An electrically controlled full hydraulic steering system and control method considering steering clearance and hydraulic nonlinearity

Abstract

The application discloses an electrically-controlled full-hydraulic steering system considering steering clearance and hydraulic nonlinearity, which comprises a full-hydraulic steering module, the full-hydraulic steering module comprises a steering oil cylinder, a full-hydraulic steering valve, a steering oil pump, a direct-current motor, an oil tank, an intelligent control system and a rotor-stator pair, the intelligent control system comprises an MPC trajectory tracking module considering the steering clearance and a full-hydraulic steering sliding mode control module considering the hydraulic nonlinearity. The application obtains the target front wheel steering angle considering the steering clearance through prediction, optimization and quadratic programming solution, then builds a full-hydraulic steering model based on the hydraulic nonlinearity and deduces a piston displacement tracking state space equation, then solves the direct-current motor voltage by using a sliding mode controller according to the piston displacement tracking state space equation, and finally controls the motor output shaft rotating speed by controlling the direct-current motor voltage to realize vehicle steering, so that the precision of vehicle tracking a reference trajectory can be effectively improved.

Description

TECHNICAL FIELD

[0001] The present application belongs to the field of vehicle steering systems, and particularly relates to an electrically-controlled full-hydraulic steering system and control method considering steering clearance and hydraulic nonlinearity. BACKGROUND

[0002] With the rapid development of intelligent technology, the intelligentization of agricultural machinery and engineering machinery and other off-highway vehicles improves work efficiency, but also needs to meet the requirements of control accuracy, so it is imminent to combine intelligent control systems with traditional hydraulic systems to realize the "integration of liquid-electric intelligence and machine" of off-highway vehicles. Due to the existence of steering clearance and the nonlinear characteristics of the hydraulic system in the full-hydraulic steering system, steering deviation will occur during steering, resulting in that the target front wheel steering angle and the actual front wheel steering angle cannot correspond one-to-one, so that the intelligent control system cannot guarantee the vehicle to travel according to the established target trajectory, thereby causing great inconvenience and safety hazards. How to solve the problem of low tracking accuracy of full-hydraulic steering intelligent vehicles has become the focus and difficulty of research in the academic and industrial circles. SUMMARY

[0003] In order to solve the problems in the prior art, the present application proposes an electrically-controlled full-hydraulic steering system and control method considering steering clearance and hydraulic nonlinearity. The intelligent control system considers the influence of the hydraulic nonlinearity and steering clearance in the full-hydraulic steering system on the tracking trajectory accuracy of intelligent vehicles, realizes the close combination of the intelligent control system and the full-hydraulic steering system, and especially solves the problem of tracking trajectory accuracy of intelligent commercial vehicles.

[0004] An electrically-controlled full-hydraulic steering system considering steering clearance and hydraulic nonlinearity, characterized in that the electrically-controlled full-hydraulic steering system comprises a full-hydraulic steering module, the full-hydraulic steering module comprises a steering oil cylinder, a full-hydraulic steering valve, a steering oil pump, a direct-current motor, an oil tank, an intelligent control system and a stator-rotor pair,

[0005] wherein,

[0006] The steering oil cylinder is provided with a first oil port and a second oil port;

[0007] The full-hydraulic steering valve is provided with an oil inlet P1, an oil outlet T1, an oil port A, an oil port B, an oil port C and an oil port D;

[0008] The stator-rotor pair is connected with the oil port C and the oil port D of the full-hydraulic steering valve;

[0009] The oil port B of the full hydraulic steering valve is connected with the second oil port of the steering cylinder through a hydraulic pipeline, the first oil port of the steering cylinder is connected with the oil port A of the full hydraulic steering valve through a hydraulic pipeline, when the full hydraulic steering valve moves right, the oil inlet P1 is connected with the oil port C, the oil liquid passes through the fixed rotor pair to reach the oil port D from the oil port C, the oil port D is connected with the oil outlet T1, when the full hydraulic steering valve moves left, the oil inlet P1 is connected with the oil port D, the oil liquid passes through the fixed rotor pair to reach the oil port C from the oil port D, the oil port C is connected with the oil outlet T1;

[0010] The steering oil pump is connected with the oil tank, and the steering oil pump supplies oil to the full hydraulic steering valve;

[0011] The rotating shaft of the direct current motor is connected with the input shaft of the fixed rotor pair;

[0012] The fixed rotor pair is connected with the full hydraulic steering valve;

[0013] The signal input end of the direct current motor is connected with the signal output end of the intelligent control system;

[0014] The intelligent control system comprises an MPC trajectory tracking module considering the steering gap and a full hydraulic steering sliding mode control module considering the hydraulic nonlinear;

[0015] The MPC trajectory tracking module considering the steering gap obtains the target front wheel steering angle according to the planned path, the vehicle state and the steering gap, and adopts an MPC-based trajectory tracking controller;

[0016] The full hydraulic steering sliding mode control module considering the hydraulic nonlinear obtains the state equation of the tracked piston displacement according to the full hydraulic dynamic model considering the hydraulic nonlinear, and adopts a sliding mode controller to obtain the direct current motor voltage.

[0017] As a preferred embodiment, step S1, the target front wheel steering angle is obtained according to the reference trajectory, the vehicle state and the steering gap, and is converted into the target piston displacement by adopting the MPC trajectory tracking controller;

[0018] Step S2, the piston displacement tracking state space equation is obtained according to the full hydraulic steering dynamic model, wherein the full hydraulic steering dynamic model considers the hydraulic nonlinear and other external factors;

[0019] Step S3, the relationship between the direct current motor voltage and the output shaft speed is obtained according to the direct current motor model;

[0020] Step S4, the direct current motor voltage is obtained by adopting the sliding mode controller according to the piston displacement tracking state space equation in step S2 and the difference between the target piston displacement and the actual piston displacement.

[0021] As a preferred embodiment, in step S1, the target front wheel steering angle calculation method calculated by the MPC trajectory tracking controller considering the steering gap is:

[0022] S1.1, the discrete state equation of vehicle dynamics is established as follows:

[0023]

[0024] Wherein: x A (k) = [v y (k) ω r (k) ψ(k) y(k)] T ,

[0025]

[0026] In the formula, ω r is the vehicle yaw rate; m is the mass of the vehicle; C f is the tire cornering stiffness of the front wheel; C r is the tire cornering stiffness of the rear wheel; a is the distance from the mass center to the front axle; b is the distance from the mass center to the rear axle; I z is the moment of inertia of the vehicle; δ f is the front wheel steering angle; v y is the lateral velocity; ψ is the heading angle; y is the lateral displacement; k is the k time; v x is the longitudinal velocity; x A is the state quantity; δ c is the front wheel steering angle caused by the steering gap; u A is the control quantity;

[0027] S1.2, the objective function and constraint condition of MPC are constructed,

[0028] The objective function is as follows:

[0029]

[0030] In the formula, r A is the expected value of the control target of the MPC controller; Q A is the MPC weight matrix; N p is the prediction step; N c is the control step;

[0031] The constraint condition is as follows:

[0032] u min (k+j)≤u A (k+j)≤u max (k+j)j = 0, 1, …, N c -1

[0033] Δu min (k+j)≤Δu A (k+j)≤Δu max (k+j)j=0,1,…,N c -1

[0034] y min (k+j)≤y A (k+j)≤y max (k+j)j=0,1,…,N p

[0035] In the formula, u min u max Δu represents the upper and lower limits of the control quantity. min , Δu max To control the upper and lower limits of the increment; y min y max These are the upper and lower limits of the output quantity;

[0036] S1.3. The objective function is solved using a quadratic programming algorithm to obtain the target front wheel steering angle. The process is as follows:

[0037]

[0038] δ f =[10...0]U A

[0039] y d =i q δ f

[0040] In the formula, U A For the target front wheel steering angle control sequence; i q y is the transmission ratio between the front wheel and the piston rod; d For the desired piston displacement;

[0041] Where H = Θ T Q A Θ+P A g = EQ A Θ; E = R A -Ψx(k)-ΘU c

[0042]

[0043] In a preferred embodiment, in step S2, the piston displacement tracking state equation of the fully hydraulic steering system based on hydraulic nonlinearity is:

[0044] S2.1 The dynamic equations for fully hydraulic steering are as follows:

[0045] my" = (p3 - p4)A - By' + f c +f(t,x1,x2)

[0046] where m is the mass of the tire load; y is the second derivative of the piston displacement; p3, p4 are the pressures in the two chambers of the steering cylinder; A is the area of the cylinder piston; B is the viscous friction coefficient; y' is the first derivative of the piston displacement; f c is the constant disturbance; f(t,x1,x2) is the concentrated disturbance caused by other external disturbances and hydraulic nonlinearities;

[0047] S2.2, the pressure dynamics equation in the double rod hydraulic cylinder is as follows:

[0048] p L = p3 - p4

[0049]

[0050] where V t is the total control volume of the steering cylinder; β e is the effective oil volume modulus; C t is the leakage coefficient of the steering cylinder; Q L is the average flow of the stator-rotor pair; Q c is the constant error; Q(t,x1,x2) is the time-varying error;

[0051] S2.3, the average flow calculation formula through the stator-rotor pair is as follows:

[0052]

[0053]

[0054] The balance equation of the output torque of the stator-rotor pair and the load torque is as follows:

[0055] T L = k m q(p1 - p2) = J m θ" m + B m θ' m + Kθ m + T a

[0056] where q is the theoretical displacement; ω m is the angular speed of the motor output shaft; k m is the transmission ratio of the reduction mechanism; Q m1 , Q m2 are the inflow and outflow of the stator-rotor pair, respectively; C im , C cmThese are the internal and external leakage coefficients of the stator and rotor pair, respectively; p1 and p2 are the inlet and outlet pressures of the stator and rotor pair, respectively; C lm T is the total leakage coefficient of the stator and rotor pair; L J is the load torque of the DC motor. m B is the equivalent rotational inertia on the stator and rotor secondary shafts; m T is the viscous damping coefficient; K is the load spring stiffness; T a θ represents the load torque of the stator and rotor pair; m The rotor-stator pair rotation angle;

[0057] S2.4 The state-space equation for piston displacement tracking in a fully hydraulic steering system based on hydraulic nonlinearity is described as follows:

[0058]

[0059] In the formula, the state variable is defined as follows:

[0060] In a preferred embodiment, the relationship between the voltage of the DC motor and the output shaft speed in step S3 is as follows:

[0061] The DC motor model is shown below:

[0062]

[0063] We can obtain:

[0064]

[0065] In the formula, u v For input voltage; C e C is the electromotive force coefficient of the DC motor; m T is the torque coefficient of the DC motor; m R is the motor time constant of the motor drive system; L is the total armature circuit resistance; L is the total armature circuit inductance; I' d a is the rate of change of armature current; m T1 is the angular acceleration of the motor output shaft; T2 is the electromagnetic time constant of the armature circuit.

[0066] In a preferred embodiment, the DC motor voltage sliding mode control calculation method in step S4 is as follows:

[0067] S4.1. Based on the difference between the target piston displacement and the actual piston displacement, a sliding mode controller is used to obtain the DC motor voltage. The construction process of the sliding mode controller is as follows:

[0068] 1) Construct the piston displacement tracking state space equation of the fully hydraulic steering system based on hydraulic nonlinearity as shown in S2.4.

[0069] 2), the sliding surface and the exponential approach law are as follows:

[0070] s=c1e1+c2e2+e3=c1(x 1d -x1)+c2(x 2d -x2)+(x 3d -x3)=0c>0

[0071] s'=-εsgn(s)-psε>0,p>0

[0072] In the formula, c1, c2, c3>0, satisfy Hurwitz condition, c1, c2, c3 are sliding mode control parameters;ε, p are exponential approach law control parameters.

[0073] 3), the control law is as follows:

[0074]

[0075] In the formula,

[0076] S4.2, the motor output shaft speed calculation method is as follows:

[0077]

[0078] Compared with the prior art, the above technical scheme has the following technical effects:

[0079] The application proposes an electrically controlled full hydraulic steering system and a control method considering steering clearance and hydraulic nonlinearity, the MPC trajectory tracker in the application considers the characteristics of the steering clearance of the steering system, and obtains the target front wheel steering angle considering the steering clearance through prediction, optimization and quadratic programming solution;The sliding mode controller obtains the piston displacement tracking state space equation according to the full hydraulic steering system model considering hydraulic nonlinearity, and then solves the direct current motor voltage. The direct current motor voltage is controlled to control the motor output shaft speed, so as to realize accurate control of vehicle steering, and solve the problem of low tracking accuracy of intelligent commercial vehicles based on the electrically controlled full hydraulic steering system. BRIEF DESCRIPTION OF DRAWINGS

[0080] Figure 1 is the structural principle diagram of the electrically controlled full hydraulic steering system;

[0081] Figure 2 is the structural schematic diagram of the intelligent control system;

[0082] Figure 3 is the control flow chart of the electrically controlled full hydraulic steering system;

[0083] 1: steering oil pump; 2: full hydraulic steering valve; 3: steering cylinder; 4: intelligent control system; 5: oil tank; 6: DC motor; 7: stator-rotor pair. DETAILED DESCRIPTION

[0084] In order to make the purpose, technical scheme and advantages of the present application more clear, the present application is further described in detail below in combination with the drawings and examples. It should be understood that the specific examples described herein are only used to explain the present application and do not limit the present application.

[0085] The electrically controlled full hydraulic steering system considering the steering gap and the hydraulic nonlinearity,

[0086] As shown in Figure 1 and Figure 2 , the electrically controlled full hydraulic steering system comprises a full hydraulic steering module, and the full hydraulic steering module comprises a steering cylinder 3, a full hydraulic steering valve 2, a steering oil pump 1, a DC motor 6, an oil tank 5, an intelligent control system 4 and a stator-rotor pair 7.

[0087] The steering cylinder 3 is provided with a first oil port 3-1 and a second oil port 3-2.

[0088] The full hydraulic steering valve 2 is provided with an oil inlet port P1, an oil outlet port T1, an oil port A, an oil port B, an oil port C and an oil port D.

[0089] The stator-rotor pair 7 is connected with the oil port C and the oil port D of the full hydraulic steering valve 2.

[0090] The oil port B of the full hydraulic steering valve 2 is connected with the second oil port 3-2 of the steering cylinder 3 through a hydraulic pipeline, the first oil port 3-1 of the steering cylinder 3 is connected with the oil port A of the full hydraulic steering valve 2 through a hydraulic pipeline, when the full hydraulic steering valve 2 moves to the right, the oil inlet port P1 is connected with the oil port C, the oil liquid reaches the oil port D from the oil port C through the stator-rotor pair 7, and the oil port D is connected with the oil outlet port T1, when the full hydraulic steering valve 2 moves to the left, the oil inlet port P1 is connected with the oil port D, the oil liquid reaches the oil port C from the oil port D through the stator-rotor pair 7, and the oil port C is connected with the oil outlet port T1.

[0091] The steering oil pump 1 is connected with the oil tank 5, and the steering oil pump 1 supplies oil to the full hydraulic steering valve 2.

[0092] The rotating shaft of the DC motor 6 is connected with the input shaft of the stator-rotor pair 7.

[0093] The stator-rotor pair 7 is connected with the full hydraulic steering valve 2.

[0094] The signal input end of the DC motor 6 is connected with the signal output end of the intelligent control system 4.

[0095] The intelligent control system 4 includes an MPC trajectory tracking module 41 considering the steering gap and a full-hydraulic steering sliding mode control module 42 considering the hydraulic nonlinearity.

[0096] The MPC trajectory tracking module 41 considering the steering gap solves the target front wheel steering angle by using an MPC-based trajectory tracking controller according to the planned path, the vehicle state and the steering gap.

[0097] The full-hydraulic steering sliding mode control module 42 considering the hydraulic nonlinearity derives a state equation for tracking the piston displacement according to a full-hydraulic dynamics model considering the hydraulic nonlinearity, and solves the DC motor voltage by using a sliding mode controller.

[0098] Based on the above-mentioned electric control full-hydraulic steering system considering the steering gap and the hydraulic nonlinearity, the application proposes a control method for the electric control full-hydraulic steering system considering the steering gap and the hydraulic nonlinearity, as shown in the figure, and the specific steps are as follows: Figure 3

[0099] Step S1, according to the reference trajectory, the vehicle state and the steering gap, the target front wheel steering angle is derived by using an MPC trajectory tracking controller, and is converted into a target piston displacement,

[0100] Specifically:

[0101] S1.1, the discrete state equation of vehicle dynamics is established, as shown below:

[0102]

[0103] Wherein: x A (k)=[v y (k)ω r (k)ψ(k)y(k)] T ,

[0104]

[0105] In the formula, ω r is the vehicle yaw rate; m is the vehicle mass; C f is the tire cornering stiffness of the front wheel; C r is the tire cornering stiffness of the rear wheel; a is the distance from the mass center to the front axle; b is the distance from the mass center to the rear axle; I z is the vehicle moment of inertia; δ f is the front wheel steering angle; v y is the lateral velocity; ψ is the heading angle; y is the lateral displacement; k is the k-th time; v x is the longitudinal velocity; x A is the state quantity; δ c is the front wheel steering angle caused by the steering gap; u A ​For control quantity.

[0106] S1.2, construct the objective function and constraint conditions of MPC,

[0107] The objective function is as follows:

[0108]

[0109] In the formula, r A is the expected value of the control target of the MPC controller; Q A is the MPC weight matrix; N p is the prediction step; N c is the control step.

[0110] The constraint condition is as follows:

[0111] u min (k+j)≤u A (k+j)≤u max (k+j)j=0,1,…,N c -1

[0112] Δu min (k+j)≤Δu A (k+j)≤Δu max (k+j)j=0,1,…,N c -1

[0113] y min (k+j)≤y A (k+j)≤y max (k+j)j=0,1,…,N p

[0114] In the formula, u min , u max are the upper and lower limits of the control quantity; Δu min , Δu max are the upper and lower limits of the control increment; y min , y max are the upper and lower limits of the output quantity.

[0115] S1.3, the objective function is solved by using a quadratic programming algorithm to obtain the target front wheel steering angle, the process is as follows:

[0116]

[0117] δ f =[10…0]U A

[0118] y d =i q δ f

[0119] where U A is the target front wheel angle control sequence; i q is the transmission ratio between the front wheel and the piston rod; y d is the desired piston displacement;

[0120] where H = Θ T Q A Θ + P A ; g = EQ A Θ; E = R A - Ψx(k) - ΘU c

[0121]

[0122] Step S2, deriving the piston displacement tracking state space equation according to the full hydraulic steering dynamics model, wherein the full hydraulic steering dynamics model considers hydraulic nonlinearities and other external factors. Specifically, it is:

[0123] S2.1, the dynamics equation of the full hydraulic steering is as follows:

[0124] my” = (p3-p4)A - By’ + f c +f(t,x1,x2)

[0125] where m is the mass of the tire load; y is the second derivative of the piston displacement; p3, p4 are the pressures in the two chambers of the steering cylinder; A is the area of the cylinder piston; B is the viscous friction coefficient; y’ is the first derivative of the piston displacement; f c is a constant disturbance; f(t,x1,x2) is a concentrated disturbance caused by other external disturbances and hydraulic nonlinearities.

[0126] S2.2, the pressure dynamics equation in the double-rod hydraulic cylinder is as follows:

[0127] p L = p3-p4

[0128]

[0129] where V t is the total control volume of the steering cylinder; β e is the effective oil volume modulus; C t is the leakage coefficient of the steering cylinder; Q L is the average flow of the fixed rotor pair; Q c is a constant error; Q(t,x1,x2) is a time-varying error.

[0130] S2.3, the average flow calculation formula through the fixed rotor pair is as follows:

[0131]

[0132] The balance equation of the stator-rotor pair output torque and load torque is as follows:

[0133] T L = k m q(p1-p2)=J m θ" m +B m θ' m +Kθ m +T a

[0134] In the formula, q is the theoretical displacement; ω m is the angular speed of the motor output shaft; k m is the transmission ratio of the reduction mechanism; Q m1 and Q m2 are the inflow and outflow of the stator-rotor pair respectively; C im and C cm are the internal and external leakage coefficients of the stator-rotor pair respectively; p1 and p2 are the inlet and outlet pressures of the stator-rotor pair respectively; C lm is the total leakage coefficient of the stator-rotor pair; T L is the load torque of the DC motor; J m is the equivalent rotational inertia on the shaft of the stator-rotor pair; B m is the viscous damping coefficient; K is the load spring stiffness; T a is the load torque of the stator-rotor pair; θ m is the rotation angle of the stator-rotor pair.

[0135] S2.4, the piston displacement tracking state space equation of the full hydraulic steering system based on hydraulic nonlinearity is described as follows:

[0136]

[0137] In the formula, the state variable is defined as

[0138] Step S3, the relationship between the DC motor voltage and the output shaft speed.

[0139] The DC motor model is as follows:

[0140]

[0141] It can be obtained that:

[0142]

[0143] In the formula, u v is the input voltage; C e is the electromotive force coefficient of the DC motor; C mT is the torque coefficient of the DC motor; m T is the motor time constant of the motor driving system; R is the total resistance of the armature circuit; L is the total inductance of the armature circuit; I' d T is the rate of change of armature current; a m T is the angular acceleration of the motor output shaft; T1 is the electromagnetic time constant of the armature circuit;

[0144] Step S4, according to the piston displacement tracking state space equation in step S2 and the difference between the target piston displacement and the actual piston displacement, a sliding mode controller is used to obtain the DC motor voltage. Specifically:

[0145] S4.1, according to the difference between the target piston displacement and the actual piston displacement, a sliding mode controller is used to obtain the DC motor voltage, and the construction process of the sliding mode controller is as follows:

[0146] 1) The piston displacement tracking state space equation of the full hydraulic steering system based on hydraulic nonlinearity is as shown in S2.4.

[0147] 2) The sliding surface and the exponential approach law are as follows:

[0148] s = c1e1 + c2e2 + e3 = c1(x 1d -x1) + c2(x 2d -x2) + (x 3d -x3) = 0 c > 0

[0149] s' = -εsgn(s) - ps ε > 0, p > 0

[0150] Wherein, c1, c2, c3 > 0, which satisfies the Hurwitz condition, c1, c2, c3 are sliding mode control parameters; ε, p are exponential approach law control parameters.

[0151] 3) The control law is as follows:

[0152]

[0153] Wherein,

[0154] S4.2, the motor output shaft speed calculation method is as follows:

[0155]

[0156] The application provides an electrically-controlled full-hydraulic steering system and a control method considering steering clearance and hydraulic nonlinearity, wherein the MPC trajectory tracker considers the characteristics of the steering system with the steering clearance, obtains the target front wheel steering angle considering the steering clearance through prediction, optimization and quadratic programming solution; the sliding mode controller obtains the piston displacement tracking state space equation according to the full-hydraulic steering system model considering the hydraulic nonlinearity, and then solves the direct-current motor voltage. The direct-current motor voltage is controlled to control the motor output shaft rotating speed, so that the vehicle steering is accurately controlled, and the problem of low tracking precision of the intelligent commercial vehicle based on the electrically-controlled full-hydraulic steering system is solved.

Claims

1. A control method of an electrically controlled full hydraulic steering system considering steering gap and hydraulic nonlinearities, comprising the following steps: Step S1, obtaining a target front wheel steering angle by using an MPC trajectory tracking controller according to a reference trajectory, a vehicle state and a steering gap, and converting the target front wheel steering angle into a target piston displacement; Step S2, obtaining a piston displacement tracking state space equation according to a full hydraulic steering dynamics model, wherein the full hydraulic steering dynamics model considers hydraulic nonlinearities; Step S3, obtaining a relationship between a direct current motor voltage and an output shaft speed according to a direct current motor model, a rotating shaft of the direct current motor (6) is connected with an input shaft of a stator-rotor pair (7), the stator-rotor pair (7) is connected with a full hydraulic steering valve (2); Step S4, obtaining the direct current motor voltage by using a sliding mode controller according to the piston displacement tracking state space equation in Step S2 and a difference between the target piston displacement and an actual piston displacement.

2. The control method of an electrically controlled full hydraulic steering system considering a turning gap and a hydraulic nonlinearity according to claim 1, characterized by, In Step S1, a target front wheel steering angle calculation method of the MPC trajectory tracking controller considering the steering gap is as follows: S1.1, establishing a discrete state equation of vehicle dynamics, as follows: where: x A (k) = [v y (k)ω r (k)ψ(k)y(k)] T , where ω r is the vehicle yaw rate; m is the vehicle mass; C f is the tire cornering stiffness of the front wheels; C r is the tire cornering stiffness of the rear wheels; a is the distance of the mass center from the front axle; b is the distance of the mass center from the rear axle; I z is the vehicle moment of inertia; δ f is the front wheel steering angle; v y is the lateral velocity; ψ is the heading angle; y is the lateral displacement; k is the k-th time instant; v x is the longitudinal velocity; x A is the state variable; δ c is the front wheel steering angle caused by the steering play; u A is the control variable; S1.2, constructing an objective function and constraint conditions of the MPC, The objective function is as follows: In the formula, r A is the control target expected value of the MPC controller; Q A is the MPC weight matrix; N p is the prediction step; N c is the control step; The constraint conditions are as follows: u min (k + j) < u A (k + j) < u max (k + j) < u c -1 Δu min (k+j)≤Δu A (k+j)≤Δu max (k+j)j = 0, 1,..., N c -1 y min (k+j)≤y A (k+j)≤y max (k+j)j=0,1,…,N p where u min , u max are upper and lower limits for the control variable; Δu min , Δu max are upper and lower limits for the control increment; y min , y max are upper and lower limits for the output variable; S1.3, obtaining the target front wheel steering angle by using a quadratic programming algorithm to solve the objective function, the process is as follows: δ f = [10...0] U A y d = i q δ f wherein U A is the target front wheel angle control sequence; i q is the transmission ratio between the front wheel and the piston rod; y d is the desired piston displacement; where H = Θ T Q A Θ + P A ; g = EQ A Θ; E = R A - Ψx(k) - ΘU c 3. The control method of an electrically controlled full hydraulic steering system considering a turning gap and a hydraulic nonlinearity according to claim 1, characterized by, In Step S2, the piston displacement tracking state space equation of the full hydraulic steering system based on hydraulic nonlinearities is as follows: my" = (p3 - p4)A - By' + f c + f(t, x1, x2) where m is the mass of the tire load; y is the second derivative of the piston displacement; p3, p4 are the pressures in the two chambers of the steering cylinder; A is the area of the cylinder piston; B is the viscous friction coefficient; y' is the first derivative of the piston displacement; f c is a constant disturbance; f(t, x1, x2) is a concentrated disturbance caused by other external disturbances and hydraulic nonlinearities; S2.1, a dynamics equation of the full hydraulic steering is as follows: p L = p3- p4 where V t is the total control volume of the steering cylinder; β e is the effective oil volume modulus; C t is the leakage coefficient of the steering cylinder; Q L is the average flow of the fixed rotor pair; Q c is the constant error; Q(t, x1, x2) is the time-varying error; S2.2, a pressure dynamics equation in a double-rod hydraulic cylinder is as follows: S2.3, an average flow calculation formula through the stator-rotor pair is as follows: T L = k m q(p1-p2) = J m θ" m + B m θ' m + Kθ m + T a where q is the theoretical displacement; ω m is the motor output shaft angular velocity; k m is the gear ratio; Q m1 , Q m2 are the stator-rotor pair inflow and outflow, respectively; C im , C cm are the stator-rotor pair internal and external leakage coefficients, respectively; p1, p2 are the stator-rotor pair inlet and outlet pressures, respectively; C lm is the stator-rotor pair total leakage coefficient; T L is the DC motor load torque; J m is the equivalent rotational inertia on the stator-rotor pair shaft; B m is the viscous damping coefficient; K is the load spring stiffness; T a is the stator-rotor pair load torque; θ m is the stator-rotor pair angle A balance equation of a stator-rotor pair output torque and a load torque is as follows: In the formula, the state variable is defined as 4. The control method of an electrically controlled full hydraulic steering system considering a turning gap and a hydraulic nonlinearity according to claim 1, characterized by, S2.4, a piston displacement tracking state space equation of the full hydraulic steering system based on hydraulic nonlinearities is described as follows: In Step S3, a relationship equation between the direct current motor voltage and the output shaft speed is as follows: A direct current motor model is as follows: wherein u v is the input voltage; C e is the electromotive force coefficient of the DC motor; C m is the torque coefficient of the DC motor; T m is the motor time constant of the motor drive system; R is the total armature circuit resistance; L is the total armature circuit inductance; I' d is the rate of change of armature current; a m is the angular acceleration of the motor output shaft; T1 is the electromagnetic time constant of the armature circuit; 5. The control method of an electrically controlled full hydraulic steering system considering a turning gap and a hydraulic nonlinearity according to claim 3, characterized by, It can be obtained that: In Step S4, a sliding mode control calculation method of the direct current motor voltage is as follows: S4.1, obtaining the direct current motor voltage by using a sliding mode controller according to a difference between the target piston displacement and the actual piston displacement, a construction process of the sliding mode controller is as follows: 1) constructing a piston displacement tracking state space equation of the full hydraulic steering system based on hydraulic nonlinearities as shown in S2.4, s = c1e1 + c2e2 + e3 = c1(x 1d - x1) + c2(x 2d - x2) + (x 3d - x3) = 0 c > 0 2) a sliding surface and an exponential reaching law are as follows: s'=-εsgn(s)-psε>0,p>0 In the formula, c1, c2, c3>0, satisfying the Hurwitz condition, c1, c2, c3 are sliding mode control parameters; ε, p are exponential reaching law control parameters, In the formulae, 3) a control law is as follows: S4.2, a motor output shaft speed calculation method is as follows: 6.An electrically controlled full hydraulic steering system considering steering gap and hydraulic nonlinearities, for realizing the control method of the electrically controlled full hydraulic steering system considering steering gap and hydraulic nonlinearities according to any one of claims 1 to 5, characterized in that, The electric control full hydraulic steering system comprises a full hydraulic steering module, wherein the full hydraulic steering module comprises a steering cylinder (3), a full hydraulic steering valve (2), a steering oil pump (1), a direct current motor (6), an oil tank (5), an intelligent control system (4) and a stator-rotor pair (7), wherein, The steering cylinder (3) is provided with a first oil port (3-1) and a second oil port (3-2); The full hydraulic steering valve (2) is provided with an oil inlet P1, an oil outlet T1, an oil port A, an oil port B, an oil port C and an oil port D; The stator-rotor pair (7) is connected with the oil port C and the oil port D of the full hydraulic steering valve (2); The oil port B of the full hydraulic steering valve (2) is connected with the second oil port (3-2) of the steering cylinder (3) through a hydraulic pipeline, the first oil port (3-1) of the steering cylinder (3) is connected with the oil port A of the full hydraulic steering valve (2) through a hydraulic pipeline, when the full hydraulic steering valve (2) moves to the right, the oil inlet P1 is connected with the oil port C, the oil liquid reaches the oil port D through the stator-rotor pair (7), and the oil port D is connected with the oil outlet T1, when the full hydraulic steering valve (2) moves to the left, the oil inlet P1 is connected with the oil port D, the oil liquid reaches the oil port C through the stator-rotor pair (7), and the oil port C is connected with the oil outlet T1; The steering oil pump (1) is connected with the oil tank (5), and the steering oil pump (1) supplies oil to the full hydraulic steering valve (2); The rotating shaft of the direct current motor (6) is connected with the input shaft of the stator-rotor pair (7); The stator-rotor pair (7) is connected with the full hydraulic steering valve (2); The signal input end of the direct current motor (6) is connected with the signal output end of the intelligent control system (4); The intelligent control system (4) comprises an MPC trajectory tracking module (41) considering a steering gap and a full hydraulic steering sliding mode control module (42) considering hydraulic nonlinearities; The MPC trajectory tracking module (41) considering the steering gap obtains a target front wheel steering angle by using an MPC-based trajectory tracking controller according to a planned path, a vehicle state and a steering gap; The full hydraulic steering sliding mode control module (42) considering hydraulic nonlinearities obtains a state equation of a tracked piston displacement according to a full hydraulic dynamics model considering hydraulic nonlinearities, and solves a direct current motor voltage by using a sliding mode controller.

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