An electro-hydraulic steering system and control method with volume / throttle speed regulation mode switching

Abstract

The present application provides a kind of volume / throttle speed regulation mode switching electro-hydraulic steering system and control method, the system includes pump control system, oil return port valve control system, signal acquisition system, electronic control system;The method uses the compound control strategy of volume speed regulation / throttle speed regulation two mode switching, in volume speed regulation mode, according to the deviation signal of steering angle to servo motor feedback control, and then control system pressure, and according to the oil return port pressure control servo proportional valve to reach the set back cavity pressure;In throttle speed regulation mode, according to the target steering angle speed feedforward and valve port pressure difference feedback to servo motor control, and according to the deviation signal of steering angle to adjust the valve opening of servo proportional valve, and then control system flow;The present application can meet the high-precision dynamic steering demand of electro-hydraulic servo steering system, achieve on-demand power matching, energy saving and consumption reduction, while also avoiding the negative pressure of system cavity, improve the overall stiffness and stability of system.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of heavy vehicle electro-hydraulic servo steering, and particularly relates to an electro-hydraulic steering system and control method for switching between volume / throttle speed regulation modes. BACKGROUND

[0002] Heavy multi-axle vehicles are widely used in many fields such as ultra-long blade transfer during large wind power equipment installation, ultra-heavy precast beam transportation during large bridge construction, and efficient transfer and positioning of intercontinental missiles in the field. They are indispensable core engineering equipment in the fields of national economic construction and military heavy industry. With the deepening of the "double carbon" strategy, the green development concept of the engineering machinery industry has become a widely accepted consensus. To design heavy multi-axle vehicles with driving safety and stability, off-road flexibility, and energy efficiency, the key is the design of the multi-axle steering system. Electro-hydraulic servo steering systems have the characteristics of fast dynamic response, large output power, and wide range, and have good performance in high-precision steering and tracking, and are the preferred solution for heavy vehicle steering. Existing electro-hydraulic servo steering systems and controls, such as the load port independent control type electro-hydraulic servo steering system described in reference patent CN202210398702.4, are based on valve control systems, introduce load port independent control technology, and use integral sliding mode controllers to improve system response speed, flow matching accuracy, and energy saving effect. The closed servo pump control system described in reference patent CN202210398704.3 allows the pump output flow to be load matched according to the actual working conditions, avoiding the inevitable throttling loss of the valve control system, and further reducing the energy consumption of the steering system.

[0003] Existing technologies have achieved certain results in the efficient energy saving and high-precision control of electro-hydraulic servo steering systems, but there is still room for improvement. The main performance is as follows:

[0004] (1) The traditional centralized valve control system commonly used in electro-hydraulic servo steering systems, although mature in application, fast in response speed, and high in control accuracy, cannot accurately match the oil pressure and flow required by each axle, resulting in large overflow and throttling losses, which can easily cause a large amount of unnecessary energy loss, making the steering system inefficient and difficult to meet the current strict energy consumption limit standards. Other energy-saving measures, such as load port independent control technology, can avoid overflow loss and reduce some throttling loss, but due to the throttling speed regulation principle of the valve port, high-precision control of the steering angle must be achieved at the expense of throttling loss, which has great limitations in the development of green energy-saving heavy vehicles.

[0005] (2) Compared with the traditional valve control system, the variable speed pump control system has the advantages of good dynamic response and low energy consumption, but due to the changeable all-terrain driving conditions of heavy vehicles, unknown external disturbances and wide load fluctuation range, etc. Factors, directly applying it to the steering system of heavy vehicles, the system cavity will produce negative pressure, the system has the problems of insufficient driving stiffness, poor stability, etc., which leads to weak anti-interference performance, which brings great difficulty to high-precision steering control.

[0006] (3) In the working process of the open variable speed pump control electro-hydraulic steering system, when the wheel steering changes, the electromagnetic reversing valve core switches position, at this time the system oil return port pressure gradually decreases, in the process of pressure drop, the pressure value fluctuates greatly, the buffeting phenomenon is obvious, which seriously affects the stability of the multi-axle heavy vehicle under the steering working condition. SUMMARY

[0007] The present application provides a volume / throttle speed regulation mode switching electro-hydraulic steering system and control method, which can ensure high-precision dynamic steering, low-energy consumption driving of heavy vehicle steering system, while avoiding negative pressure in the cavity, meeting the driving stiffness required by the system and improving the stability of the steering system.

[0008] The present application adopts the following technical solutions.

[0009] A volume / throttle speed regulation mode switching electro-hydraulic steering system for vehicle steering, comprising an open variable speed pump control system, an oil return port valve control system, a signal acquisition system and an electronic control system; the open variable speed pump control system controls the steering assist cylinders of the left and right wheels through the servo motor pump unit and the electromagnetic reversing valve; the oil return port valve control system is used to adjust the back cavity pressure of the oil return port and control the flow of the system; the signal acquisition system is used to acquire the steering angle of the steering wheel and the pressure of each specified circuit position; the electronic control system outputs control instructions to control the steering system of the vehicle in real time according to the target angle signal, the wheel angle sensor signal and the pressure sensor signal of each specified circuit position.

[0010] The open variable speed pump control system comprises a servo motor (2), a one-way constant displacement pump (20), an electromagnetic reversing valve (5), a first hydraulic control one-way valve (6), a second hydraulic control one-way valve (14), a first overflow valve (3), a second overflow valve (7), a third overflow valve (13), a fourth overflow valve (17), a first electromagnetic reversing ball valve (15) and an oil tank (1);

[0011] The oil return port valve control system comprises a servo proportional valve, and the oil return port valve control system adjusts the back cavity pressure of the oil return port by controlling the displacement of the servo proportional valve core;

[0012] The signal acquisition system comprises a first rotation angle sensor, a first pressure sensor, a second pressure sensor and a third pressure sensor.

[0013] The electronic control system comprises a controller and a servo motor driver, which calculates and outputs corresponding control voltage signals according to the target rotation angle required by the wheel, the actual rotation angle of the steering wheel and the pressure of the pressure sensor of each specified loop, and switches between the pump control volume speed regulation mode and the pump-valve combined control throttling speed regulation mode according to the actual system loop pressure, so as to control the electro-hydraulic steering system.

[0014] The servo motor and the one-way constant-displacement pump are connected in combination through a connecting key to form a servo motor pump, and the servo motor is electrically connected with the servo driver (18) to change the rotation speed of the motor through a voltage control signal, so as to adjust the output flow and pressure of the motor pump.

[0015] The rodless cavity of the left steering assist cylinder (9) and the rod cavity of the right steering assist cylinder (10) of the vehicle steering system are connected with the B port of the first hydraulic control one-way valve, and the A port of the first hydraulic control one-way valve and the A port of the first electromagnetic reversing valve are connected to form the first working loop.

[0016] The rod cavity of the left steering assist cylinder and the rodless cavity of the right steering assist cylinder are connected with the B port of the second hydraulic control one-way valve, and the A port of the second hydraulic control one-way valve and the B port of the first electromagnetic reversing valve are connected to form the second working loop.

[0017] The B port of the first hydraulic control one-way valve is connected with the A port of the second overflow valve, and the T port of the second overflow valve is connected with the oil tank; the B port of the second hydraulic control one-way valve is connected with the A port of the third overflow valve, and the T port of the third overflow valve is connected with the oil tank; the P port of the first electromagnetic reversing valve is connected with the outlet of the constant-displacement pump, and the A port of the first overflow valve is connected in parallel, and the T port of the first overflow valve is connected with the oil tank.

[0018] The T port of the first electromagnetic reversing valve is connected with the P port of the servo proportional valve (19), the A port of the servo proportional valve is connected with the oil tank, and the T port and the B port of the servo proportional valve are closed with plugs.

[0019] The K ports of the first hydraulic control one-way valve and the second hydraulic control one-way valve are connected with the A port of the first electromagnetic reversing ball valve, the P port of the first electromagnetic reversing ball valve is connected with the P port of the first electromagnetic reversing valve, and the T port of the first electromagnetic reversing ball valve is connected with the oil tank.

[0020] In the hydraulic system of the open variable-speed pump control system, the first hydraulic control one-way valve, the second hydraulic control one-way valve and the first electromagnetic reversing ball valve form a hydraulic lock loop; the protection effect of the first overflow valve, the second overflow valve, the third overflow valve and the fourth overflow valve on the loop prevents the system from running in danger due to excessive loop pressure, and when the pressure of each node of the system exceeds the protection pressure set by each overflow valve, the excess pressure part of the corresponding oil will flow back to the oil tank through the overflow valve.

[0021] A control method of an electro-hydraulic steering system with volume / throttle speed regulation mode switching, the control method adopts a compound control strategy with volume speed regulation / throttle speed regulation mode switching, in the volume speed regulation mode, the servo motor is feedback controlled according to the turning angle deviation signal of the wheel steering angle, to control the system pressure, and the servo proportional valve is controlled according to the oil return port pressure to achieve the set back cavity pressure; in the throttle speed regulation mode, the servo motor is controlled according to the target turning angle speed feedforward and the valve port pressure difference feedback, and the servo proportional valve opening is adjusted according to the turning angle deviation signal, to control the system flow, to meet the high-precision dynamic steering requirements of the electro-hydraulic servo steering system, and to achieve on-demand power supply, energy saving and consumption reduction, to avoid negative pressure in the system cavity, and to improve the overall stiffness and stability of the system.

[0022] The control method comprises the following steps:

[0023] Step S1: system startup, establishing an electro-hydraulic steering system mathematical model and control guide model with volume / throttle speed regulation mode switching in the controller, inputting the wheel target turning angle signal and the target back cavity pressure signal in the volume speed regulation mode to the controller, inputting the analog signals collected by the right wheel angle sensor and the pressure sensor to the controller, and calculating the pressure difference between the servo proportional valve inlet and outlet, and the deviation between the right wheel actual turning angle and the target turning angle;

[0024] Step S2: starting the volume speed regulation mode, taking the servo motor pump as the main control object, according to the deviation signal of the right wheel target turning angle and the actual turning angle, the controller sends signals to the servo motor and the electromagnetic reversing valve to control their work, to control the variable speed pump output pressure, to form the hydraulic driving force, to make the left and right steering cylinders extend and retract and make the wheel reach the target turning angle;

[0025] Step S3: according to the deviation signal of the system oil return port back cavity pressure and the expected back cavity pressure, the controller outputs a suitable voltage signal to drive the servo proportional valve to work, to dynamically adjust the pressure at the oil return port to reach the set back cavity pressure;

[0026] Step S4: judging whether the electromagnetic reversing valve spool switches position, if yes, jumping to step S5, if not, continuing to execute steps S2-S3;

[0027] Step S5: starting the throttle speed regulation mode, taking the oil return port servo proportional valve as the main control object, calculating the angular velocity according to the right wheel target turning angle signal, controlling the servo motor pump through angular velocity feedforward and servo proportional valve port pressure difference feedback, to make the motor pump output the expected flow;

[0028] Step S6: According to the deviation signal of the right wheel target angle and the actual angle, the appropriate voltage value is calculated in the controller, the servo proportional valve spool is driven to produce corresponding displacement, the valve opening size of the servo proportional valve is adjusted, and then the system flow is controlled, so that the left and right steering cylinders are extended and retracted and the wheels reach the target angle;

[0029] Step S7: Determine whether the system oil return port pressure reaches the set back cavity pressure required by the system at this time. If yes, jump to step S2, if not, continue to execute steps S5-S6;

[0030] Step S8: When the controller receives the stop command, the electromagnetic reversing valve and the servo proportional valve are in the middle position, and the servo motor is stopped.

[0031] The target back cavity pressure value in the volume speed regulation mode is obtained as follows: in the hydraulic control system, the liquid guide is a measure of the valve opening under the predetermined flow coefficient and oil density. The larger the valve opening, the larger the flow capacity. Therefore, from the perspective of the change of the liquid guide characteristic, the selection range of the back cavity pressure is analyzed, and the relationship between the liquid guides is as follows:

[0032]

[0033]

[0034] In formula one, A is the area of the rodless cavity, a is the area of the rod cavity, p1 and p2 are the pressures of the two cavities, v is the piston speed of the hydraulic cylinder, p s is the pump source pressure, K1 and K2 are the liquid guides of the inlet and outlet valves, and the tank pressure is 0;

[0035] In order to have a real solution of formula two, the change range of the system liquid guide is R 1.5 K e <K1<∞,K e <K2<∞, and the liquid guide equivalent K e is proportional to the piston speed v of the hydraulic cylinder. When the target angle signal is determined, the hydraulic cylinder speed v is obtained, the liquid guide curve is calculated, and the selection range of the back cavity pressure is obtained.

[0036] When starting the volume speed regulation mode, the control signal sent by the controller to the servo motor is obtained as follows:

[0037]

[0038] In formula three, u m1 is the servo motor control signal, g and f are the simplified expressions of the system variable polynomial, d can be regarded as the total disturbance of the system, s m1 is the sliding mode function related to the angle tracking error, and ξ m1is the function characteristic coefficient, -ε m11 s m1 -ε m12 sgn(s m1 ) is the switching law and the exponential approach law.

[0039] The control signal sent by the controller to the servo proportional valve is obtained as follows:

[0040]

[0041] In formula four, u v is the servo proportional valve control signal, c and h are the simplified expression forms of the system variable polynomials, is the derivative of the actual displacement of the servo proportional valve spool, s2 is the sliding mode function related to the back cavity pressure deviation, -ε3s2-ε4sgn(s2) is the switching law and the exponential approach law, and b is the oil compression coefficient.

[0042] The method for obtaining the signal for judging whether the electromagnetic reversing valve spool switches positions and whether the speed regulation mode is switched is as follows:

[0043] The left wheel turning is set as positive, the deviation angle between the target turning angle and the actual turning angle of the right steering wheel is taken as the feedback signal, and the positive and negative values of the feedback signal are taken as the control signal of the electromagnetic reversing valve switching, and the control signal function is as follows:

[0044]

[0045] In formula five, u z is the electromagnetic reversing valve control signal, θ q is the turning angle error, θ s is the actual turning angle, k is the set switching signal, and u z =0 when the system stops or the power steering cylinder is locked.

[0046] After the direction of the wheel turning angle changes once, that is, the electromagnetic reversing valve receives the switching voltage signal of the opposite value, the controller switches the volume speed regulation mode to the throttle speed regulation mode, and when the return port pressure drops to the target back cavity pressure, the controller switches the throttle speed regulation mode to the volume speed regulation mode; in the next turning angle period, the switching control method is repeated.

[0047] When the throttle speed regulation mode is started, the control signal sent by the controller to the servo motor is obtained as follows:

[0048]

[0049] In formula six, u m2 is the servo motor control signal, H1, F1, and G1 are the simplified expression forms of the system variable polynomials, s m2ξ m2 is a function characteristic coefficient, -ε m21 sgn(s m2 )-ε m22 s m2 is a switching law and an exponential reaching law.

[0050] The control signal sent by the controller to the servo proportional valve is obtained as follows:

[0051]

[0052] In formula seven, u v is a servo proportional valve control signal, G2, F2 are simplified expressions of system variable polynomials, ξ v2 is a function characteristic coefficient, s v2 is a sliding mode function related to the right wheel angle deviation, -ε v21 sgn(s v2 )-ε v22 s v2 is a switching law and an exponential reaching law. In step S1, a kind of volume / throttle speed regulating mode switching electro-hydraulic steering system mathematical model is established in the controller, and the method is as follows:

[0053] Step S1-1: The mechanical trapezoidal steering structure modeling of the electro-hydraulic steering bench described in the present example is as follows:

[0054] For the electro-hydraulic servo steering system controlled by the electromagnetic reversing valve to drive the double steering cylinder to drive the wheel rotation, the relationship between the left and right tire steering angles is represented as:

[0055]

[0056] In formula (1), α and β are the steering angles of the left and right wheels respectively, m is the length of the two sides of the steering arm, γ is the included angle between the steering arm and the axle, L is the distance between the two main pins of the axle, and B is the distance between the two main pins of the axle.

[0057]

[0058] In formula (2), F L is the thrust of the left steering cylinder; F R is the thrust of the right steering cylinder; p1 is the pressure of the left steering cylinder rodless cavity and the right steering cylinder rod cavity; p2 is the pressure of the left steering cylinder rod cavity and the right steering cylinder rodless cavity; A is the piston area of the cylinder rodless cavity; a is the piston area of the cylinder rod cavity.

[0059] Based on the Lagrange equation, the dynamics equation of the steering trapezoidal mechanism is established:

[0060]

[0061] In the formula:

[0062]

[0063]

[0064] In formula (3-5): T is the generalized kinetic energy of the steering system; D is the generalized dissipation energy of the steering system; J L is the equivalent moment of inertia of the left wheel, knuckle and steering trapezoidal arm around the left kingpin; J R is the equivalent moment of inertia of the right wheel, knuckle and steering trapezoidal arm around the left kingpin; C L , C R are the equivalent damping coefficients of the left and right tires and their related components, respectively; Q is the generalized force of the steering system, and

[0065]

[0066]

[0067] In formula (6-9): v L , v R are the speeds of the action points of the left and right steering assist cylinders on the steering trapezoidal arms, respectively, n is the distance between the action point of the steering cylinder and the kingpin; s1 is the distance from the assist cylinder to the bracket hinge on the bracket crossbeam, s2 is the distance from the assist cylinder to the bracket hinge on the knuckle; θ3', θ3 are the angles between the left and right steering assist cylinder forces and the action point speeds, respectively; T L , T R are the left and right tire resistance moments, respectively;

[0068] Through further calculation of each term of the dynamic equation, the following equations are obtained:

[0069]

[0070]

[0071]

[0072]

[0073]

[0074] To simplify the system model, the tank pressure is set to zero, and the distance between the valve and the assist cylinder is ignored;

[0075] Step S1-2: modeling the hydraulic system of the electro-hydraulic steering bench, the method is specifically:

[0076] Step S1-2-1: Establish the quantitative pump output flow equation as:

[0077] Set the system left turn as positive motion, at this time the reversing valve signal u s >0; right turn u s <0; stop u s =0; For ease of modeling, define ω1, ω2 to distinguish the left and right turning states of the steering system:

[0078] The servo motor pump used is a constant displacement-variable speed pump, and the output flow is related to the speed and pressure. The flow equation is as follows:

[0079]

[0080] The servo motor pump used is a constant displacement-variable speed pump, and the output flow is related to the speed and pressure. The flow equation is as follows:

[0081] Q p =D p K m u m -C i (p s -p T )-C e p s (16)

[0082] In equation (16), Q m is the output flow of the quantitative pump; D p is the displacement of the gear pump, and D p =1.6×10 -5 m 3 / r; K m is the voltage-speed conversion coefficient, and K m =200 / 60, i.e. 1V corresponds to 200r / min; u m is the input voltage of the servo motor; C i , C e are the internal and external leakage coefficients of the servo motor pump; the tank pressure p T =0, let C p =C i +C e , then:

[0083] Q p =D p K m u m -C p (ω1p1+ω2p2) (17)

[0084] In equation (17), p1 and p2 are the working pressures on the left and right sides of the reversing valve port, respectively;

[0085] Step S1-2-2: Establish the hydraulic assist cylinder flow equation:

[0086] According to the flow continuity equation, the inlet and outlet flow of the double steering assist cylinder is:

[0087]

[0088]

[0089] In formula (18-19), β e represents the effective bulk modulus; A is the rodless cavity area of the assist steering cylinder, and a is the rod cavity area of the assist steering cylinder; and represent the velocities of the left and right assist cylinders; C ip and C ep represent the internal and external leakage coefficients of the assist cylinder; V t is the total volume of the steering assist cylinder;

[0090] Step S1-2-3: Electromagnetic reversing valve, servo proportional valve port flow equation:

[0091] Assuming that the electromagnetic reversing valve port throttling loss is small, in order to reduce the complexity of the model, it is assumed that the hydraulic cylinder back cavity pressure chamber is directly connected to the servo proportional valve; the servo proportional valve port flow equation is:

[0092]

[0093] In formula (20), C d is the valve port flow coefficient of the servo proportional valve; w is the valve port area gradient of the servo proportional valve; ρ is the oil density; Δp is the pressure difference before and after the valve port of the servo proportional valve; therefore, the inlet and outlet flow of the steering assist cylinder is:

[0094]

[0095] In formula (21), q1 is the A port flow of the first reversing valve; q2 is the B port flow of the first reversing valve; x v is the servo proportional valve spool displacement;

[0096] Step S1-2-4: The steering assist cylinder two-cavity working pressure dynamic equation can be obtained by simultaneous equations:

[0097]

[0098] The target back cavity pressure of the volume speed regulation mode of the input controller is:

[0099] The hydraulic guide in the hydraulic control system can be used as a measure of the valve port opening degree under certain flow coefficient and oil density. The greater the hydraulic guide, the greater the valve port opening degree and the stronger the flow capacity. Therefore, the selection range of the back cavity pressure is analyzed from the perspective of the change of the hydraulic guide characteristics, and the relationship between the hydraulic guides is as follows:

[0100]

[0101]

[0102] In formula (23), A is the area of the rodless cavity, a is the area of the rod cavity, p1 and p2 are the pressures of the two cavities respectively, v is the speed of the hydraulic cylinder piston, p s is the pump source pressure, K1 and K2 are the hydraulic guides of the inlet and outlet oil valves respectively, and the tank pressure is 0.

[0103] In order to have a real solution of formula (24), the change range of the system hydraulic guide is R 1.5 K e <K1<∞,K e <K2<∞, and the hydraulic guide equivalent K e is proportional to the speed v of the hydraulic cylinder piston. When the target rotation angle signal is determined, the speed v of the hydraulic cylinder can be obtained, the hydraulic guide curve can be calculated, and the selection range of the back cavity pressure can be obtained.

[0104] In the steps S2-S3, the volume speed regulation mode is started, and the controller sends control signals to the servo motor and the servo proportional valve respectively. The specific method is as follows:

[0105] Step S2 / 3-1: The mathematical model of the electro-hydraulic servo steering system is converted into a state space model, and the corresponding system state space equation is as follows:

[0106] The system state variables are defined as Each state is: the right wheel rotation angle β, the right wheel rotation angle speed The back cavity pressure p t , the system input is the servo motor voltage u m1 , and the servo proportional valve voltage output system is the right wheel rotation angle back cavity pressure system state space equation as follows:

[0107]

[0108] The input variable y1=x1 is defined, and the third order derivative is obtained as follows:

[0109]

[0110] In formula (26):

[0111]

[0112]

[0113]

[0114] Define input variable y2=x3, the second derivative is:

[0115]

[0116] In equation (30):

[0117]

[0118]

[0119]

[0120] Step S2 / 3-2: define the corner control error e m1 is:

[0121] e m1 =y1-y 1d (34)

[0122] Construct a sliding mode function s related to the corner tracking error m1 , s m1 =0 is a sliding surface, by designing conditions to make the state converge to the sliding surface s m1 =0, then it can be guaranteed that the corner tracking error e1 also converges to 0, the sliding surface is as follows:

[0123]

[0124] In equation (35), n is the order of the system, according to the corner control input-output linearization expression, n=3, so:

[0125]

[0126] In equation (36), λ>0, the derivative of the sliding surface is obtained:

[0127]

[0128] Step S3 / 4-3: according to the motor voltage-corner input-output relationship, establish the relationship between the sliding surface s m1 and the controller input u m1 :

[0129]

[0130] In equation (38),

[0131] In order to make the control continuous The control law is obtained from the formula:

[0132]

[0133] In order to make the state converge to the sliding surface s m1 = 0 The switching law is added to the sliding mode dynamics, and the control law is obtained as:

[0134]

[0135] In the control process, in order to ensure the approximation speed and suppress the sliding mode chattering phenomenon, a larger ε m12 and a smaller ε m11 should be taken The improved servo motor control law is obtained as:

[0136]

[0137] In formula (41), u m1 is the servo motor control signal, g and f are the simplified expressions of the system variable polynomial, d can be regarded as the total disturbance of the system, s m1 is the sliding mode function related to the rotation angle tracking error, ξ m1 is the function characteristic coefficient, -ε m11 s m1 -ε m12 sgn(s m1 ) is the switching law and exponential reaching law

[0138] Step S2 / 3-4: Design the back cavity pressure sliding mode control law, define the back cavity pressure control error e v1 as:

[0139] e v1 = y2-y 2d (42)

[0140] Define the sliding mode function s v1 , according to the input-output linearization expression of the back cavity pressure control, n = 1, so s v1 = e v1 ;

[0141] The derivative of the sliding mode function is:

[0142]

[0143] Similarly, select the switching law and exponential reaching law:

[0144]

[0145] The servo proportional valve control law is obtained as:

[0146]

[0147] In formula (45), u v is a servo proportional valve control signal, c and h are simplified expressions of system variable polynomials, is a derivative of an actual displacement of a servo proportional valve spool, s2 is a sliding mode function related to a back cavity pressure deviation, -ε3s2-ε4sgn(s2) is a switching law and an exponential approach law, and b is an oil compression coefficient;

[0148] In step S4, a signal for judging whether the electromagnetic reversing valve spool switches position and whether the speed regulation mode switches is:

[0149] The left wheel turning is set as positive, the deviation angle between the right turning wheel target turning angle and the actual turning angle is taken as a feedback signal, and the positive and negative values of the feedback signal are taken as the electromagnetic reversing valve switching control signal. The control signal function is as follows:

[0150]

[0151] In formula (46), u z is an electromagnetic reversing valve control signal, θ q is a turning angle error, θ s is an actual turning angle, k is a set switching signal, and u z =0 when the system stops or the power steering cylinder is locked.

[0152] After the wheel turning direction changes once, that is, the electromagnetic reversing valve receives the switching voltage signal of the opposite value, the controller switches the volume speed regulation mode to the throttle speed regulation mode. Until the return port pressure drops to the target back cavity pressure, the controller switches the throttle speed regulation mode to the volume speed regulation mode. In the next turning period, the switching control method is repeated.

[0153] In steps S5-S6, the throttle speed regulation mode is started, and the control signals sent by the controller to the servo motor and the servo proportional valve are:

[0154] Step S5 / 6-1: The mathematical model of the electro-hydraulic servo steering system is converted into a state space model, and the corresponding system state space equation is as follows:

[0155] X=[x1,x2,x3,x4,x5] T The system state variables are defined, which include: the right wheel turning angle x1=β, the right wheel turning speed The servo proportional valve inlet valve port pressure x3=p1, the servo proportional valve outlet valve port pressure x4=p2, and the pump source pressure x5=p s The following state space equation is obtained:

[0156]

[0157] The system output variables are defined as y1=x1 and y2=Ap, and the input u is obtained by taking the derivatives of both m2 , u v2 and the outputs y1 and y2:

[0158]

[0159]

[0160]

[0161]

[0162] The relationship between the input and output of the rotation control system is obtained by the above equations (48-51):

[0163]

[0164] For y2=Ap, the oil inlet valve port pressure difference Ap has two states of left rotation and right rotation:

[0165]

[0166] The equation can be expressed by the state variable as:

[0167]

[0168] The following equations are obtained by taking the derivative of y2 under the conditions of left rotation and right rotation:

[0169] Left rotation condition:

[0170]

[0171] Right rotation condition:

[0172]

[0173] F2, G2, and H2 are simplified expressions of polynomials containing system variables, and the relationship between the input and output of the valve port pressure difference control system is obtained by the above process:

[0174]

[0175] Step S5 / 6-2: The sliding surface of the rotation control system designed by the control guide model is:

[0176]

[0177] In equation (58), λ1>0, e v2 =y1-y1d is the tracking error of the steering angle;

[0178] Taking the derivative of the above equation, the dynamic of the sliding surface is obtained as:

[0179]

[0180] In equation (59),

[0181] In order to satisfy the condition that the sliding surface s v2 = 0, i.e. And in order to further improve the dynamic performance of the system in the reaching motion stage, the switching law and the exponential reaching law are added to the dynamic of the sliding surface respectively:

[0182]

[0183] By solving equations (59-60) simultaneously, F2+G2u v2 + ξ v2 = -ε v21 sgn(s v2 )-ε v22 s v2 The sliding mode control law of the steering angle tracking system is finally obtained as:

[0184]

[0185] Step S5 / 6-3: The sliding surface of the valve port pressure difference control system is:

[0186] s m2 = e m2 (62)

[0187] In equation (63), e m2 = y2-y 2d is the difference between the actual value and the expected value of the valve port pressure difference;

[0188] Taking the derivative of the above equation, the dynamic of the sliding surface is obtained as:

[0189]

[0190] In equation (63),

[0191] Similarly, the switching law and the exponential reaching law are added to the dynamic of the sliding surface:

[0192]

[0193] By solving equations (63-64) simultaneously, F1+G1u v2 +H1u m2 + ξ m2 = -ε m21sgn(s m2 )-ε m22 s m2 Finally, the sliding mode control law of the valve port differential pressure control system can be obtained as follows:

[0194]

[0195] Compared with existing technologies, the present invention has the following advantages:

[0196] (1) In response to the problems of large energy loss and low energy efficiency of traditional centralized valve control systems, and to avoid the commutation shock caused by the frequent forward and reverse rotation of the servo motor in closed pump control, a variable speed pump control system with adjustable pump source pressure and on-demand supply is proposed. The overall structure is relatively simple, the heat dissipation effect is good, and the energy saving effect is obvious. Starting from the power supply matching point, the loss is controlled from the source, and there is no need to consider the system flow compensation problem, which reduces the sensitivity of the system to oil impurities. In addition, the servo motor runs in one direction and there is no commutation shock problem, so the life of the servo motor is guaranteed.

[0197] (2) To address the limitation of low drive stiffness in the pump-controlled steering system of heavy vehicles, a back pressure valve is designed to be installed at the return port of the steering system. The optimal back pressure parameter selection principle and a two-input two-output nonlinear sliding mode control method are proposed. The back pressure is controlled by adjusting the displacement of the servo proportional valve core through the pressure feedback of the return port of the power steering cylinder. According to the actual working conditions, a reasonable back pressure value is selected so that the system can maintain excellent dynamic characteristics with extremely low energy consumption, effectively improve the robustness and control accuracy of the steering control system, dynamically compensate for the negative pressure of the system, and increase the system damping and drive stiffness.

[0198] (3) To address the issues of large pressure fluctuations and significant vibration during the pressure drop at the return port of the steering system, a novel steering system and control scheme based on switching between pump-controlled volumetric speed regulation and return port throttling speed regulation modes is proposed. By employing a two-mode switching strategy of volumetric speed regulation and throttling speed regulation, the large fluctuations during the pressure drop at the return port are avoided, the significant pressure vibration is eliminated, and the stability of the steering system is greatly improved. Attached Figure Description

[0199] The present invention will be further described in detail below with reference to the accompanying drawings and specific embodiments:

[0200] Appendix Figure 1 This is a schematic diagram of the electro-hydraulic steering system with volumetric speed regulation / throttling speed regulation mode switching as described in this invention;

[0201] Appendix Figure 2 This is a schematic diagram of the control method for switching between volumetric speed regulation and throttling speed regulation modes as described in this invention.

[0202] AppendixFigure 3 is a control method flowchart of the volumetric speed regulation mode described in the present application;

[0203] attached Figure 4 is a control method flowchart of the throttling speed regulation mode described in the present application;

[0204] In the figure: 1. oil tank, 2. servo motor, 3. first overflow valve, 4. first pressure sensor, 5. electromagnetic reversing valve, 6. first hydraulic control check valve, 7. second overflow valve, 8. second pressure sensor, 9. left steering assist cylinder, 10. right steering assist cylinder, 11. first angle sensor, 12. third pressure sensor, 13. third overflow valve, 14. second hydraulic control check valve, 15. first electromagnetic reversing ball valve, 16. controller, 17. fourth overflow valve, 18. servo driver, 19. servo proportional valve, 20. one-way constant displacement pump. DETAILED DESCRIPTION

[0205] As shown in the figure, an electro-hydraulic steering system for switching between volumetric and throttling speed regulation modes is used for vehicle steering, comprising an open variable speed pump control system, a return port valve control system, a signal acquisition system, and an electronic control system; the open variable speed pump control system controls the steering assist cylinders of the left and right wheels through a servo motor pump unit and an electromagnetic reversing valve; the return port valve control system is used to adjust the back cavity pressure of the return port and control the flow of the system; the signal acquisition system is used to acquire the steering angle of the steering wheel and the position pressure of each designated circuit; the electronic control system outputs control instructions to control the steering system of the vehicle in real time according to the target angle signal, the wheel angle sensor signal, and the pressure sensor signal of each designated circuit.

[0206] The open variable speed pump control system comprises a servo motor 2, a one-way constant displacement pump 20, an electromagnetic reversing valve 5, a first hydraulic control check valve 6, a second hydraulic control check valve 14, a first overflow valve 3, a second overflow valve 7, a third overflow valve 13, a fourth overflow valve 17, a first electromagnetic reversing ball valve 15, and an oil tank 1.

[0207] The return port valve control system comprises a servo proportional valve, which adjusts the back cavity pressure of the return port by controlling the displacement of the servo proportional valve spool; the signal acquisition system comprises a first angle sensor, a first pressure sensor, a second pressure sensor, and a third pressure sensor.

[0208] The electronic control system comprises a controller and a servo motor driver, which calculates and outputs corresponding control voltage signals according to the target angle required by the wheels, the actual steering angle of the steering wheel, and the pressure sensor pressure of each designated circuit position, and switches between the pump control volumetric speed regulation mode and the pump-valve combined throttling speed regulation mode according to the actual system circuit pressure, thereby controlling the electro-hydraulic steering system.

[0209] The servo motor and the one-way constant displacement pump are combined into a servo motor pump through a connecting key, and the servo motor is electrically connected with the servo driver 18, so as to change the motor speed through a voltage control signal, thereby adjusting the output flow and pressure of the motor pump;

[0210] The rodless cavity of the left steering assist cylinder 9 and the rod cavity of the right steering assist cylinder 10 of the vehicle steering system are connected with the B port of the first hydraulic control check valve, the A port of the first hydraulic control check valve and the A port of the first electromagnetic reversing valve are connected, forming a first working circuit;

[0211] The rod cavity of the left steering assist cylinder and the rodless cavity of the right steering assist cylinder are connected with the B port of the second hydraulic control check valve, the A port of the second hydraulic control check valve and the B port of the first electromagnetic reversing valve are connected, forming a second working circuit;

[0212] The B port of the first hydraulic control check valve is connected with the A port of the second overflow valve, the T port of the second overflow valve is connected with the oil tank; the B port of the second hydraulic control check valve is connected with the A port of the third overflow valve, the T port of the third overflow valve is connected with the oil tank; the P port of the first electromagnetic reversing valve is connected with the outlet of the constant displacement pump, and the A port of the first overflow valve is connected in parallel, and the T port of the first overflow valve is connected with the oil tank;

[0213] The T port of the first electromagnetic reversing valve is connected with the P port of the servo proportional valve 19, the A port of the servo proportional valve is connected with the oil tank, and the T port and the B port of the servo proportional valve are closed with plugs;

[0214] The K port of the first hydraulic control check valve and the K port of the second hydraulic control check valve are connected with the A port of the first electromagnetic reversing ball valve, the P port of the first electromagnetic reversing ball valve is connected with the P port of the first electromagnetic reversing valve, and the T port of the first electromagnetic reversing ball valve is connected with the oil tank.

[0215] In the hydraulic system of the open variable speed pump control system, the first hydraulic control check valve, the second hydraulic control check valve and the first electromagnetic reversing ball valve form a hydraulic lock circuit; through the protection of the first overflow valve, the second overflow valve, the third overflow valve and the fourth overflow valve to the circuit, the risk of overlarge circuit pressure to the system operation is prevented, when the pressure of each node of the system exceeds the protection pressure set by each overflow valve, the corresponding oil of the excess pressure part will flow back to the oil tank through the overflow valve.

[0216] The application discloses a control method of an electro-hydraulic steering system with volume / throttle speed regulation mode switching, which adopts a compound control strategy with volume speed regulation / throttle speed regulation mode switching. In the volume speed regulation mode, a servo motor is controlled according to a rotation angle deviation signal of a wheel steering angle, so as to control system pressure, and a servo proportional valve is controlled according to an oil return port pressure, so as to reach a set back cavity pressure. In the throttle speed regulation mode, the servo motor is controlled according to a target rotation speed feedforward and a valve port pressure difference feedback, and a valve port opening of the servo proportional valve is adjusted according to a rotation angle deviation signal, so as to control system flow, so as to meet high-precision dynamic steering requirements of the electro-hydraulic servo steering system, to realize on-demand power supply and energy saving and consumption reduction, to avoid negative pressure in a system cavity, and to improve overall rigidity and stability of the system.

[0217] The control method comprises the following steps.

[0218] Step S1: system starting, establishing an electro-hydraulic steering system mathematical model and a control guide model with volume / throttle speed regulation mode switching in a controller, inputting a wheel target rotation angle signal and a target back cavity pressure signal in a volume speed regulation mode into the controller, inputting analog signals collected by a right wheel angle sensor and a pressure sensor into the controller, and calculating a pressure difference between inlet and outlet of a servo proportional valve and a deviation between a right wheel actual rotation angle and a target rotation angle.

[0219] Step S2: starting the volume speed regulation mode, taking a servo motor pump as a main control object, and controlling the servo motor and an electromagnetic reversing valve according to a deviation signal of a right wheel target rotation angle and an actual rotation angle, so as to control variable speed pump output pressure, form a hydraulic driving force, and make left and right steering assist cylinders extend and retract and a wheel reach a target rotation angle.

[0220] Step S3: outputting a suitable voltage signal to drive the servo proportional valve to work according to a deviation signal of a system oil return port back cavity pressure and an expected back cavity pressure, and dynamically adjusting the pressure at the oil return port to reach the set back cavity pressure.

[0221] Step S4: judging whether a valve core of the electromagnetic reversing valve is switched, and if yes, jumping to step S5, and if not, continuing to execute steps S2-S3.

[0222] Step S5: starting the throttle speed regulation mode, taking an oil return port servo proportional valve as a main control object, calculating an angular velocity according to a right wheel target rotation angle signal, controlling a servo motor pump through angular velocity feedforward and servo proportional valve port pressure difference feedback, and making the motor pump output an expected flow.

[0223] Step S6: According to the deviation signal of the right wheel target angle and the actual angle, the appropriate voltage value is calculated in the controller, the servo proportional valve spool is driven to produce corresponding displacement, the valve opening size of the servo proportional valve is adjusted, and then the system flow is controlled, so that the left and right steering cylinders are extended and retracted and the wheels reach the target angle;

[0224] Step S7: Determine whether the system oil return port pressure reaches the set back cavity pressure required by the system at this time. If yes, jump to step S2, if not, continue to execute steps S5-S6;

[0225] Step S8: When the controller receives the stop command, the electromagnetic reversing valve and the servo proportional valve are in the middle position, and the servo motor is stopped.

[0226] The target back cavity pressure value in the volume speed regulation mode is obtained as follows: in the hydraulic control system, the liquid guide is a measure of the valve opening under the predetermined flow coefficient and oil density. The larger the valve opening, the larger the flow capacity. Therefore, from the perspective of the change of the liquid guide characteristic, the selection range of the back cavity pressure is analyzed, and the relationship between the liquid guides is as follows:

[0227]

[0228]

[0229] In formula one, A is the area of the rodless cavity, a is the area of the rod cavity, p1 and p2 are the pressures of the two cavities, v is the piston speed of the hydraulic cylinder, p s is the pump source pressure, K1 and K2 are the liquid guides of the inlet and outlet valves, and the tank pressure is 0;

[0230] In order to have a real solution of formula two, the change range of the system liquid guide is R 1.5 K e <K1<∞,K e <K2<∞, and the liquid guide equivalent K e is proportional to the piston speed v of the hydraulic cylinder. When the target angle signal is determined, the hydraulic cylinder speed v is obtained, the liquid guide curve is calculated, and the selection range of the back cavity pressure is obtained.

[0231] When starting the volume speed regulation mode, the control signal sent by the controller to the servo motor is obtained as follows:

[0232]

[0233] In formula three, u m1 is the servo motor control signal, g and f are the simplified expressions of the system variable polynomial, d can be regarded as the total disturbance of the system, s m1 is the sliding mode function related to the angle tracking error, and ξ m1is the function characteristic coefficient, -ε m11 s m1 -ε m12 sgn(s m1 ) is the switching law and the exponential approach law; the control signal sent by the controller to the servo proportional valve is obtained as follows:

[0234]

[0235] In formula four, u v is the servo proportional valve control signal, c and h are the simplified expressions of the system variable polynomials, is the derivative of the servo proportional valve spool actual displacement, s2 is the sliding mode function related to the back cavity pressure deviation, -ε3s2-ε4sgn(s2) is the switching law and the exponential approach law, and b is the oil compression coefficient.

[0236] The method for obtaining the signal for judging whether the electromagnetic reversing valve spool switches positions and whether the speed regulation mode is switched is as follows:

[0237] The left wheel rotation is set as positive, the deviation angle between the right steering wheel target rotation angle and the actual rotation angle is taken as the feedback signal, and the positive and negative values of the feedback signal are taken as the control signal of the electromagnetic reversing valve switching, and the control signal function is as follows:

[0238]

[0239] In formula five, u z is the electromagnetic reversing valve control signal, θ q is the rotation error, θ s is the actual rotation angle, k is the set switching signal, and u z =0 when the system stops or the power steering cylinder is locked.

[0240] After the wheel rotation direction changes once, that is, the electromagnetic reversing valve receives the switching voltage signal of the opposite value, the controller switches the volume speed regulation mode to the throttle speed regulation mode, and when the return port pressure drops to the target back cavity pressure, the controller switches the throttle speed regulation mode to the volume speed regulation mode; in the next rotation period, the switching control method is repeated.

[0241] When starting the throttle speed regulation mode, the control signal sent by the controller to the servo motor is obtained as follows:

[0242]

[0243] In formula six, u m2 is the servo motor control signal, H1, F1, and G1 are the simplified expressions of the system variable polynomials, s m2 is the sliding mode function related to the valve port pressure difference, and ξ m2is the function characteristic coefficient, s m21 sgn(s m2 is the function characteristic coefficient, s m22 is the function characteristic coefficient, s m2 is the switching law and the exponential reaching law.

[0244] The control signal sent by the controller to the servo proportional valve is obtained as follows:

[0245]

[0246] In formula seven, u v is the servo proportional valve control signal, G2, F2 are simplified forms of the system variable polynomial, ξ v2 is the function characteristic coefficient, s v2 is the function characteristic coefficient, s v21 is the function characteristic coefficient, s v2 is the function characteristic coefficient, s v22 is the function characteristic coefficient, s v2 is the switching law and the exponential reaching law.

[0247] In step S1, a kind of volume / throttle speed regulation mode switching electro-hydraulic steering system mathematical model is established in the controller, and the method is as follows:

[0248] Step S1-1: the mechanical trapezoidal steering structure modeling of the electro-hydraulic steering rack described in the example is as follows:

[0249] For the electro-hydraulic servo steering system controlled by electromagnetic reversing valve to drive the double steering cylinder to rotate, the relationship between the left and right tire steering angles is represented as:

[0250]

[0251] In formula (1), α and β are the steering angles of the left and right wheels respectively, m is the length of the two sides of the steering arm, γ is the included angle between the steering arm and the axle, L is the distance between the two main pins of the axle, and B is the distance between the two main pins of the axle.

[0252]

[0253] In formula (2), F L is the thrust of the left steering cylinder; F R is the thrust of the right steering cylinder; p1 is the pressure of the left steering cylinder rodless cavity and the right steering cylinder rod cavity; p2 is the pressure of the left steering cylinder rod cavity and the right steering cylinder rodless cavity; A is the piston area of the cylinder rodless cavity; a is the piston area of the cylinder rod cavity.

[0254] Based on the Lagrange equation, the dynamics equation of the steering trapezoidal mechanism is established:

[0255] Based on the Lagrange equation, the dynamics equation of the steering trapezoidal mechanism is established:

[0256] In the formula:

[0257]

[0258]

[0259] In the formula (3-5): T is the generalized kinetic energy of the steering system; D is the generalized dissipation energy of the steering system; J L JLis the equivalent moment of inertia of the left wheel, knuckle and steering trapezoidal arm around the left kingpin; JRis the equivalent moment of inertia of the right wheel, knuckle and steering trapezoidal arm around the left kingpin; C R Cis the equivalent moment of inertia of the right wheel, knuckle and steering trapezoidal arm around the left kingpin; C L , C R are the equivalent damping coefficients of the left and right tires and their related components respectively; Q is the generalized force of the steering system, and

[0260]

[0261]

[0262]

[0263]

[0264] In the formula (6-9): v L , v R are the speeds of the action points of the left and right steering assistance cylinders respectively, n is the distance between the action point of the steering cylinder and the kingpin; s1 is the distance from the assistance cylinder to the bracket hinge on the bracket crossbeam, s2 is the distance from the assistance cylinder to the bracket hinge on the knuckle; θ3', θ3 are the angles between the left and right steering assistance cylinder forces and the action point speeds respectively; T L , T R are the left and right tire resistance moments respectively;

[0265] Through further calculation of each term of the dynamic equation, the following equations are obtained:

[0266]

[0267]

[0268]

[0269]

[0270]

[0271] To simplify the system model, the tank pressure is set to zero, and the distance between the valve and the assistance cylinder is ignored.

[0272] Step S1-2: Modeling the hydraulic system of the electro-hydraulic steering test bench, the method is as follows:

[0273] Step S1-2-1: Establishing the output flow equation of the fixed displacement pump as follows:

[0274] It is defined that the system is in positive motion when steering left, at this time the reversing valve signal u s >0; when steering right, u s <0; and when stopping, u s =0; for the convenience of modeling, ω1 and ω2 are defined to distinguish the left and right steering states of the steering system:

[0275] The servo motor pump used is a fixed displacement- variable speed pump, the output flow is related to the speed and pressure, and the flow equation is as follows:

[0276]

[0277] The servo motor pump used is a fixed displacement- variable speed pump, the output flow is related to the speed and pressure, and the flow equation is as follows:

[0278] Q p =D p K m u m -C i (p s -p T )-C e p s (16)

[0279] In formula (16), Q m is the output flow of the fixed displacement pump; D p is the displacement of the gear pump, and D p =1.6×10 -5 m 3 / r; K m is the voltage- speed conversion coefficient, and K m =200 / 60, that is, 1V corresponds to 200r / min; u m is the input voltage of the servo motor; C i and C e are the internal and external leakage coefficients of the servo motor pump; the tank pressure p T =0, C p =C i +C e , then:

[0280] Q p =D p K m u m -C p (ω1p1+ω2p2) (17)

[0281] In formula (17), p1 and p2 are the working pressures on the left and right sides of the port of the reversing valve, respectively;

[0282] Step S1-2-2: Establish the flow equation of the hydraulic power cylinder:

[0283] According to the flow continuity equation, the inlet and outlet flow rates of the double-reversing power cylinder are:

[0284]

[0285]

[0286] In formula (18-19), β e represents the effective bulk modulus; A is the rodless cavity area of the power steering cylinder, and a is the rod cavity area of the power steering cylinder; and represent the velocities of the left and right power cylinders; C ip and C ep represent the internal and external leakage coefficients of the power cylinder; V t is the total volume of the power cylinder;

[0287] Step S1-2-3: Flow equation of the electromagnetic reversing valve and the servo proportional valve:

[0288] Suppose that the throttling loss of the electromagnetic reversing valve is small, and the back cavity pressure chamber of the hydraulic cylinder is directly connected to the servo proportional valve to reduce the complexity of the model; the flow equation of the servo proportional valve is:

[0289]

[0290] In formula (20), C d is the flow coefficient of the servo proportional valve; w is the area gradient of the servo proportional valve; ρ is the oil density; Δp is the pressure difference before and after the valve port of the servo proportional valve; and therefore the inlet and outlet flow rates of the power cylinder are:

[0291]

[0292] In formula (21), q1 is the flow rate of port A of the first reversing valve; q2 is the flow rate of port B of the first reversing valve; x v is the displacement of the servo proportional valve spool;

[0293] Step S1-2-4: The working pressure dynamic equation of the two cavities of the power cylinder can be obtained by simultaneous equations:

[0294]

[0295] The target back cavity pressure of the volume speed regulation mode of the input controller is:

[0296] The hydraulic control system is provided, and the flow guide can be used as a measure of the valve port opening degree under certain flow coefficient and oil density. The greater the valve port flow guide is, the greater the valve port opening degree and the stronger the flow capacity are. Therefore, the selection range of the back cavity pressure is analyzed from the perspective of the change of the flow guide characteristics, and the relationship between the flow guides is as follows:

[0297]

[0298]

[0299] In formula (23), A is the rodless cavity action area, a is the rod cavity action area, p1 and p2 are the pressures of the two cavities respectively, v is the hydraulic cylinder piston movement speed, p s is the pump source pressure, K1 and K2 are the inlet and outlet oil valve port flow guides, and the tank pressure is 0;

[0300] In order to have a real solution of formula (24), the change range of the system flow guide is R 1.5 K e <K1<∞,K e <K2<∞, and the flow guide equivalent K e is proportional to the hydraulic cylinder piston speed v. When the target rotation angle signal is determined, the hydraulic cylinder speed v can be obtained, the flow guide curve can be calculated, and the selection range of the back cavity pressure can be obtained.

[0301] In the embodiment, the target back cavity pressure in the volume speed regulation mode is 0.5 MPa.

[0302] Figure 3 The control method flow chart of the volume speed regulation mode of the system involved in the application.

[0303] In steps S2-S3, the volume speed regulation mode is started, and the controller sends control signals to the servo motor and the servo proportional valve respectively. The specific method is as follows:

[0304] Step S2 / 3-1: The mathematical model of the electro-hydraulic servo steering system is converted into a state space model, and the corresponding system state space equation is as follows:

[0305] The system state variables are defined as Each state is: the right wheel rotation angle β, the right wheel rotation angle speed The back cavity pressure p t , the system input is the servo motor voltage u m1 , and the servo proportional valve voltage output system is the right wheel rotation angle back cavity pressure system state space equation as follows:

[0306]

[0307] The input variable y1=x1 is defined, and the third order derivative is obtained:

[0308]

[0309] In formula (26):

[0310]

[0311]

[0312]

[0313] Define input variable y2=x3, and the second derivative is:

[0314]

[0315] In formula (30):

[0316]

[0317]

[0318]

[0319] Step S2 / 3-2: define the rotation angle control error e m1 is:

[0320] e m1 =y1-y 1d (34)

[0321] Construct a sliding mode function s related to the rotation angle tracking error m1 , s m1 =0 is a sliding surface, and by designing conditions to make the state converge to the sliding surface s m1 =0, then the rotation angle tracking error e1 can also converge to 0, and the sliding surface is as follows:

[0322]

[0323] In formula (35), n is the order of the system, and according to the rotation angle control input-output linearization expression, n=3, so:

[0324]

[0325] In formula (36), λ>0, and the sliding surface dynamics is obtained by deriving the sliding surface:

[0326]

[0327] Step S3 / 4-3: according to the motor voltage-rotation angle input-output relationship, establish the sliding surface s m1 and the controller input u m1The relationship is:

[0328]

[0329] In formula (38),

[0330] In order to make the control continuous The control law is obtained by formula:

[0331]

[0332] In order to make the state converge to the sliding mode surface s m1 =0 The switching law is added to the sliding mode dynamics, and the control law of the servo motor is obtained:

[0333]

[0334] In the control process, in order to ensure the approximation speed and suppress the sliding mode chattering phenomenon, a larger ε m12 and a smaller ε m11 should be taken, and the improved servo motor control law is obtained:

[0335]

[0336] In formula (41), u m1 is the servo motor control signal, g and f are the simplified expressions of the polynomial of system variables, d can be regarded as the total disturbance of the system, s m1 is the sliding mode function related to the tracking error of the rotation angle, ξ m1 is the characteristic coefficient of the function, -ε m11 s m1 -ε m12 sgn(s m1 ) is the switching law and exponential reaching law.

[0337] Step S2 / 3-4: Design the back cavity pressure sliding mode control law, define the back cavity pressure control error e v1 as:

[0338] e v1 =y2-y 2d (42)

[0339] Define the sliding mode function s v1 , according to the input-output linearization expression of the back cavity pressure control, n=1, so s v1 =e v1 ;

[0340] The derivative of the sliding mode function is:

[0341]

[0342] The switching law and the exponential approach law are also selected:

[0343]

[0344] The servo proportional valve control law is obtained as:

[0345]

[0346] In formula (45), u v is a servo proportional valve control signal, c and h are simplified expressions of system variable polynomials, is a derivative of the actual displacement of the servo proportional valve spool, s2 is a sliding mode function related to the back cavity pressure deviation, -ε3s2-ε4sgn(s2) is a switching law and an exponential approach law, and b is an oil compression coefficient;

[0347] In step S4, whether the electromagnetic reversing valve spool switches position and whether the speed regulation mode switches are judged by a signal:

[0348] The left wheel turning is set as positive, the deviation angle between the right turning wheel target turning angle and the actual turning angle is taken as a feedback signal, and the positive and negative values of the feedback signal are taken as the control signal of the electromagnetic reversing valve switching, and the control signal function is as follows:

[0349]

[0350] In formula (46), u z is an electromagnetic reversing valve control signal, θ q is a turning angle error, θ s is an actual turning angle, k is a set switching signal, and u z =0 when the system stops or the power steering cylinder is locked.

[0351] After the wheel turning direction changes once, that is, the electromagnetic reversing valve receives the switching voltage signal of the opposite value, the controller switches the volume speed regulation mode to the throttle speed regulation mode, and when the return port pressure drops to the target back cavity pressure, the controller switches the throttle speed regulation mode to the volume speed regulation mode; in the next turning period, the switching control method is repeated.

[0352] Figure 4 The flow chart of the control method of the throttle speed regulation mode of the system involved in the present application is shown.

[0353] In steps S5-S6, the throttle speed regulation mode is started, and the control signals sent by the controller to the servo motor and the servo proportional valve are:

[0354] Step S5 / 6-1: the mathematical model of the electro-hydraulic servo steering system is converted into a state space model, and the corresponding system state space equation is as follows:

[0355] With X = [x1, x2, x3, x4, x5] T Define the state variables of the system, which include: right wheel steering angle x1 = β, right wheel steering speed Servo proportional valve inlet valve port pressure x3 = p1, servo proportional valve outlet valve port pressure x4 = p2, pump source pressure x5 = p s , the following state space equation is obtained:

[0356]

[0357] Define the output variables of the system y1 = x1, y2 = Δp, and multiple derivatives of both can obtain the input u m2 , u v2 and the relationship between the output y1, y2:

[0358]

[0359]

[0360]

[0361]

[0362] The relationship between the input and output of the steering angle control system is obtained by the above equations (48-51):

[0363]

[0364] For y2 = Δp, the inlet valve port pressure difference Δp has two states of left turn and right turn:

[0365]

[0366] The equation can be expressed as:

[0367]

[0368] Under the conditions of left turn and right turn, the derivative of y2 is obtained as follows:

[0369] Left turn condition:

[0370]

[0371] Right turn condition:

[0372]

[0373] F2, G2, H2 are simplified expressions of polynomials containing system variables, and the relationship between the input and output of the valve pressure difference control system is obtained through the above process:

[0374]

[0375] Step S5 / 6-2: The sliding surface of the steering angle control system is designed by the control guide model:

[0376]

[0377] In formula (58), λ1>0, e v2 =y1-y 1d is the steering angle tracking error;

[0378] Taking the derivative of the above formula, the sliding surface dynamic is:

[0379]

[0380] In formula (59),

[0381] In order to meet the condition of reaching the sliding surface s v2 =0, that is, And in order to further improve the dynamic performance of the system in the reaching motion stage, a switching law and an exponential reaching law are added to the sliding surface dynamic, respectively:

[0382]

[0383] By solving (59-60) simultaneously, F2+G2u v2 +ξ v2 =-ε v21 sgn(s v2 )-ε v22 s v2 , the sliding mode control law of the steering angle tracking system is finally obtained:

[0384]

[0385] Step S5 / 6-3: The sliding surface of the valve pressure difference control system is:

[0386] s m2 =e m2 (62)

[0387] In the formula, e m2 =y2-y 2d is the difference between the actual value and the expected value of the valve pressure difference;

[0388] Taking the derivative of the above formula, the sliding surface dynamic is:

[0389]

[0390] In formula (63),

[0391] Similarly, the switching law and the exponential approach law are added in the sliding mode surface dynamics:

[0392]

[0393] Through the simultaneous equations (63-64), F1+G1u can be obtained v2 +H1u m2 +ξ m2 =-ε m21 s m2 gn(s m22 s m2 , and finally the sliding mode control law of the valve pressure difference control system is:

[0394]

[0395] In this example, Figure 2 is the flow chart of the steering control method of the system involved in the present application.

[0396] In this embodiment, a heavy vehicle electro-hydraulic steering bench is taken as an example, and the related parameters are as follows:

[0397]

[0398] Preferably, in this embodiment, the right wheel target angle curve y d = 20sin(0.2πt) is set for the cyclic steering working condition, and the steering angle amplitude is ±20°. The above only describes the preferred embodiments of the present application, and is not intended to limit the other forms of the present application. Any skilled person in the art can use the disclosed technical content to make changes or modifications to equivalent embodiments. However, any simple modification, equivalent change and modification made according to the technical essence of the present application without departing from the technical solution content of the present application still belongs to the protection scope of the present application.

Claims

1. An electro-hydraulic steering system with volumetric / throttle speed regulation mode switching, used for vehicle steering, characterized in that: The system includes an open-type variable speed pump control system, a return port valve control system, a signal acquisition system, and an electronic control system. The open-type variable speed pump control system controls the power steering cylinders of the left and right wheels through a servo motor pump unit and an electromagnetic reversing valve. The return port valve control system is used to adjust the pressure in the return port back chamber and control the flow rate of the system. The signal acquisition system is used to acquire the steering angle of the steering wheels and the pressure at each designated loop position. The electronic control system outputs control commands to control the vehicle's steering system in real time based on its input target steering angle signal, wheel steering angle sensor signal, and pressure sensor signal at each designated loop position. The open variable speed pump control system includes a servo motor (2), a one-way quantitative pump (20), an electromagnetic reversing valve (5), a first hydraulic control check valve (6), a second hydraulic control check valve (14), a first overflow valve (3), a second overflow valve (7), a third overflow valve (13), a fourth overflow valve (17), a first electromagnetic reversing ball valve (15), and an oil tank (1). The return port valve control system includes a servo proportional valve, which adjusts the pressure in the return port back chamber by controlling the displacement of the servo proportional valve core. The signal acquisition system includes a first angle sensor, a first pressure sensor, a second pressure sensor, and a third pressure sensor; The electronic control system includes a controller and a servo motor driver. It calculates and outputs a corresponding control voltage signal based on the target turning angle required by the wheel, the actual turning angle of the steering wheel, and the pressure of the pressure sensors at each designated loop position. It also switches between two modes, namely pump-controlled volumetric speed regulation mode and pump-valve joint control throttling speed regulation mode, based on the actual system loop pressure, thereby controlling the electro-hydraulic steering system. The servo motor and the unidirectional metering pump are connected by a key to form a servo motor pump. The servo motor is electrically connected to the servo driver (18). The motor speed is changed by the voltage control signal, thereby adjusting the output flow and pressure of the motor pump. The rodless chamber of the left power steering cylinder and the rod chamber of the right power steering cylinder of the vehicle steering system are connected to port B of the first hydraulic check valve. Port A of the first hydraulic check valve is connected to port A of the first solenoid directional ball valve to form the first working circuit. The rod chamber of the left steering assist cylinder and the rodless chamber of the right steering assist cylinder are connected to port B of the second hydraulic check valve. Port A of the second hydraulic check valve is connected to port B of the first solenoid directional ball valve to form a second working circuit. The first hydraulic control check valve B port is connected to the second relief valve A port, and the second relief valve T port is connected to the oil tank; the second hydraulic control check valve B port is connected to the third relief valve A port, and the third relief valve T port is connected to the oil tank; the first electromagnetic reversing ball valve P port is connected to the outlet of the metering pump and is connected in parallel to the first relief valve A port, and the first relief valve T port is connected to the oil tank. The T port of the first electromagnetic reversing ball valve is connected to the P port of the servo proportional valve (19), the A port of the servo proportional valve is connected to the oil tank, and the T port and B port of the servo proportional valve are both sealed with plugs. The K port of the first hydraulic control check valve and the K port of the second hydraulic control check valve are both connected to the A port of the first solenoid directional ball valve. The P port of the first solenoid directional ball valve is connected to the P port of the first solenoid directional ball valve. The T port of the first solenoid directional ball valve is connected to the oil tank.

2. The electro-hydraulic steering system with volumetric / throttling speed regulation mode switching according to claim 1, characterized in that: In the hydraulic system of the open variable speed pump control system, the first hydraulically controlled check valve, the second hydraulically controlled check valve, and the first solenoid directional ball valve form a hydraulic lock-up circuit. By setting the first relief valve, the second relief valve, the third relief valve, and the fourth relief valve to protect the circuit, excessive circuit pressure is prevented from causing danger to the system operation. When the pressure at each node of the system exceeds the protection pressure set by each relief valve, the oil corresponding to the excess pressure will flow back to the oil tank through the relief valve.

3. A control method for an electro-hydraulic steering system with volumetric / throttle speed regulation mode switching, employing the electro-hydraulic steering system with volumetric / throttle speed regulation mode switching as described in claim 1, characterized in that: The control method employs a composite control strategy that switches between volumetric speed regulation and throttling speed regulation modes. In volumetric speed regulation mode, the servo motor is controlled based on the angular deviation signal of the wheel steering angle, thereby controlling the system pressure. The servo proportional valve is then controlled based on the return oil port pressure to achieve the set back cavity pressure. In throttling speed regulation mode, the servo motor is controlled based on the target angular velocity feedforward and valve port pressure difference feedback. The valve opening of the servo proportional valve is adjusted based on the angular deviation signal, thereby controlling the system flow rate. This meets the high-precision dynamic steering requirements of the electro-hydraulic servo steering system and achieves on-demand power supply, energy saving and consumption reduction, avoids negative pressure in the system cavity, and improves the overall rigidity and stability of the system.

4. The control method for an electro-hydraulic steering system with volumetric / throttling speed regulation mode switching according to claim 3, characterized in that: The control method includes the following steps; Step S1: The system starts up. A mathematical model and control guidance model of an electro-hydraulic steering system with volumetric / throttle speed regulation mode switching are established in the controller. The target wheel turning angle signal and the target back cavity pressure signal under the volumetric speed regulation mode are input to the controller. The analog signals collected by the right wheel angle sensor and pressure sensor are input to the controller. The pressure difference between the inlet and outlet of the servo proportional valve and the deviation between the actual turning angle of the right wheel and the target turning angle are calculated. Step S2: Start the volumetric speed regulation mode, with the servo motor pump as the main control object. Based on the deviation signal between the target turning angle and the actual turning angle of the right wheel, the controller sends a signal to the servo motor and the solenoid reversing valve to control their operation, thereby controlling the output pressure of the variable speed pump to form hydraulic driving force, so that the left and right steering assist cylinders extend and retract and allow the wheel to reach the target turning angle. Step S3: Based on the deviation signal between the system return port back cavity pressure and the desired back cavity pressure, the controller outputs a suitable voltage signal to drive the servo proportional valve to work, dynamically adjusting the pressure at the return port to achieve the set back cavity pressure. Step S4: Determine whether the valve core of the solenoid directional valve has switched positions. If yes, proceed to step S5; otherwise, continue with steps S2-S3. Step S5: Start the throttling speed regulation mode, take the return oil port servo proportional valve as the main control object, calculate the angular velocity according to the target rotation angle signal of the right wheel, and control the servo motor pump through angular velocity feedforward and servo proportional valve port pressure difference feedback, so that the motor pump outputs the desired flow rate. Step S6: Calculate a suitable voltage value in the controller based on the deviation signal between the target turning angle and the actual turning angle of the right wheel, drive the servo proportional valve core to produce a corresponding displacement, adjust the valve opening of the servo proportional valve, and thus control the system flow rate, so that the left and right steering assist cylinders extend and retract and the wheel reaches the target turning angle. Step S7: Determine whether the system return port pressure has reached the set back cavity pressure required by the system at this time. If yes, proceed to step S2; otherwise, continue to execute steps S5-S6. Step S8: When the controller receives the stop command, both the solenoid directional valve and the servo proportional valve are in the neutral position, and the servo motor stops.

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