the following steps:

[0173] 1) Collect the steering angle signal and torque signal of the vehicle steering wheel and transmit the signals to the electronic control unit;

[0174] 2) The electronic control unit calculates the vehicle stability margin according to the collected steering angle signal and torque signal, and determines the state of the electromagnetic actuator according to the vehicle stability margin. If the state of the electromagnetic actuator is disengaged, go to step 3); if the state of the electromagnetic actuator is locked, go to step 4);

[0175] Among them, the specific steps of determining the locking or disengaging of the electromagnetic actuator according to the vehicle stability margin in step 2) are as follows:

[0176] 21) Calculate the vehicle stability coefficient as follows:

[0177]

[0178] In the formula, K is the vehicle stability coefficient; m is the vehicle mass; l f is the distance from the vehicle front axle to the center of mass; l r is the distance from the vehicle rear axle to the center of mass; C f is the cornering stiffness of the front axle; C r is the cornering stiffness of the rear axle;

[0179] 22) Calculate the yaw rate corresponding to the vehicle steady-state equilibrium point as follows:

[0180]

[0181] In the formula, is the yaw rate corresponding to the vehicle steady-state equilibrium point; v x is the vehicle longitudinal speed; δ f is the front wheel steering angle;

[0182] 23) Calculate the maximum value of the vehicle yaw rate as follows:

[0183] ω r-max = μg / v x

[0184] In the formula, ω r-max is the maximum value of the vehicle yaw rate; μ is the road adhesion coefficient; g is the acceleration due to gravity;

[0185] 24) Calculate the vehicle stability margin as follows:

[0186]

[0187] In the formula, is the vehicle stability margin; I z is the moment of inertia of the vehicle about the z-axis;

[0188] 25) Calculate the longitudinal energy and lateral energy of the vehicle. The longitudinal energy and lateral energy of the vehicle are expressed as:

[0189]

[0190] In the formula, E x and E y are the longitudinal energy and lateral energy of the vehicle respectively; v y is the vehicle lateral speed; ω r is the vehicle yaw rate;

[0191] 26) Calculate the lateral energy ratio of the vehicle. The lateral energy ratio of the vehicle is expressed as:

[0192]

[0193] In the formula, S y-x is the lateral energy ratio of the vehicle;

[0194] 27) Compare the lateral energy ratio of the vehicle with the vehicle stability margin. If the lateral energy ratio of the vehicle is less than the vehicle stability margin, the steering system is in the normal working condition and the electromagnetic actuator is in the locked state; if the lateral energy ratio of the vehicle is greater than the vehicle stability margin, the steering system is in the limit working condition and the electromagnetic actuator is in the disengaged state.

[0195] 3) The electronic control unit calculates the required transmission ratio under the limit working condition; calculates the required rotation angle value of the auxiliary motor and sends a control signal to the auxiliary motor; calculates the required assist torque value of the main motor and sends a control signal to the main motor, and enters step 5);

[0196] The assist characteristic of the main motor is linear.

[0197] Among them, step 3) specifically includes:

[0198] 31) Calculate the ideal angular transmission ratio of the steering system under extreme conditions as follows:

[0199]

[0200] In the formula, i is the ideal angular transmission ratio; L is the wheelbase of the vehicle; K s is the steering sensitivity coefficient, and its value range is 0.12 - 0.417;

[0201] 32) Calculate the required transmission ratio of the double-row planetary gear module as follows:

[0202]

[0203] In the formula, i eg is the transmission ratio of the double-row planetary gear module; i g is the transmission ratio of the planetary roller screw steering gear;

[0204] 33) Calculate the relationship between the proportional coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle and the required transmission ratio of the double-row planetary gear module as follows:

[0205] The rotation angle of the auxiliary motor and the steering wheel rotation angle satisfy the following relationship:

[0206]

[0207] In the formula, is the rotation angle of the auxiliary motor, k is the proportional coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle, is the steering wheel rotation angle;

[0208] The rotation angle of the auxiliary motor and the rotation angle of the upper row gear ring satisfy the following relationship:

[0209]

[0210] In the formula, i u is the transmission ratio of the upper row gear of the auxiliary motor and the upper row gear ring, is the rotation angle of the upper row gear ring;

[0211] The rotation angle of the auxiliary motor and the rotation angle of the lower row gear ring satisfy the following relationship:

[0212]

[0213] In the formula, i d is the transmission ratio of the lower row gear of the auxiliary motor and the lower row gear ring, is the rotation angle of the lower row gear ring;

[0214] The upper row planetary gear satisfies the following relationship:

[0215]

[0216] Wherein, is the upper row sun gear rotation angle, is the common planet carrier rotation angle, α u is the tooth number ratio of the upper row ring gear to the upper row sun gear;

[0217] The lower row planetary gears satisfy the following relationship:

[0218]

[0219] Wherein, is the lower row sun gear rotation angle, α d is the tooth number ratio of the lower row ring gear to the lower row sun gear;

[0220] The steering wheel rotation angle is the upper row sun gear rotation angle, as follows:

[0221]

[0222] Thus, the relationship between the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle and the required transmission ratio of the double row planetary gear module is as follows:

[0223]

[0224] 34) Calculate the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle as follows:

[0225]

[0226] 35) Calculate the required rotation angle of the auxiliary motor through the proportionality coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle, as follows:

[0227]

[0228] 36) Calculate the steering resistance torque and the steering wheel output torque, as follows:

[0229]

[0230] Wherein, T r is the steering resistance torque, f is the sliding friction coefficient between the vehicle tires and the road surface, G1 is the front axle load of the vehicle, and p is the tire pressure of the vehicle;

[0231]

[0232] Wherein, T h is the steering wheel output torque, η + is the transmission efficiency;

[0233] 37) Calculate the assist torque value required by the main motor, and output the assist torque of the main motor according to the main motor assist curve, specifically including:

[0234] Calculate the maximum assist torque of the main motor as follows:

[0235] T amax = T h - T i

[0236] In the formula, T amax is the maximum assist torque of the main motor, and T i is the average torque of the steering wheel;

[0237] Calculate the vehicle speed induction coefficient as follows:

[0238]

[0239] In the formula, K V (V) is the vehicle speed induction coefficient, T dmax is the steering wheel torque corresponding to the maximum assist torque of the main motor, and T d0 is the steering wheel dead zone torque;

[0240] Refer to Figure 3 As shown, the main motor assist curve under extreme conditions in the present invention is obtained from the following main motor assist torque characteristics:

[0241] The main motor assist torque characteristics are:

[0242]

[0243] In the formula, T a is the main motor assist torque, and T d is the steering wheel torque.

[0244] 4) The electronic control unit calculates the torque distribution values of the auxiliary motor and the main motor under normal conditions, and sends them to the auxiliary motor and the main motor respectively, and enters step 5);

[0245] The assist characteristics of the auxiliary motor and the main motor are in a broken line shape.

[0246] Among them, step 4) specifically includes:

[0247] 41) Calculate the transmission ratios of the upper planetary gear and the lower planetary gear as follows:

[0248] The upper planetary gear and the lower planetary gear satisfy the following:

[0249]

[0250] The transmission ratio of the upper sun gear and the upper ring gear is:

[0251]

[0252] Wherein, i ueg is the transmission ratio between the upper row sun gear and the upper row ring gear;

[0253] The transmission ratio between the lower row ring gear and the lower row sun gear is:

[0254]

[0255] Wherein, i deg is the transmission ratio between the lower row ring gear and the lower row sun gear;

[0256] 42) Calculate the transmission ratio from the steering wheel to the wheel, the transmission ratio from the auxiliary motor to the wheel, and the transmission ratio from the main motor to the wheel as follows:

[0257] The transmission ratio from the steering wheel to the wheel is as follows:

[0258]

[0259] Wherein, i sw is the transmission ratio from the steering wheel to the wheel;

[0260] The transmission ratio from the auxiliary motor to the wheel is as follows:

[0261] i f = i d i deg i g

[0262] Wherein, i f is the transmission ratio from the auxiliary motor to the wheel;

[0263] The transmission ratio from the main motor to the wheel is as follows:

[0264] i z = i g

[0265] Wherein, i z is the transmission ratio from the main motor to the wheel;

[0266] 43) Allocate the assisting torques of the auxiliary motor and the main motor according to the transmission ratio from the auxiliary motor to the wheel and the transmission ratio from the main motor to the wheel as follows:

[0267] Calculate the steering resistance torque and the output torque of the steering wheel as follows:

[0268]

[0269] Calculate the sum of the maximum assisting torques of the auxiliary motor and the main motor as follows:

[0270] T smax = T h - Ti

[0271] Wherein, T smax is the sum of the maximum boost torques of the auxiliary motor and the main motor;

[0272] Calculate the vehicle speed induction coefficient as follows:

[0273]

[0274] Wherein, K sV (V) is the total vehicle speed induction coefficient of the sum of the boost torques of the auxiliary motor and the main motor;

[0275] The sum of the boost torques of the auxiliary motor and the main motor is:

[0276]

[0277] Wherein, T s is the sum of the boost torques of the auxiliary motor and the main motor;

[0278] Calculate the vehicle speed induction coefficient of the auxiliary motor and the vehicle speed induction coefficient of the main motor respectively as follows:

[0279]

[0280] Wherein, K fV1 (V), K fV2 (V) are both the vehicle speed induction coefficients of the auxiliary motor; K zV1 (V), K zV2 (V) are both the vehicle speed induction coefficients of the main motor;

[0281] Refer to Figure 4 As shown, in the present invention, the auxiliary motor boost curve under normal conditions is obtained from the following auxiliary motor boost torque distribution values:

[0282] The auxiliary motor boost torque distribution value is:

[0283]

[0284] Wherein, T f is the auxiliary motor boost torque distribution value; T d1 is the steering wheel torque corresponding to the turning point of the auxiliary motor and main motor boost torque distribution value;

[0285] Refer to Figure 5 As shown, in the present invention, the main motor boost curve under normal conditions is obtained from the following main motor boost torque distribution values:

[0286] The main motor boost torque distribution value is:

[0287]

[0288] Wherein, Tz It is the assist torque distribution value of the main motor.

[0289] 5) According to the control signal in step 3) or step 4), output the corresponding torque or rotation angle, transmit it to the wheels, and complete the vehicle steering; specifically including:

[0290] Take the error value e in the auxiliary motor and the main motor as the input of the closed-loop control. The output expression of the controller is as follows:

[0291]

[0292] In the formula, u(t) is the output control quantity of the auxiliary motor and the main motor; e(t) is the input of the closed-loop control; K is the proportionality coefficient; T is the integral constant; T d is the differential constant; u0 is the initial control output.

[0293] The specific application ways of the present invention are numerous. The above description is only the preferred embodiment of the present invention. It should be pointed out that for those of ordinary skill in the art of this technology, several improvements can be made without departing from the principle of the present invention, and these improvements should also be regarded as the protection scope of the present invention.

Claims

1. A multi-functional dual-motor coupled active steering system for commercial vehicles, characterized in that, Including: A steering wheel module, a double-row planetary gear module, a steering gear module and a control module; The steering wheel module includes a steering wheel, a steering column and a steering column housing. The steering wheel is fixedly connected to the input end of the steering column, and the steering column is arranged inside the steering column housing; The double-row planetary gear module includes an upper-row planetary gear, a lower-row planetary gear, a common planet carrier, an auxiliary motor and an electromagnetic actuator; The input shaft of the upper-row planetary gear is fixedly connected to the output end of the steering column; The upper-row planetary gear includes an upper-row ring gear, upper-row planet gears and an upper-row sun gear. The upper-row planet gears are installed on the common planet carrier and are arranged between the upper-row ring gear and the upper-row sun gear; the input shaft of the upper-row planetary gear is fixedly connected to the upper-row sun gear; The lower-row planetary gear includes a lower-row ring gear, lower-row planet gears and a lower-row sun gear. The lower-row planet gears are installed on the common planet carrier and are arranged between the lower-row ring gear and the lower-row sun gear; the output shaft of the lower-row planetary gear is fixedly connected to the lower-row sun gear; The auxiliary motor includes a force / displacement auxiliary motor, an upper-row gear of the auxiliary motor and a lower-row gear of the auxiliary motor; the upper-row gear of the auxiliary motor and the lower-row gear of the auxiliary motor are installed on the output shaft of the force / displacement auxiliary motor; the upper-row gear of the auxiliary motor meshes with the outer side of the upper-row ring gear; the lower-row gear of the auxiliary motor meshes with the outer side of the lower-row ring gear; The electromagnetic actuator includes a coil and a fixed magnetic core. The coil is arranged outside the fixed magnetic core. A magnetic field is generated by the current passing through the coil and interacts with the fixed magnetic core, and an electromagnetic force is generated on the fixed magnetic core to disengage and lock with the common planet carrier; The steering gear module includes a planetary roller screw steering gear, a main motor, a steering arm, a steering drag link, a trapezoidal arm, a steering cross link, a steering knuckle, a wheel hub and a wheel; The planetary roller screw steering gear includes a screw, planetary rollers, a cage, an internal gear ring, a nut and a tooth sector; the upper end of the screw is fixedly connected to the output shaft of the lower-row planetary gear; the planetary rollers are installed on the cage and are arranged between the screw and the nut; the internal gear ring is installed at the upper and lower ends of the planetary rollers and is arranged between the cage and the nut; the nut meshes with the tooth sector; the tooth sector is fixedly connected to the steering arm; The steering arm is hinged to the steering drag link; The steering drag link is hinged to the trapezoidal arm; The trapezoidal arm is hinged to the steering cross link; The steering cross link is hinged to the steering knuckle; The steering knuckle is fixedly connected to the wheel hub; The wheel is fixedly connected to the wheel hub; The control module includes an electronic control unit, a steering angle sensor and a torque sensor. The steering angle sensor and the torque sensor are installed between the steering column and the steering column housing and are electrically connected to the electronic control unit; the electronic control unit is also electrically connected to the electromagnetic actuator, the auxiliary motor and the main motor.

2. A control method for a multi-functional dual-motor coupled active steering system of a commercial vehicle, based on the system described in claim 1, characterized in that, The method steps are as follows: 1) Collect the steering angle signal and torque signal of the vehicle steering wheel and transmit the signals to the electronic control unit; 2) The electronic control unit calculates the vehicle stability margin according to the collected steering angle signal and torque signal, and determines the state of the electromagnetic actuator according to the vehicle stability margin. If the state of the electromagnetic actuator is disengaged, go to step 3); if the state of the electromagnetic actuator is locked, go to step 4); 3) The electronic control unit calculates the required transmission ratio under extreme conditions; calculates the required rotation angle value of the auxiliary motor, and sends a control signal to the auxiliary motor; calculates the required assist torque value of the main motor, and sends a control signal to the main motor, then enters step 5). 4) The electronic control unit calculates the torque distribution values of the auxiliary motor and the main motor under normal conditions, and sends them to the auxiliary motor and the main motor respectively, then enters step 5). 5) According to the control signal in step 3) or step 4), output the corresponding torque or rotation angle, and transfer it to the wheels to complete vehicle steering.

3. The control method of the multi-functional dual-motor coupled active steering system for commercial vehicles according to claim 2, characterized in that, In step 2), determining the locking or disengaging of the electromagnetic actuator according to the vehicle stability margin is specifically as follows: 21) Calculate the vehicle stability coefficient as follows: Wherein, K is the vehicle stability coefficient; m is the vehicle mass; l f is the distance from the front axle of the vehicle to the center of mass; l r is the distance from the rear axle of the vehicle to the center of mass; C f is the cornering stiffness of the front axle; C r is the cornering stiffness of the rear axle; 22) Calculate the yaw rate corresponding to the vehicle's steady-state equilibrium point as follows: In the formula, is the yaw rate corresponding to the vehicle's steady-state equilibrium point; v x is the vehicle's longitudinal speed; δ f is the front-wheel steering angle; 23) Calculate the maximum value of the vehicle's yaw rate as follows: ω r-max = μg / v x where ω r-max is the maximum value of the vehicle's yaw rate; μ is the road adhesion coefficient; g is the acceleration due to gravity; 24) Calculate the vehicle stability margin as follows: In the formula, is the vehicle stability margin; I z is the moment of inertia of the vehicle about the z-axis; 25) Calculate the longitudinal energy and lateral energy of the vehicle. The longitudinal energy and lateral energy of the vehicle are expressed as: where, E x and E y are the longitudinal energy and lateral energy of the vehicle respectively; v y is the lateral speed of the vehicle; ω r is the yaw rate of the vehicle; 26) Calculate the lateral energy ratio of the vehicle. The lateral energy ratio of the vehicle is expressed as: Where S y-x is the lateral energy ratio of the vehicle; 27) Compare the lateral energy ratio of the vehicle with the vehicle stability margin. If the lateral energy ratio of the vehicle is less than the vehicle stability margin, the steering system is in the normal condition and the electromagnetic actuator is in the locked state; if the lateral energy ratio of the vehicle is greater than the vehicle stability margin, the steering system is in the extreme condition and the electromagnetic actuator is in the disengaged state.

4. The control method of the multi-functional dual-motor coupled active steering system for commercial vehicles according to claim 3, characterized in that, Step 3) specifically includes: 31) Calculate the ideal angular transmission ratio of the steering system under extreme conditions as follows: where i is the ideal angular transmission ratio; L is the wheelbase of the vehicle; K s is the steering sensitivity coefficient, and the value range is 0.12 - 0.417; 32) Calculate the required transmission ratio of the double-row planetary gear module as follows: where, i eg is the transmission ratio of the double-row planetary gear module; i g is the transmission ratio of the planetary roller screw steering gear; 33) Calculate the relationship between the proportional coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle and the required transmission ratio of the double-row planetary gear module as follows: The relationship between the auxiliary motor rotation angle and the steering wheel rotation angle satisfies the following: In the formula, is the rotation angle of the auxiliary motor, k is the proportionality coefficient between the rotation angle of the auxiliary motor and the rotation angle of the steering wheel, is the rotation angle of the steering wheel; The relationship between the auxiliary motor rotation angle and the rotation angle of the upper row ring gear satisfies the following: where \(i\) u is the transmission ratio of the upper row gear and the upper row gear ring of the auxiliary motor, is the rotation angle of the upper row gear ring; The relationship between the auxiliary motor rotation angle and the rotation angle of the lower row ring gear satisfies the following: where i d is the transmission ratio of the lower row gear of the auxiliary motor to the lower row gear ring, is the rotation angle of the lower row gear ring; The upper row planetary gear satisfies the following: In the formula, is the rotation angle of the upper row sun gear, is the rotation angle of the common planetary carrier, and α u is the tooth number ratio of the upper row ring gear to the upper row sun gear; The lower row planetary gear satisfies the following: In the formula, is the rotation angle of the lower row sun gear, and α d is the tooth number ratio of the lower row ring gear to the lower row sun gear; The steering wheel rotation angle is the rotation angle of the upper row sun gear, as follows: Thus, the relationship between the proportional coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle and the required transmission ratio of the double-row planetary gear module is as follows: 34) Calculate the proportional coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle as follows: 35) Calculate the required rotation angle of the auxiliary motor through the proportional coefficient of the auxiliary motor rotation angle and the steering wheel rotation angle as follows: 36) Calculate the steering resistance torque and the steering wheel output torque as follows: where T r is the steering resistance moment, f is the sliding friction coefficient between the vehicle tires and the road surface, G1 is the front axle load of the vehicle, and p is the tire pressure of the vehicle; Where, T h is the output torque of the steering wheel, and η + is the transmission efficiency; 37) Calculate the required assist torque value of the main motor, and output the main motor assist torque according to the main motor assist curve, specifically including: Calculate the maximum assist torque of the main motor as follows: T amax = T h - T i where T amax is the maximum assist torque of the main motor, and T i is the average torque of the steering wheel; Calculate the vehicle speed induction coefficient as follows: where K V (V) is the vehicle speed sensing coefficient, T dmax is the steering wheel torque corresponding to the maximum assist torque of the main motor, T d0 is the steering wheel dead zone torque; The main motor assist torque characteristic is: where T a is the assist torque of the main motor, and T d is the steering wheel torque.

5. The control method of the multi-functional dual-motor coupled active steering system for commercial vehicles according to claim 4, wherein Step 4) specifically includes: 41) Calculate the transmission ratios of the upper row planetary gear and the lower row planetary gear as follows: The upper row planetary gear and the lower row planetary gear satisfy the following: The transmission ratio of the upper row sun gear and the upper row ring gear is: where i ueg is the transmission ratio between the upper sun gear and the upper ring gear; The transmission ratio of the lower row ring gear and the lower row sun gear is: where i deg is the transmission ratio between the lower row ring gear and the lower row sun gear; 42) Calculate the transmission ratio from the steering wheel to the wheels, the transmission ratio from the auxiliary motor to the wheels, and the transmission ratio from the main motor to the wheels as follows: The transmission ratio from the steering wheel to the wheels is as follows: where i sw is the transmission ratio from the steering wheel to the wheel; The transmission ratio from the auxiliary motor to the wheels is as follows: i f = i d i deg i g where i f is the transmission ratio from the auxiliary motor to the wheel; The transmission ratio from the main motor to the wheels is as follows: i z = i g where i z is the transmission ratio from the main motor to the wheels; 43) Distribute the auxiliary motor and main motor assist torques according to the transmission ratio from the auxiliary motor to the wheels and the transmission ratio from the main motor to the wheels, as follows: Calculate the steering resistance torque and the steering wheel output torque, as follows: Calculate the sum of the maximum assist torques of the auxiliary motor and the main motor, as follows: T smax = T h - T i where T smax is the sum of the maximum assist torques of the auxiliary motor and the main motor; Calculate the vehicle speed induction coefficient, as follows: where K sV (V) is the total vehicle speed induction coefficient of the sum of the auxiliary motor and the main motor boost torque; The sum of the auxiliary motor and main motor assist torques is: where T s is the sum of the assisting torques of the auxiliary motor and the main motor; Calculate the auxiliary motor vehicle speed induction coefficient and the main motor vehicle speed induction coefficient respectively, as follows: where K fV1 (V) and K fV2 (V) are both the speed sensitivity coefficients of the auxiliary motor; K zV1 (V) and K zV2 (V) are both the speed sensitivity coefficients of the main motor; The assigned value of the auxiliary motor assist torque is: where T f is the auxiliary motor assist torque distribution value; T d1 is the steering wheel torque corresponding to the turning point of the auxiliary motor and main motor assist torque distribution value. The assigned value of the main motor assist torque is: where T z is the main motor assist torque distribution value.

6. The control method of the multi-functional dual-motor coupled active steering system for commercial vehicles according to claim 5, characterized in that, The specific content of step 5) includes: Use the error value e in the auxiliary motor and the main motor as the input of the closed-loop control, and the output expression of the controller is as follows: Wherein, u(t) is the output control quantity of the auxiliary motor and the main motor; e(t) is the input of the closed-loop control; K is the proportionality coefficient; T is the integral constant; T d is the differential constant; u0 is the initial control output.

7. The control method of the multi-functional dual-motor coupled active steering system for commercial vehicles according to claim 2, characterized in that, The assist characteristic of the main motor in step 3) is linear.

8. The control method of the multi-functional dual-motor coupled active steering system for commercial vehicles according to claim 2, characterized in that, The assist characteristics of the auxiliary motor and the main motor in step 4) are polyline-shaped.

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