A six-wheel vehicle hydraulic semi-active suspension and control method thereof

Abstract

The present invention discloses a hydraulic semi-active suspension for a six-wheeled vehicle and a control method thereof. The tire knuckle and the suspension module are both installed on the vehicle body, the cross arm in the suspension module is connected to each tire, the suspension actuator is connected to the hydraulic module, the hydraulic module is used to control the elastic force and damping force between the suspension actuator and the vehicle body, and then the six-wheeled vehicle is controlled to travel smoothly under different driving conditions through the suspension module; the method includes constructing a multi-body dynamics model of the six-wheeled vehicle and a semi-active suspension control strategy, using the multi-body dynamics model of the six-wheeled vehicle to obtain the dynamic stroke and dynamic stroke speed of each suspension actuator, and using the semi-active suspension control strategy to set the damping and stiffness of the suspension actuator. The present invention uses the hydraulic module to adjust the stiffness coefficient and damping coefficient of the suspension in real time, effectively suppresses the vertical acceleration of the vehicle body, the pitch angle of the vehicle body, and the roll angle of the vehicle body during the driving of the six-wheeled vehicle, and effectively improves the driving stability of the six-wheeled vehicle.

Description

Technical Field

[0001] The invention belongs to the field of vehicle suspension system control, and in particular relates to a hydraulic semi-active suspension for a six-wheel vehicle and a control method thereof. Background Art

[0002] At present, vehicle suspension mainly includes three types: passive, semi-active, and active suspension. Among them, passive suspension has problems such as poor dynamic performance and fixed performance that cannot be adjusted according to driving conditions, while active suspension has problems such as slow response speed, high energy consumption, and low energy efficiency. Therefore, semi-active suspension has become a solution to the current problems of insufficient performance and high energy consumption of vehicle suspension systems, and is currently a hot research area for universities and major automobile companies.

[0003] Generally speaking, a semi-active suspension system usually includes sensors for monitoring the vehicle's motion state and road conditions, such as vehicle speed, acceleration, etc. The data collected by these sensors are sent to a control unit, which analyzes the current road conditions and vehicle status based on these data. Based on the analysis results, the control unit can adjust the parameters of the suspension system, such as damping force or elastic force, to adapt to current driving needs and road conditions. By dynamically adjusting the suspension parameters, the semi-active suspension system can optimize the vehicle's stability, handling and ride comfort, while reducing the vehicle's bumpy feeling on poor road conditions and improving the vehicle's grip. The prior art lacks a hydraulic semi-active suspension that can improve the driving stability of six-wheeled vehicles. Summary of the invention

[0004] In order to solve the problems existing in the background technology, the purpose of the present invention is to provide a hydraulic semi-active suspension for a six-wheeled vehicle and a control method thereof, so as to solve the problem of poor driving stability of six-wheeled vehicles in the prior art.

[0005] The technical solution adopted by the present invention is:

[0006] 1. A hydraulic semi-active suspension for a six-wheel vehicle, characterized in that:

[0007] It includes a vehicle body, six tire knuckles, six suspension modules and six hydraulic modules; the six tire knuckles and the six suspension modules are all installed on the vehicle body, the cross arms in each suspension module are connected to each tire knuckle, the suspension actuators in each suspension module are connected to each hydraulic module, the hydraulic modules are used to control the elastic force and damping force between the suspension actuator and the vehicle body, and then control the smooth driving of the six-wheeled vehicle under different driving conditions through the suspension modules.

[0008] The suspension module includes a suspension spring, a suspension actuator, an upper cross arm and a lower cross arm; the upper cross arm is located above the lower cross arm, the upper cross arm is hinged between the middle part of the vehicle body and the tire knuckle, the lower cross arm is hinged between the lower part of the vehicle body and the tire knuckle, the retractable suspension actuator is hinged between the top of the vehicle body and the middle part of the lower cross arm, the suspension spring is wound on the outer side wall of the suspension actuator, and the top end of the suspension spring is hinged at the connection node between the suspension actuator and the vehicle body, and the bottom end of the suspension spring is hinged at the connection node between the suspension actuator and the lower cross arm.

[0009] The hydraulic module includes a stiffness adjustment module and a damping adjustment module. The rodless chamber of the suspension actuator in the suspension module is connected to one end of the damping adjustment module through a pipeline, and the other end of the damping adjustment module is connected to the stiffness adjustment module. The damping adjustment module is used to adjust the damping force of the suspension actuator. The stiffness adjustment module is mainly composed of a primary accumulator, a secondary accumulator and an electromagnetic switch valve. The damping adjustment module is connected to the secondary accumulator of the stiffness adjustment module through a pipeline. The primary accumulator and the electromagnetic switch valve are sequentially arranged on the pipeline from the damping adjustment module to the secondary accumulator. The electromagnetic switch valve is used to adjust the equivalent stiffness of the suspension actuator.

[0010] The damping adjustment module and the electromagnetic switch valve are both externally connected to a control system, and the control system is used to control the opening and closing state of the electromagnetic switch valve and adjust the equivalent damping coefficient of the damping adjustment module.

[0011] The suspension actuator adopts an asymmetric hydraulic cylinder, and the damping adjustment module adopts an adjustable flow valve.

[0012] 2. A control method for a hydraulic semi-active suspension of a six-wheel vehicle, comprising the following steps:

[0013] Step S1, firstly obtaining the geometric characteristics, speed and spring load information of the six-wheeled vehicle during driving;

[0014] Step S2: Next, a multi-body dynamics model of the six-wheeled vehicle and a semi-active suspension control strategy are constructed in a computer according to the geometric characteristics, speed and spring load information of the six-wheeled vehicle;

[0015] Step S3, using a six-wheel vehicle multi-body dynamics model to obtain the dynamic stroke and dynamic stroke speed of each suspension actuator;

[0016] Step S4, inputting the dynamic stroke and dynamic stroke speed of step S3 into the semi-active suspension control strategy, and using the processor to obtain target values ​​of the equivalent damping coefficient and equivalent stiffness coefficient of the suspension actuator;

[0017] Step S5: Use the control system to control the opening and closing state of the electromagnetic switch valve and the opening of the adjustable flow valve in the damping adjustment module in real time, so that the equivalent damping coefficient and the equivalent stiffness coefficient of the suspension actuator reach the target value of step S4, so that the six-wheeled vehicle can run smoothly under different road conditions.

[0018] The multi-body dynamics model of the six-wheeled vehicle in step S2 is as follows:

[0019] m u,i ×a u,i +F HS,i + F RS,i +k t (z u,i -z r,i )=0,i∈{1,2,3,4,5,6}

[0020] m s ×a s =∑ 6 i=1 F HS,i +∑ 6 i=1 F RS,i

[0021] I xx ×a φ =(F HS,1 +F HS,3 +F HS,5 )b / 2-(F HS,2 +F HS,4 +F HS,6 )b / 2+(F RS,1 +F RS,3 +F RS,5 )b / 2-(F RS,2 +F RS,4 +F RS,6 )b / 2

[0022] I yy ×a θ =-(F HS,1 +F HS,2 ) f -(F HS,3 +F HS,4 ) m +(F HS,5 +F HS,6 ) r -(F RS,1 +F RS,2 ) f -(F RS,3 +F RS,4 ) m +(F RS,5 +F RS,6) r

[0023] Among them, m u,i represents the unsprung mass at the position of the i-th tire; i represents the ordinal number of the tire, and the 1st to 6th tires are the left front tire, right front tire, left middle tire, right middle tire, left rear tire and right rear tire of the six-wheel vehicle respectively; z u,i represents the vertical displacement of the unsprung mass at the i-th tire position; a u,i represents the vertical displacement z of the unsprung mass at the i-th tire position u,i Find the second-order derivative with respect to time t; F HS,i represents the vertical force of the suspension actuator at the position of the i-th tire; F RS,i represents the vertical force of the suspension spring at the position of the i-th tire; k t Represents the equivalent stiffness coefficient of each tire; z r,i represents the road input displacement at the i-th tire position; m s Indicates the sprung mass of the vehicle; s represents the vertical displacement of the sprung mass center; a s is the vertical displacement z at the center of mass of the sprung mass s Find the second-order derivative with respect to time t; I xx I represents the moment of inertia of the sprung mass around the vehicle's x-axis; yy represents the moment of inertia of the sprung mass about the vehicle's y-axis; a θ represents the angular acceleration of the pitch angle at the center of mass of the sprung mass; a φ represents the angular acceleration of the roll angle at the center of mass of the sprung mass; b represents the distance between the two tires in the width direction of the six-wheeled vehicle; l f It represents the distance between the front wheel and the center of mass of the sprung mass in the length direction of the six-wheel vehicle; l m It represents the distance between the middle wheel and the center of mass of the sprung mass in the length direction of a six-wheeled vehicle; l r It represents the distance between the rear wheels and the center of mass of the sprung mass in the length direction of a six-wheeled vehicle.

[0024] In the multi-body dynamics model of the six-wheeled vehicle, the vertical force F of the suspension actuator is HS,i , the vertical force F of the suspension spring RS,i According to the following formula:

[0025] F HS,i =-c i (v s,i -v u,i )-k c,i (z s,i -z u,i )

[0026] FRS,i =-k b (z s,i -z u,i )

[0027] Among them, c i represents the equivalent damping coefficient of the suspension actuator at the i-th tire position; z s,i represents the vertical displacement of the connection between the suspension actuator and the vehicle body at the position of the i-th tire; v s,i represents the vertical displacement z of the connection between the suspension actuator and the vehicle body at the i-th tire position s,i Find the first-order derivative with respect to time; v u,i represents the vertical displacement z of the unsprung mass at the i-th tire position u,i Find the first-order derivative with respect to time t; k c,i represents the equivalent stiffness coefficient of the suspension actuator at the i-th tire position; k b Indicates the stiffness of the suspension spring.

[0028] The semi-active suspension control strategy in step S4 includes equivalent damping control and equivalent stiffness control of the suspension actuator. The equivalent damping control of the suspension actuator includes:

[0029] If the vertical velocity v at the connection between the suspension actuator and the vehicle body at the i-th tire position s,i The suspension travel speed of the suspension actuator at the i-th tire position is opposite to that of the suspension actuator, so the equivalent damping coefficient c of the suspension actuator is i The target value is the maximum value of the equivalent damping coefficient c max ;

[0030] If the vertical velocity v at the connection between the suspension actuator and the vehicle body at the i-th tire position s,i The target value of the equivalent damping coefficient of the suspension actuator is c, which is the same as the suspension travel speed of the suspension actuator at the i-th tire position. i Take the minimum value of equivalent damping coefficient c min .

[0031] The equivalent stiffness control of the suspension actuator includes:

[0032] If the pitch angle θ at the center of mass of the vehicle body sprung mass is greater than 0, the target value of the equivalent stiffness coefficient of the suspension actuator at the first and second tire positions is the maximum value of the equivalent stiffness coefficient k cmax , the target value of the equivalent stiffness coefficient of the suspension actuator at the 5th and 6th tire positions is the minimum value of the equivalent stiffness coefficient k cmin ;

[0033] If the pitch angle θ at the center of mass of the vehicle body sprung mass is less than 0, the target value of the equivalent stiffness coefficient of the suspension actuator at the first and second tire positions is the minimum value of the equivalent stiffness coefficient kcmin , the target value of the equivalent stiffness coefficient of the suspension actuator at the 5th and 6th tire positions is the maximum value of the equivalent stiffness coefficient k cmax ;

[0034] If the pitch angle θ at the center of mass of the vehicle body sprung mass is 0, the target values ​​of the equivalent stiffness coefficients of the suspension actuators at the positions of the 1st, 2nd, 5th, and 6th tires are all taken as the minimum value of the equivalent stiffness coefficient k cmin ;

[0035] If the roll angle φ at the center of mass of the vehicle body sprung mass is greater than 0, the equivalent stiffness coefficient k of the suspension actuator at the third tire position is c,3 The target value is the minimum value of the equivalent stiffness coefficient k cmin , the equivalent stiffness coefficient k of the suspension actuator at the 4th tire position c,4 The target value is the maximum value of the equivalent stiffness coefficient k cmax ;

[0036] If the roll angle φ at the center of mass of the vehicle body sprung mass is less than 0, the equivalent stiffness coefficient k of the suspension actuator at the third tire position is c,3 The target value is the maximum value of the equivalent stiffness coefficient k cmax , the equivalent stiffness coefficient k of the suspension actuator at the 4th tire position c,4 The target value is the minimum value of the equivalent stiffness coefficient k cmin ;

[0037] If the roll angle φ at the center of mass of the sprung mass of the vehicle body is 0, the target value of the equivalent stiffness coefficient of the suspension actuator at the third and fourth tire positions is the minimum value of the equivalent stiffness coefficient k cmin .

[0038] The present invention firstly constructs a multi-body dynamics model of a six-wheel vehicle in a computer according to the geometric characteristics, speed and sprung load information of the six-wheel vehicle; secondly, according to the characteristics of the hydraulic suspension, a semi-active suspension control strategy of stiffness-damping coordinated control is designed based on the driving stability of the six-wheel vehicle; according to the performance parameters of the suspension module, a suspension hydraulic model is built, a joint simulation interface is established, and a Simulink-AMESim joint simulation platform is constructed; finally, the joint simulation platform is used to compare the performance of the semi-active suspension and the passive suspension.

[0039] The beneficial effects of the present invention are:

[0040] 1. The present invention utilizes a hydraulic module to adjust the stiffness coefficient and damping coefficient of any suspension in real time, effectively suppressing the vertical acceleration, pitch angle, and roll angle of the vehicle body during the driving of the six-wheeled vehicle, and effectively improving the driving stability of the six-wheeled vehicle.

[0041] 2. The semi-active suspension control method of the present invention, which is based on the multi-body dynamics design of the six-wheeled vehicle, realizes the coordinated control of the stiffness and damping of each suspension actuator of the six-wheeled vehicle. Compared with the control method of controlling the stiffness or damping separately, the six-wheeled vehicle has better stability and controllability.

[0042] 3. The present invention constructs a six-wheel vehicle joint simulation model in a computer according to the geometric characteristics of the six-wheel vehicle and the performance parameters of the suspension module, and develops and verifies a semi-active suspension control method through the joint simulation model, which effectively improves the research and development efficiency and reduces the research and development cost, and provides a reliable technical path for the research and development of the semi-active suspension control method. BRIEF DESCRIPTION OF THE DRAWINGS

[0043] Figure 1 is a schematic diagram of the semi-active suspension structure of the present invention;

[0044] Figure 2 is a schematic diagram of a hydraulic module of the present invention;

[0045] Figure 3 is a flow chart of the control method of the present invention;

[0046] Figure 4 This is a schematic diagram of the hydraulic principle of the one-sixth suspension;

[0047] Figure 5 This is a comparison chart of vehicle body acceleration simulation;

[0048] Figure 6 This is a comparison chart of the vehicle body pitch angle simulation;

[0049] Figure 7 This is a comparison chart of the vehicle body roll angle simulation;

[0050] Figure 8 This is a comparison chart of suspension dynamic travel simulation;

[0051] Fig. 9 This is a comparison chart of tire dynamic load simulation.

[0052] In the figure: 1. Vehicle body; 2. Suspension spring; 3. Suspension actuator; 4. Upper wishbone; 5. Lower wishbone; 6. Tire knuckle; 7. Stiffness adjustment module; 8. Damping adjustment module. DETAILED DESCRIPTION

[0053] The present invention is described in detail below in conjunction with specific implementation cases. The following implementation cases will help those skilled in the art to further understand the present invention, but will not limit the present invention in any form.

[0054] like Figure 1As shown, the semi-active suspension includes a vehicle body 1, six tire knuckles 6, six suspension modules and six hydraulic modules; the six tire knuckles 6 and the six suspension modules are all installed on the vehicle body 1, the cross arms in each suspension module are connected to each tire knuckle 6, the suspension actuators 3 in each suspension module are connected to each hydraulic module, and the hydraulic modules are used to control the elastic force and damping force between the suspension actuators 3 and the vehicle body 1, thereby controlling the smooth driving of the six-wheeled vehicle under different driving conditions through the suspension modules.

[0055] The vehicle body 1, the tire knuckle 6, the suspension module and the hydraulic module together constitute the six-wheel vehicle of the present invention. The tire knuckle 6 is mainly composed of the wheels and the knuckles on the six-wheel vehicle, and each knuckle is fixedly connected to each wheel. In a specific implementation, the front direction of the semi-active suspension is specifically the forward direction of the six-wheel vehicle, and the left and right directions of the semi-active suspension are respectively the left and right directions of the six-wheel vehicle. The six wheels are respectively installed at the left front, right front, left middle, right middle, left rear and right rear positions of the vehicle body 1 of the six-wheel vehicle. Each suspension module is only connected to one tire knuckle 6, and each suspension module is only connected to one hydraulic module.

[0056] The suspension module includes a suspension spring 2, a suspension actuator 3, an upper cross arm 4 and a lower cross arm 5; the upper cross arm 4 is located above the lower cross arm 5, the upper cross arm 4 is hinged between the middle part of the vehicle body 1 and the knuckle in the tire knuckle 6, the lower cross arm 5 is hinged between the lower part of the vehicle body 1 and the knuckle in the tire knuckle 6, the retractable suspension actuator 3 is hinged between the top of the vehicle body 1 and the middle part of the lower cross arm 5, the suspension spring 2 is wound on the outer side wall of the suspension actuator 3, and the top end of the suspension spring 2 is hinged at the connection node between the suspension actuator 3 and the vehicle body 1, and the bottom end of the suspension spring 2 is hinged at the connection node between the suspension actuator 3 and the lower cross arm 5.

[0057] Specifically, the top end of the suspension actuator 3 is an end of the rodless cavity, and the bottom end of the suspension actuator 3 is an end of the rod cavity.

[0058] like Figure 2 As shown, the hydraulic module includes a stiffness adjustment module 7 and a damping adjustment module 8. The rodless cavity of the suspension actuator 3 in the suspension module is connected to one end of the damping adjustment module 8 through a pipeline, and the other end of the damping adjustment module 8 is connected to the stiffness adjustment module 7. The damping adjustment module 8 is used to adjust the damping force of the suspension actuator 3. The stiffness adjustment module 7 is mainly composed of a primary accumulator, a secondary accumulator and an electromagnetic switch valve. The other end of the damping adjustment module 8 is connected to the secondary accumulator of the stiffness adjustment module 7 through a pipeline. The primary accumulator and the electromagnetic switch valve are sequentially arranged on the pipeline from the damping adjustment module 8 to the secondary accumulator. The electromagnetic switch valve is used to adjust the equivalent stiffness of the suspension actuator 3.

[0059] The damping adjustment module 8 and the electromagnetic switch valve are both externally connected to a control system, and the control system is used to control the opening and closing state of the electromagnetic switch valve and adjust the equivalent damping coefficient of the damping adjustment module 8.

[0060] When the electromagnetic switch valve is opened, the primary accumulator and the secondary accumulator are connected, and the equivalent stiffness coefficient of the suspension actuator 3 reaches the minimum value; when the electromagnetic switch valve is closed, the primary accumulator and the secondary accumulator are not connected, the suspension actuator 3 is only connected to the primary accumulator, and the secondary accumulator does not participate in the stiffness adjustment of the suspension actuator 3. At this time, the equivalent stiffness coefficient of the suspension actuator 3 reaches the maximum value.

[0061] The suspension actuator 3 adopts an asymmetric hydraulic cylinder, and the damping adjustment module 8 adopts an adjustable flow valve.

[0062] The embodiment of the present invention includes the following steps: Figure 3 As shown:

[0063] Step S1, firstly obtaining the geometric characteristics, speed and spring load information of the six-wheeled vehicle during driving;

[0064] Step S2: Next, a multi-body dynamics model of a six-wheeled vehicle and a semi-active suspension control strategy are constructed in a computer according to the geometric characteristics, speed and spring load information of the six-wheeled vehicle. The semi-active suspension control strategy is used to obtain the equivalent damping and equivalent stiffness of the suspension actuator 3 according to the dynamic stroke and dynamic stroke speed of the suspension actuator 3;

[0065] Among them, the geometric characteristic information of the six-wheel vehicle includes the width and length of the six-wheel vehicle; the sprung information includes the sprung mass, sprung mass center of mass, etc. of the vehicle.

[0066] Step S3, using a six-wheel vehicle multi-body dynamics model according to a Simulink simulation platform to obtain the dynamic stroke and dynamic stroke speed of each suspension actuator 3 during the driving process;

[0067] Step S4, inputting the dynamic stroke and dynamic stroke speed of the suspension actuator 3 into the semi-active suspension control strategy, and using the processor in the calculation to obtain the target values ​​of the equivalent damping coefficient and the equivalent stiffness coefficient of the suspension actuator 3;

[0068] Step S5, using the control system connected to the hydraulic module to control the opening and closing state of the electromagnetic switch valve in real time, and to adjust the opening of the adjustable flow valve in the damping adjustment module 8 in real time, so that the equivalent damping coefficient and the equivalent stiffness coefficient of the suspension actuator 3 reach the target values ​​of the equivalent damping coefficient and the equivalent stiffness coefficient in step S4, thereby realizing the control of the stiffness and damping of the suspension actuator 3, and allowing the six-wheeled vehicle to travel smoothly under different road conditions.

[0069] The multi-body dynamics model of the six-wheeled vehicle in step S2 is as follows: Figure 4As shown:

[0070] m u,i ×a u,i +F HS,i + F RS,i +k t (z u,i -z r,i )=0,i=1,2,3,4,5,6

[0071] m s ×a s =∑ 6 i=1 F HS,i +∑ 6 i=1 F RS,i

[0072] I xx ×a φ =(F HS,1 +F HS,3 +F HS,5 )b / 2-(F HS,2 +F HS,4 +F HS,6 )b / 2+(F RS,1 +F RS,3 +F RS,5 )b / 2-(F RS,2 +F RS,4 +F RS,6 )b / 2

[0073] I yy ×a θ =-(F HS,1 +F HS,2 ) f -(F HS,3 +F HS,4 ) m +(F HS,5 +F HS,6 ) r -(F RS,1 +F RS,2 ) f -(F RS,3 +F RS,4 ) m +(F RS,5 +F RS,6 ) r

[0074] Among them, m u,i represents the unsprung mass at the position of the i-th tire; i represents the ordinal number of the tire, and the 1st to 6th tires are the left front tire, right front tire, left middle tire, right middle tire, left rear tire and right rear tire of the six-wheel vehicle respectively; zu,i represents the vertical displacement of the unsprung mass at the i-th tire position; a u,i represents the vertical displacement z of the unsprung mass at the i-th tire position u,i Find the second-order derivative with respect to time t; F HS,i represents the vertical force of the suspension actuator 3 at the position of the i-th tire; F RS,i represents the vertical force of suspension spring 2 at the position of the i-th tire; k t Represents the equivalent stiffness coefficient of each tire; z r,i represents the road input displacement at the i-th tire position; m s Indicates the sprung mass of the vehicle; s represents the vertical displacement of the sprung mass center; a s is the vertical displacement z at the center of mass of the sprung mass s Find the second-order derivative with respect to time t; I xx I represents the moment of inertia of the sprung mass around the vehicle's x-axis; yy represents the moment of inertia of the sprung mass around the vehicle y-axis, the vehicle's x-axis is parallel to the length direction of the vehicle body 1, the vehicle's y-axis is parallel to the width direction of the vehicle body 1, and the vehicle's z-axis is perpendicular to the xoy plane; a θ represents the angular acceleration of the sprung mass at the pitch angle (θ); a φ represents the angular acceleration of the roll angle (φ) at the center of mass of the sprung mass; b represents the distance between two adjacent tires in the width direction of the six-wheeled vehicle; l f It represents the distance between the front wheel and the center of mass of the sprung mass in the length direction of the six-wheel vehicle; l m It represents the distance between the middle wheel and the center of mass of the sprung mass in the length direction of a six-wheeled vehicle; l r It represents the distance between the rear wheels and the center of mass of the sprung mass in the length direction of a six-wheeled vehicle.

[0075] In the multi-body dynamics model of a six-wheeled vehicle, the vertical force F acting on the suspension actuator 3 is HS,i , vertical force F of suspension spring 2 RS,i According to the following formula:

[0076] F HS,i =-c i (v s,i -v u,i )-k c,i (z s,i -z u,i )

[0077] F RS,i =-k b (z s,i -z u,i )

[0078] Among them, c i represents the equivalent damping coefficient of the suspension actuator 3 at the i-th tire position; z s,i represents the vertical displacement of the connection between the suspension actuator 3 and the vehicle body 1 at the position of the i-th tire; v s,i represents the vertical displacement z of the connection between the suspension actuator 3 and the vehicle body 1 at the position of the i-th tire s,i Find the first-order derivative with respect to time; v u,i represents the vertical displacement z of the unsprung mass at the i-th tire position u,i Find the first-order derivative with respect to time t; k c,i represents the equivalent stiffness coefficient of the suspension actuator 3 at the i-th tire position; k b represents the stiffness of suspension spring 2;

[0079] Wherein, the road input displacement z at the i-th tire position is r,i According to the following formula, we can get

[0080] z r,i =2πf0z r,i (t)+ 2π(G0v) 1 / 2 w(t)

[0081] Where f0 represents the cut-off frequency; z r,i (t) represents the input displacement of the road surface at the position of the i-th tire at time t; G0 represents the road roughness coefficient; v represents the vehicle speed; w(t) represents the Gaussian white noise signal with zero amplitude mean at time t.

[0082] The semi-active suspension control strategy in step S4 includes equivalent damping control and equivalent stiffness control of the suspension actuator 3. The equivalent damping control of the suspension actuator 3 includes:

[0083] If the vertical velocity v at the connection between the suspension actuator 3 and the vehicle body 1 at the i-th tire position s,i and the suspension dynamic stroke speed (v s,i -v u,i ) On the contrary, the equivalent damping coefficient c of the suspension actuator 3 at the i-th tire position is i The target value is the maximum value of the equivalent damping coefficient c max ;

[0084] If the vertical velocity v at the connection between the suspension actuator 3 and the vehicle body 1 at the i-th tire position s,i and the suspension dynamic stroke speed (v s,i -v u,i ) is the same, then the target value c of the equivalent damping coefficient of the suspension actuator 3 at the i-th tire position is i Take the minimum value of equivalent damping coefficient cmin .

[0085] The equivalent damping coefficient of the suspension actuator 3 is set by changing the opening of the adjustable throttle valve of the damping adjustment module 8. max 、c min are preset constants, c max 、c min They are respectively the maximum and minimum values ​​to which the suspension actuator 3 can be adjusted.

[0086] The equivalent stiffness control of the suspension actuator 3 includes:

[0087] If the pitch angle θ at the center of mass of the vehicle body sprung mass is greater than 0, that is, the height of the front axle of the vehicle body sprung mass is lower than the height of the rear axle, then the equivalent stiffness coefficient k of the suspension actuator 3 at the position of the first and second tires is c,1 , k c,2 The target value is the maximum value of the equivalent stiffness coefficient k cmax , the equivalent stiffness coefficient k of the suspension actuator 3 at the 5th and 6th tire positions c,5 , k c,6 The target value is the minimum value of the equivalent stiffness coefficient k cmin ;

[0088] If the pitch angle θ at the center of mass of the vehicle body sprung mass is less than 0, that is, the height of the front axle of the vehicle body sprung mass is higher than the height of the rear axle, then the equivalent stiffness coefficient k of the suspension actuator 3 at the position of the first and second tires is c,1 , k c,2 The target value is the minimum value of the equivalent stiffness coefficient k cmin , the equivalent stiffness coefficient k of the suspension actuator 3 at the 5th and 6th tire positions c,5 , k c,6 The target value is the maximum value of the equivalent stiffness coefficient k cmax ;

[0089] If the pitch angle θ at the center of mass of the vehicle body sprung mass is 0, that is, the height of the front axle of the vehicle body sprung mass is equal to the height of the rear axle, then the equivalent stiffness coefficient k of the suspension actuator 3 at the positions of the 1st, 2nd, 5th, and 6th tires is c,1 , k c,2 , k c,5 , k c,6 The target value of the equivalent stiffness coefficient is the minimum value k cmin ;

[0090] If the roll angle φ at the center of mass of the vehicle body sprung mass is greater than 0, that is, the height of the left side of the vehicle body sprung mass is higher than that of the right side, then the equivalent stiffness coefficient k of the suspension actuator 3 at the position of the third tire is c,3 The target value is the minimum value of the equivalent stiffness coefficient k cmin , the equivalent stiffness coefficient k of the suspension actuator 3 at the 4th tire position c,4The target value is the maximum value of the equivalent stiffness coefficient k cmax ;

[0091] If the roll angle φ at the center of mass of the vehicle body sprung mass is less than 0, that is, the height of the left side of the vehicle body sprung mass is lower than that of the right side, then the equivalent stiffness coefficient k of the suspension actuator 3 at the position of the third tire is c,3 The target value is the maximum value of the equivalent stiffness coefficient k cmax , the equivalent stiffness coefficient k of the suspension actuator 3 at the 4th tire position c,4 The target value is the minimum value of the equivalent stiffness coefficient k cmin ;

[0092] If the roll angle φ at the center of mass of the sprung mass of the vehicle body is 0, that is, the height of the left side of the sprung mass of the vehicle body is equal to the height of the right side, then the equivalent stiffness coefficient k of the suspension actuator 3 at the position of the 3rd and 4th tires is c,3 , k c,4 The target value of the equivalent stiffness coefficient is the minimum value k cmin .

[0093] The equivalent stiffness coefficient of the suspension actuator 3 is regulated by changing the opening and closing state of the electromagnetic switch valve. When the electromagnetic switch valve is opened, the primary accumulator and the secondary accumulator are connected. At this time, the equivalent stiffness coefficient of the suspension actuator 3 reaches the minimum value k cmin ; When the electromagnetic switch valve is closed, the primary accumulator and the secondary accumulator are not connected, and the equivalent stiffness coefficient of the suspension actuator 3 reaches the maximum value k cmax The suspension actuators 3 at the 1st to 6th tire positions are the left front, right front, left center, right center, left rear, and right rear suspension actuators 3, respectively. The equivalent stiffness coefficients of the left front, right front, left center, right center, left rear, and right rear suspension actuators 3 are k c1 , k c2、 k c3 , k c4、 k c5 , k c6 .

[0094] The expression of the semi-active suspension control strategy is as follows:

[0095] If v s,i (v s,i -v u,i )>0, then c i =c max ; if v s,i (v s,i -v u,i )<0, then c i =c min

[0096] If θ>0, then k c,1 =k c,2 =kcmax , k c,5 =k c,6 =k cmin ;

[0097] If θ<0, then k c,1 =k c,2 =k cmin , k c,5 =k c,6 = k cmax ;

[0098] If θ=0, then k c,1 =k c,2 = k c,5 =k c,6 =k cmin

[0099] If φ>0, then k c,3 = k cmin , k4= k cmax ;

[0100] If φ<0, then k c,3 = k cmax , k4= k cmin ;

[0101] If φ=0, then k c,3 =k c,4 =k cmin

[0102] The control performance of the semi-active suspension of a six-wheeled vehicle is analyzed and compared using the joint simulation platform Simulink-AMESim. The specific implementation is as follows:

[0103] Step 1: Build a multi-body dynamics model of a six-wheel vehicle in Simulink, calculate and solve the dynamic stroke and dynamic stroke speed of each suspension actuator, and transmit them to the simulation software AMESim;

[0104] Step 2: Build a hydraulic module in AMESim, calculate and solve the output force of the suspension actuator and transmit it to Simulink;

[0105] Step 3: Establish a semi-active suspension control strategy in Simulink and set simulation conditions.

[0106] Assume that the road surface grade is E and the road surface roughness coefficient is 4.096×10-3m 3 , the vehicle passes at a constant speed of 10km / h, and the vehicle body acceleration is obtained as Figure 5 As shown, the vehicle body pitch angle is Figure 6 As shown, the body roll angle is Figure 7 As shown, the suspension travel is Figure 8 As shown, the tire dynamic load is Fig. 9 As shown in Table 1, by comparing the semi-active suspension with the passive suspension, the semi-active control method proposed in the present invention reduces the root mean square value of the vehicle acceleration by 25.9%, the root mean square value of the vehicle pitch angle by 22.3%, the root mean square value of the vehicle roll angle by 13.8%, the root mean square value of the suspension travel by 22.6%, and the root mean square value of the tire dynamic load by 5.6%.

[0107] Table 1 Comparison results between semi-active suspension and passive suspension

[0108] <![CDATA[Root mean square value of vehicle body vertical acceleration (m / s 2 )]]> RMS value of vehicle pitch angle (rad) Body roll angle root mean square value (rad) Suspension dynamic travel RMS value (m) Tire dynamic load RMS value (N) Passive suspension 1.5621 0.0103 0.0123 0.0133 <![CDATA[3.331×10 3 ]]> Semi-active suspension 1.1571 0.0080 0.0106 0.0103 <![CDATA[3.518×10 3 ]]> optimization(%) 25.9 22.3 13.8 22.6 -5.6

[0109] According to the above simulation results, it can be seen that the use of the hydraulic semi-active suspension of the present invention can suppress changes in vehicle body posture and improve the driving smoothness of the six-wheel vehicle.

[0110] Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention, rather than to limit it. Although the present invention has been described in detail with reference to the aforementioned embodiments, those skilled in the art should understand that they can still modify the technical solutions described in the aforementioned embodiments, or replace some or all of the technical features therein by equivalents. However, these modifications or replacements do not make the essence of the corresponding technical solutions deviate from the scope of the technical solutions of the embodiments of the present invention.

Claims

1. A hydraulic semi-active suspension for a six-wheel vehicle, characterized in that: The vehicle comprises a vehicle body (1), six tire knuckles (6), six suspension modules and six hydraulic modules; the six tire knuckles (6) and the six suspension modules are all mounted on the vehicle body (1); a transverse arm in each suspension module is connected to each tire knuckle (6); a suspension actuator (3) in each suspension module is connected to each hydraulic module; the hydraulic module is used to control the elastic force and damping force between the suspension actuator (3) and the vehicle body (1), thereby controlling the smooth driving of the six-wheeled vehicle under different driving conditions through the suspension module; The suspension module comprises a suspension spring (2), a suspension actuator (3), an upper cross arm (4) and a lower cross arm (5); the upper cross arm (4) is located above the lower cross arm (5), the upper cross arm (4) is hinged between the middle of the vehicle body (1) and the tire knuckle (6), the lower cross arm (5) is hinged between the lower part of the vehicle body (1) and the tire knuckle (6), the retractable suspension actuator (3) is hinged between the top of the vehicle body (1) and the middle of the lower cross arm (5), the suspension spring (2) is wound around the outer side wall of the suspension actuator (3), the top end of the suspension spring (2) is hinged to the connection node between the suspension actuator (3) and the vehicle body (1), and the bottom end of the suspension spring (2) is hinged to the connection node between the suspension actuator (3) and the lower cross arm (5); The hydraulic module comprises a stiffness adjustment module (7) and a damping adjustment module (8); the rodless chamber of the suspension actuator (3) in the suspension module is connected to one end of the damping adjustment module (8) via a pipeline; the other end of the damping adjustment module (8) is connected to the stiffness adjustment module (7); the damping adjustment module (8) is used to adjust the damping force of the suspension actuator (3); the stiffness adjustment module (7) is mainly composed of a primary accumulator, a secondary accumulator and an electromagnetic switch valve; the damping adjustment module (8) is connected to the secondary accumulator of the stiffness adjustment module (7) via a pipeline; the primary accumulator and the electromagnetic switch valve are sequentially arranged on the pipeline from the damping adjustment module (8) to the secondary accumulator; the electromagnetic switch valve is used to adjust the equivalent stiffness of the suspension actuator (3); The damping adjustment module (8) and the electromagnetic switch valve are both externally connected to a control system, and the control system is used to control the opening and closing state of the electromagnetic switch valve and adjust the equivalent damping coefficient of the damping adjustment module (8); When the electromagnetic switch valve is opened, the first-stage accumulator and the second-stage accumulator are connected, and at this time, the equivalent stiffness coefficient of the suspension actuator (3) reaches a minimum value; when the electromagnetic switch valve is closed, the first-stage accumulator and the second-stage accumulator are not connected, the suspension actuator (3) is only connected to the first-stage accumulator, and the second-stage accumulator does not participate in the stiffness adjustment of the suspension actuator (3), and at this time, the equivalent stiffness coefficient of the suspension actuator (3) reaches a maximum value.

2. A six-wheel vehicle hydraulic semi-active suspension according to claim 1, characterized in that: The suspension actuator (3) adopts an asymmetric hydraulic cylinder, and the damping adjustment module (8) adopts an adjustable flow valve.

3. A control method for a six-wheeled vehicle hydraulic semi-active suspension as claimed in any one of claims 1-2, characterized in that: The following steps are involved: Step S1, firstly obtaining the geometric characteristics, speed and spring load information of the six-wheeled vehicle during driving; Step S2: Next, a multi-body dynamics model of the six-wheeled vehicle and a semi-active suspension control strategy are constructed in a computer according to the geometric characteristics, speed and spring load information of the six-wheeled vehicle; Step S3, using a six-wheel vehicle multi-body dynamics model to obtain the dynamic stroke and dynamic stroke speed of each suspension actuator (3); Step S4, inputting the dynamic stroke and dynamic stroke speed of step S3 into the semi-active suspension control strategy, and using the processor to obtain target values ​​of the equivalent damping coefficient and the equivalent stiffness coefficient of the suspension actuator (3); Step S5: Using the control system to control the opening and closing state of the electromagnetic switch valve and the opening of the adjustable flow valve in the damping adjustment module (8) in real time, so that the equivalent damping coefficient and the equivalent stiffness coefficient of the suspension actuator (3) reach the target values ​​of step S4, so that the six-wheeled vehicle can travel smoothly under different road conditions.

4. The control method of a hydraulic semi-active suspension of a six-wheel vehicle according to claim 3, characterized in that: The multi-body dynamics model of the six-wheeled vehicle in step S2 is as follows: m u,i ×a u,i +F HS,i + F RS,i +k t (z u,i - z r,i )=0,i∈{1,2,3,4,5,6} m s ×a s =∑ 6 i=1 F HS,i +∑ 6 i=1 F RS,i I xx ×a φ =(F HS,1 +F HS,3 +F HS,5 )b / 2-(F HS,2 +F HS,4 +F HS,6 )b / 2+(F RS,1 +F RS,3 +F RS,5 )b / 2-(F RS,2 +F RS,4 +F RS,6 )b / 2 I yy ×a θ =-(F HS,1 +F HS,2 )l f -(F HS,3 +F HS,4 )l m +(F HS,5 +F HS,6 )l r -(F RS,1 +F RS,2 )l f -(F RS,3 +F RS,4 )l m +(F RS,5 +F RS,6 )l r Among them, m u,i represents the unsprung mass at the position of the i-th tire; i represents the ordinal number of the tire, and the 1st to 6th tires are the left front tire, right front tire, left middle tire, right middle tire, left rear tire and right rear tire of the six-wheel vehicle respectively; z u,i represents the vertical displacement of the unsprung mass at the i-th tire position; a u,i represents the vertical displacement z of the unsprung mass at the i-th tire position u,i Find the second-order derivative with respect to time t; F HS,i represents the vertical force of the suspension actuator (3) at the position of the i-th tire; F RS,i represents the vertical force of the suspension spring (2) at the position of the i-th tire; k t Represents the equivalent stiffness coefficient of each tire; z r,i represents the road input displacement at the i-th tire position; m s Indicates the sprung mass of the vehicle; s represents the vertical displacement at the center of mass of the sprung mass; a s is the vertical displacement z at the center of mass of the sprung mass s Find the second-order derivative with respect to time t; I xx I represents the moment of inertia of the sprung mass around the vehicle's x-axis; yy represents the moment of inertia of the sprung mass about the vehicle's y-axis; a θ represents the angular acceleration of the pitch angle at the center of mass of the sprung mass; a φ represents the angular acceleration of the roll angle at the center of mass of the sprung mass; b represents the distance between the two tires in the width direction of the six-wheeled vehicle; l f It represents the distance between the front wheel and the center of mass of the sprung mass in the length direction of the six-wheel vehicle; l m It represents the distance between the middle wheel and the center of mass of the sprung mass in the length direction of a six-wheeled vehicle; l r It represents the distance between the rear wheels and the center of mass of the sprung mass in the length direction of a six-wheeled vehicle.

5. The control method of a hydraulic semi-active suspension of a six-wheel vehicle according to claim 4, characterized in that: In the multi-body dynamics model of the six-wheeled vehicle, the vertical force F of the suspension actuator (3) is HS,i , the vertical force F of the suspension spring (2) RS,i According to the following formula: F HS,i =-c i (v s,i - v u,i )-k c,i (z s,i - z u,i ) F RS,i =-k b (With s,i -With u,i ) Among them, c i represents the equivalent damping coefficient of the suspension actuator (3) at the i-th tire position; z s,i represents the vertical displacement of the connection between the suspension actuator (3) and the vehicle body (1) at the position of the i-th tire; v s,i represents the vertical displacement z of the connection between the suspension actuator (3) and the vehicle body (1) at the position of the i-th tire s,i Find the first-order derivative with respect to time; v u,i represents the vertical displacement z of the unsprung mass at the i-th tire position u,i Find the first-order derivative with respect to time t; k c,i represents the equivalent stiffness coefficient of the suspension actuator (3) at the i-th tire position; k b Indicates the stiffness of the suspension spring (2).

6. The control method of a hydraulic semi-active suspension of a six-wheel vehicle according to claim 3, characterized in that: The semi-active suspension control strategy in step S4 includes equivalent damping control and equivalent stiffness control of the suspension actuator (3), and the equivalent damping control of the suspension actuator (3) includes: If the vertical velocity v at the connection between the suspension actuator (3) and the vehicle body (1) at the i-th tire position s,i The suspension travel speed of the suspension actuator (3) at the i-th tire position is opposite to that of the suspension actuator (3), and the equivalent damping coefficient c of the suspension actuator (3) is i The target value is the maximum value of the equivalent damping coefficient c max ; If the vertical velocity v at the connection between the suspension actuator (3) and the vehicle body (1) at the i-th tire position s,i The target value c of the equivalent damping coefficient of the suspension actuator (3) is the same as the suspension travel speed of the suspension actuator (3) at the i-th tire position. i Take the minimum value of equivalent damping coefficient c min .

7. The control method of a hydraulic semi-active suspension of a six-wheel vehicle according to claim 3, characterized in that: The equivalent stiffness control of the suspension actuator (3) includes: If the pitch angle θ at the center of mass of the vehicle body sprung mass is greater than 0, the target value of the equivalent stiffness coefficient of the suspension actuator (3) at the first and second tire positions is the maximum value of the equivalent stiffness coefficient k cmax , the target value of the equivalent stiffness coefficient of the suspension actuator (3) at the 5th and 6th tire positions is the minimum value of the equivalent stiffness coefficient k cmin ; If the pitch angle θ at the center of mass of the vehicle body sprung mass is less than 0, the target value of the equivalent stiffness coefficient of the suspension actuator (3) at the first and second tire positions is the minimum value of the equivalent stiffness coefficient k cmin , the target value of the equivalent stiffness coefficient of the suspension actuator (3) at the 5th and 6th tire positions is the maximum value of the equivalent stiffness coefficient k cmax ; If the pitch angle θ at the center of mass of the sprung mass of the vehicle body is 0, the target value of the equivalent stiffness coefficient of the suspension actuator (3) at the positions of the 1st, 2nd, 5th and 6th tires is the minimum value of the equivalent stiffness coefficient k cmin ; If the roll angle φ at the center of mass of the sprung mass of the vehicle body is greater than 0, the equivalent stiffness coefficient k of the suspension actuator (3) at the third tire position is c,3 The target value is the minimum value of the equivalent stiffness coefficient k cmin , the equivalent stiffness coefficient k of the suspension actuator (3) at the 4th tire position c,4 The target value is the maximum value of the equivalent stiffness coefficient k cmax ; If the roll angle φ at the center of mass of the sprung mass of the vehicle body is less than 0, the equivalent stiffness coefficient k of the suspension actuator (3) at the third tire position is c,3 The target value is the maximum value of the equivalent stiffness coefficient k cmax , the equivalent stiffness coefficient k of the suspension actuator (3) at the 4th tire position c,4 The target value is the minimum value of the equivalent stiffness coefficient k cmin ; If the roll angle φ at the center of mass of the sprung mass of the vehicle body is 0, the target values ​​of the equivalent stiffness coefficients of the suspension actuator (3) at the positions of the third and fourth tires are both taken as the minimum value of the equivalent stiffness coefficient k cmin .

8. A computer device comprising a memory and a processor, wherein the memory stores a computer program, wherein: When the processor executes the computer program, the steps of the control method of the hydraulic semi-active suspension according to any one of claims 3 to 7 are implemented.

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