A method for position axis control of an electro-hydraulic proportional servo valve considering output performance safety

Abstract

The present invention discloses a method for position axis control of an electro-hydraulic proportional servo valve taking into account output performance safety. The axis control method innovatively introduces a preset performance safety function, integrates the concept of dynamic surface control, and designs a nonlinear robust position axis control controller with output preset performance. Aiming at the problem of position axis control of an electro-hydraulic proportional servo valve, the present invention can not only realize preset control of the transient performance and steady-state performance of the system output, ensure safe and reliable operation of the system, but also avoid the differential explosion problem in the traditional backstepping control of the electro-hydraulic proportional servo valve axis control system, reduce the influence of measurement noise on control accuracy, and achieve high-precision tracking performance.

Description

Technical Field

[0001] The present invention relates to the technical field of electromechanical servo control, and in particular to an electro-hydraulic proportional servo valve position axis control method (OPSRC) taking output performance safety into consideration. Background Art

[0002] The electro-hydraulic proportional servo valve axis control system plays an important role in the fields of robots, heavy machinery, high-performance loading test equipment, etc. due to its high power density, large force / torque output, and fast dynamic response. The electro-hydraulic proportional servo valve axis control system is a typical nonlinear system, which contains many nonlinear characteristics and modeling uncertainties. The nonlinear characteristics include input nonlinearities such as hysteresis and saturation, proportional servo valve flow pressure nonlinearity, friction nonlinearity, etc. The modeling uncertainty includes parameter uncertainty and uncertainty nonlinearity, among which the parameter uncertainty mainly includes load mass, actuator viscous friction coefficient, leakage coefficient, servo valve flow gain, hydraulic oil elastic modulus, etc. The uncertainty nonlinearity mainly includes unmodeled friction dynamics, system high-order dynamics, external interference and unmodeled leakage. When the electro-hydraulic proportional servo valve axis control system develops towards high precision and high frequency response, the nonlinear characteristics of the system will have a more significant impact on the system performance, and the existence of modeling uncertainty will make the controller designed with the system nominal model unstable or downgraded. Therefore, the nonlinear characteristics and modeling uncertainty of the electro-hydraulic proportional servo valve axis control system are important factors that limit the improvement of system performance. With the continuous advancement of technology in the industrial and defense fields, the controllers designed based on traditional linear theory have gradually failed to meet the high performance requirements of the system. Therefore, it is necessary to study more advanced nonlinear control strategies based on the nonlinear characteristics of the electro-hydraulic proportional servo valve axis control system.

[0003] Many methods have been proposed to solve the nonlinear control problem of electro-hydraulic proportional servo valve axis control system. Among them, the adaptive control method is a very effective method for dealing with parameter uncertainty problems, and can obtain steady-state performance of asymptotic tracking. However, it is powerless for uncertainty nonlinearity such as external load interference. When the uncertainty nonlinearity is too large, the system may become unstable. The actual electro-hydraulic proportional servo valve axis control system has uncertainty nonlinearity, so the adaptive control method cannot obtain high-precision control performance in practical applications; as a robust control method, classical sliding mode control can effectively deal with any bounded modeling uncertainty and obtain steady-state performance of asymptotic tracking, but the discontinuous controller designed by classical sliding mode control is prone to cause the chatter problem of the sliding surface, thereby deteriorating the tracking performance of the system; in order to solve the problems of parameter uncertainty and uncertainty nonlinearity at the same time, the adaptive robust control method is proposed. This control method can enable the system to obtain certain transient and steady-state performance when the two modeling uncertainties exist at the same time. If high-precision tracking performance is to be obtained, the feedback gain must be increased to reduce the tracking error. Due to the existence of measurement noise, the gain is too large, which often leads to high-gain feedback, thereby causing chattering of the control input, thereby deteriorating the control performance and even causing system instability. Summary of the invention

[0004] The purpose of the present invention is to provide an electro-hydraulic proportional servo valve position axis control method with safe limited system output performance, strong anti-interference ability and high tracking performance, which can not only realize the preset control of the system output transient performance and steady-state performance, ensure the safe and reliable operation of the system, but also avoid the differential explosion problem in the traditional backstepping control of the electro-hydraulic proportional servo valve axis control system, reduce the influence of measurement noise on control accuracy, and achieve high-precision tracking performance.

[0005] The technical solution to achieve the purpose of the present invention is: a method for controlling the position axis of an electro-hydraulic proportional servo valve taking into account the safety of output performance, comprising the following steps:

[0006] Step 1: Establish a mathematical model of the electro-hydraulic proportional servo valve position axis control system, and then proceed to step 2.

[0007] Step 2: Based on the mathematical model of the electro-hydraulic proportional servo valve position axis control system, design a nonlinear robust position axis control controller with output preset performance, and then proceed to step 3.

[0008] Step 3: Use Lyapunov stability theory to prove the stability of the nonlinear robust position axis control controller and obtain the result that the system tracking error is asymptotically stable.

[0009] Compared with the prior art, the present invention has the following significant advantages: (1) it realizes preset control of the transient performance and steady-state performance of the system output, ensuring safe and reliable operation of the system; (2) it avoids the differential explosion problem in the traditional backstepping control of the electro-hydraulic proportional servo valve axis control system, reduces the influence of measurement noise on control accuracy, and achieves high-precision tracking performance. The simulation results verify its effectiveness. BRIEF DESCRIPTION OF THE DRAWINGS

[0010] Figure 1 It is a schematic diagram of the principle of the electro-hydraulic proportional servo valve position axis control method considering output performance safety of the present invention.

[0011] Figure 2 It is a schematic diagram of the principle of the axis control system of the electro-hydraulic proportional servo valve of the present invention.

[0012] Figure 3 It is a curve diagram of the tracking process of the system output to the expected instruction under the action of the OPSRC controller designed by the present invention.

[0013] Figure 4 It is a curve diagram showing the tracking error of the system changing with time under the action of the OPSRC controller designed by the present invention.

[0014] Figure 5 It is a tracking error comparison curve diagram of the system under the action of the OPSRC controller designed by the present invention and the traditional PID controller.

[0015] Figure 6 It is a control input curve diagram of the system under the action of the OPSRC controller designed by the present invention. DETAILED DESCRIPTION

[0016] The present invention is further described in detail below with reference to the accompanying drawings and specific embodiments.

[0017] Combination Figure 1 and Figure 2 The present invention considers the output performance safety of the electro-hydraulic proportional servo valve position axis control method, comprising the following steps:

[0018] Step 1: Establish a mathematical model of the electro-hydraulic proportional servo valve position axis control system.

[0019] Step 1-1, the electro-hydraulic proportional servo valve position axis control system is applied to the linear motion of large industrial heavy-load mechanical equipment, wherein the load is fixedly connected to the piston rod on the hydraulic cylinder, and the electro-hydraulic proportional servo valve controls the movement of the piston rod on the hydraulic cylinder, thereby driving the load to move;

[0020] According to Newton's second law, the force balance equation of the electro-hydraulic proportional servo valve position axis control system is:

[0021]

[0022] In formula (1), m represents the mass of the load, y represents the displacement of the hydraulic cylinder piston rod, Indicates the speed of the hydraulic cylinder piston rod. represents the acceleration of the hydraulic cylinder piston rod, A represents the effective area of ​​the hydraulic cylinder piston, P 1 Indicates the oil pressure in the hydraulic cylinder oil inlet chamber, P 2 represents the oil pressure of the hydraulic cylinder oil outlet chamber, B represents the viscous damping coefficient of the hydraulic cylinder, d 1 (t) represents the unmodeled mechanical disturbance of the system, and t represents time.

[0023] Then formula (1) can be rewritten as:

[0024]

[0025] In the electro-hydraulic proportional servo valve position axis control system, ignoring the leakage of the cylinder oil, the pressure dynamic equation is:

[0026]

[0027] In formula (3), β e Indicates the effective elastic modulus of the oil, C t Indicates the leakage coefficient of the hydraulic cylinder, the oil pressure difference P on both sides of the cylinder in and out of the oil chamber L =P 1 -P 2 , the control volume V of the oil inlet chamber 1 =V 01 +Ay, the controlled volume V of the oil outlet cavity 2 =V 02 -Ay, V 01 Represents the initial volume of the oil inlet chamber, V 02 Indicates the initial volume of the oil chamber, Q 1 Indicates the flow rate of the oil inlet cavity, Q 2 Indicates the flow rate of the oil chamber, q 1 Indicates P 1 The unmodeled disturbance, q 2 Indicates P 2 The unmodeled disturbance Indicates P 1 The first derivative of Indicates P 2 The first derivative of .

[0028] Q 1 , Q 2 Respectively with the displacement x of the electro-hydraulic proportional servo valve core v There are the following relationships:

[0029]

[0030] Among them, the electro-hydraulic proportional servo valve coefficient C d Indicates the flow coefficient of the electro-hydraulic proportional servo valve, w 0 represents the valve core area gradient of the electro-hydraulic proportional servo valve, ρ represents the oil density, P s Indicates the oil supply pressure, P r represents the return oil pressure, s(·) represents the function of the intermediate variable ·, and is defined as:

[0031]

[0032] Ignoring the dynamics of the electro-hydraulic proportional servo valve spool, assume that the control input u acting on the spool and the spool displacement x v Proportional relationship, that is, satisfying x v =k i u, where k i represents the voltage-valve core displacement gain coefficient, so equation (4) is rewritten as:

[0033]

[0034] Formula (6), intermediate variable k u =k q k i , intermediate variable Intermediate variables

[0035] Step 1-2, define state variables: Among them, the intermediate variable x 1 =y, intermediate variable Intermediate variable x 3 =(AP 1 -AP 2 ) / m, then transform equation (2) into the state equation:

[0036]

[0037] Formula (7), Represents x 1 The first derivative of Represents x 2 The first derivative of Represents x 3 The first derivative of the system is unknown dynamics Intermediate variables Intermediate variables Intermediate variables System unknown dynamics

[0038] To facilitate the design of the controller, the following assumptions are made:

[0039] Assumption 1: The system is expected to track the position command x d It is second-order continuous, and the system expects that the position command, velocity command and acceleration command are all bounded;

[0040] Assumption 2: The system has unknown dynamics D 1 With D 2 satisfy:

[0041]

[0042] Formula (8), δ 1 and δ 2 are all unknown positive constants;

[0043] Go to step 2.

[0044] Step 2: Based on the mathematical model of the electro-hydraulic proportional servo valve position axis control system, a nonlinear robust position axis control controller with output preset performance is designed. The specific steps are as follows:

[0045] Step 2-1. Define error ξ 1 =z 1 / ρ 1 (t), where z 1 =x 1 -x d represents the system tracking error, x d is the position command that the system expects to track, ρ 1 (t) represents the preset performance safety function;

[0046] Tracking error z 1 Meet the preset range:

[0047] -ρ 1 (t)<z 1 <ρ 1 (t) (9),

[0048] Formula (9), preset performance safety function ρ 1 (t)=(ρ 0 -ρ ∞ ) -ct +ρ ∞ , where ρ 0 Represents a positive constant, ρ ∞ represents a constant that is always positive, c represents a constant that is always positive, and satisfies the condition ρ 0 >ρ ∞ .

[0049] In order to facilitate the design of the controller, the present invention innovatively proposes the following nonlinear filter:

[0050]

[0051] Formula (10), filter gain τ 1 >0,α 1 Represents x 2 Virtual control, α 1f Represents α 1 The filtered signal, α 1f With x 2 The error z 2 =x 2 -α 1f , α 1 The filtering error ε 1 =α 1f -α 1 , σ 1 (t) represents a function that is always positive and satisfies Where ν represents the integral variable, represents a positive constant, Represents α 1 The first derivative of Represents α 1f The first derivative of 1 >0 means The upper bound of .

[0052] Yes 1 Taking the derivative, we get:

[0053]

[0054] Formula (11), Represents the preset performance safety function ρ 1 The first derivative of (t).

[0055] Select the Lyapunov function We can get:

[0056]

[0057] Design Virtual Control α 1 for:

[0058]

[0059] Formula (13), gain k 1 >0, then

[0060]

[0061] Step 2-2, innovatively propose the following nonlinear filter:

[0062]

[0063] Formula (15), filter gain τ 2 >0,α2 Represents x 3 Virtual control, α 2f Represents α 2 The filtered signal, α 2f With x 3 The error z 3 =x 3 -α 2f , α 2 The filtering error ε 2 =α 2f -α 2 , σ 2 (t) represents a function that is always positive and satisfies Where ν represents the integral variable, represents a positive constant, Represents α 2 The first derivative of Represents α 2f The first derivative of 2 >0 means The upper bound of .

[0064] Yes 2 The derivative is:

[0065]

[0066] Select the Lyapunov function We can get:

[0067]

[0068] Design Virtual Control α 2 for:

[0069]

[0070] Formula (18), gain k 2 >0,α 2a represents the model-based compensation term, α 2s represents the robust term, α 2s1 represents the linear robust term, α 2s2 represents the nonlinear robust term, σ 3 (t) represents a function that is always positive and satisfies Where ν represents the integral variable, Represents a positive constant, l 3 >0 means D 1 The upper bound of .

[0071] Substituting formula (18) into formula (17), we get:

[0072]

[0073] Step 2-3, 3 The derivative is:

[0074]

[0075] Select the Lyapunov function We can get:

[0076]

[0077] According to formula (21), the control input of the valve core, that is, the nonlinear robust position axis control controller u considering the output preset performance safety, is:

[0078]

[0079] Formula (22), gain k 3 >0,u a represents the model-based compensation term, u s represents the robust term, u s1 represents the linear robust term, u s2 represents the nonlinear robust term, σ 4 (t) represents a function that is always positive and satisfies Where ν represents the integral variable, Represents a positive constant, l 4 >0 means D 2 The upper bound of .

[0080] Substituting formula (22) into formula (21), we get:

[0081]

[0082] Go to step 3.

[0083] Step 3, using Lyapunov stability theory to prove the stability of the nonlinear robust position axis control controller, the result of asymptotic stability of the system tracking error is obtained, as follows:

[0084] The Lyapunov function is defined as follows:

[0085]

[0086] Deriving equation (24) and substituting equation (10), equation (15) and equation (23) into it, we can obtain:

[0087]

[0088] Considering |D 1 |≤l 3 and |D 2 |≤l 4, we can get the expression:

[0089]

[0090] Notice

[0091]

[0092] Substituting equation (27) into equation (26), we can obtain

[0093]

[0094] Formula (28), intermediate variable Intermediate variable Ψ=σ 1 (t)+σ 2 (t)+σ 3 (t)+σ 4 (t).

[0095] Integrating both sides of equation (28) we can obtain:

[0096]

[0097] Formula (29), intermediate variable

[0098] From equation (29), we can see that V is bounded, Φ is integrally bounded, and then we can conclude that all signals in the system are bounded, so Φ is uniformly continuous. According to Barbalat’s lemma, when time tends to positive infinity, the tracking error z 1 Tends to 0.

[0099] Therefore, it is concluded that by adjusting the gain k 1 , k 2 , k 3 and filter gain τ 1 , τ 2 The nonlinear robust position axis control controller designed for the electro-hydraulic proportional servo valve position axis control system with preset output performance and safety can enable the system to innovatively obtain the result that the tracking error converges to 0 asymptotically. The principle diagram of the nonlinear robust position axis control controller for the electro-hydraulic proportional servo valve position axis control system is shown in the figure. Figure 1 shown.

[0100] Example

[0101] In order to evaluate the performance of the designed controller, the physical parameters of the electro-hydraulic proportional servo valve position axis control system in the simulation are shown in Table 1:

[0102] Table 1 System physical parameters

[0103] Physical parameters Numeric Physical parameters Numeric <![CDATA[A(m 2 )]]> <![CDATA[2×10 -4 ]]> <![CDATA[β e (Well)]]> <![CDATA[2×10 8 ]]> m(kg) 40 B(N·s / m) 80 <![CDATA[C t (m 5 / (N·s))]]> <![CDATA[7×10 -12 ]]> <![CDATA[k u (m / V)]]> <![CDATA[4×10 -8 ]]> <![CDATA[V 01 (m 3 )]]> <![CDATA[1×10 -3 ]]> <![CDATA[V 02 (m 3 )]]> <![CDATA[1×10 -3 ]]> <![CDATA[P s (MPa)]]> 7 <![CDATA[P r (MPa)]]> 0

[0104] Given a system with the expected instruction x d =0.02sin(πt)×(1-e -t )m.

[0105] The following controllers are used for comparison in the simulation:

[0106] Considering the output performance safety of the electro-hydraulic proportional servo valve position axis control controller (OPSRC): take the gain k 1 =100, k 2 =50, k 3 =20,γ 1 =0.01,γ 2 =1,γ 3 =1,γ 4 =1,τ 1 =2000,τ 2 =2000,ρ 1 (t) = 0.01e -1.1t +0.001.

[0107] PID controller: The steps for selecting PID controller parameters are: first, ignoring the nonlinear dynamics of the electro-hydraulic proportional servo valve axis control system, a set of controller parameters are obtained through the PID parameter self-tuning function in Matlab, and then the self-tuning parameters are fine-tuned after adding the nonlinear dynamics of the system to obtain the best tracking performance. The selected controller parameters are k P =1000, k I =500, k D =1.

[0108] The expected command of the system, the tracking error of the OPSRC controller, and the tracking error comparison between the OPSRC controller and the PID controller are shown as follows: Figure 3 , Figure 4 and Figure 5 As shown. Figure 4 It can be seen that under the action of the OPSRC controller, the position output of the proportional servo valve axis control system has a high tracking accuracy for the command, and the amplitude of the steady-state tracking error is about 5.4×10 -4 m. From Figure 5 The comparison of the tracking errors of the two controllers shows that the tracking error of the OPSRC controller proposed in the present invention is much smaller than that of the PID controller, and the tracking performance is more superior.

[0109] Figure 6 It is a curve diagram of the control input of the electro-hydraulic proportional servo valve axis control system changing with time under the action of the OPSRC controller. It can be seen from the figure that the obtained control input is a low-frequency continuous signal, which is more conducive to execution in practical applications.

Claims

1. A method for controlling the position axis of an electro-hydraulic proportional servo valve taking into account output performance safety, characterized in that: The following steps are involved: Step 1: Establish a mathematical model of the electro-hydraulic proportional servo valve position axis control system, as follows: Step 1-1, the electro-hydraulic proportional servo valve position axis control system is applied to the linear motion of large industrial heavy-duty mechanical equipment, in which the load is fixedly connected to the piston rod on the hydraulic cylinder, and the electro-hydraulic proportional servo valve controls the movement of the piston rod on the hydraulic cylinder, thereby driving the load to move. According to the dynamic characteristics of the load, hydraulic cylinder and electro-hydraulic proportional servo valve, the mathematical model of the electro-hydraulic proportional servo valve position axis control system is obtained; the details are as follows: According to Newton's second law, the force balance equation of the electro-hydraulic proportional servo valve position axis control system is: In formula (1), m represents the mass of the load, y represents the displacement of the hydraulic cylinder piston rod, Indicates the speed of the hydraulic cylinder piston rod. represents the acceleration of the piston rod of the hydraulic cylinder, A represents the effective action area of ​​the piston of the hydraulic cylinder, P1 represents the oil pressure of the oil inlet chamber of the hydraulic cylinder, P2 represents the oil pressure of the oil outlet chamber of the hydraulic cylinder, B represents the viscous damping coefficient of the hydraulic cylinder, d1(t) represents the unmodeled mechanical disturbance of the system, and t represents time; Then formula (1) can be rewritten as: In the electro-hydraulic proportional servo valve position axis control system, ignoring the leakage of the cylinder oil, the pressure dynamic equation is: In formula (3), β e Indicates the effective elastic modulus of the oil, C t Indicates the leakage coefficient of the hydraulic cylinder, the oil pressure difference P on both sides of the cylinder in and out of the oil chamber L =P1-P2, the control volume of the oil inlet chamber V1 = V 01 +Ay, the control volume of the oil outlet chamber V2 = V 02 -Ay, V 01 Represents the initial volume of the oil inlet chamber, V 02 represents the initial volume of the oil outlet chamber, Q1 represents the flow rate of the oil inlet chamber, Q2 represents the flow rate of the oil outlet chamber, q1 represents the unmodeled disturbance of P1, q2 represents the unmodeled disturbance of P2, represents the first-order derivative of P1, represents the first derivative of P2; Q1 and Q2 are respectively related to the displacement x of the valve core of the electro-hydraulic proportional servo valve v There are the following relationships: Among them, the electro-hydraulic proportional servo valve coefficient C d represents the flow coefficient of the electro-hydraulic proportional servo valve, w0 represents the valve core area gradient of the electro-hydraulic proportional servo valve, ρ represents the oil density, P s Indicates the oil supply pressure, P r represents the return oil pressure, s(·) represents the function of the intermediate variable ·, and is defined as: Ignoring the dynamics of the electro-hydraulic proportional servo valve spool, assume that the control input u acting on the spool and the spool displacement x v Proportional relationship, that is, satisfying x v =k i u, where k i represents the voltage-valve core displacement gain coefficient, so equation (4) is rewritten as: Formula (6), intermediate variable k u =k q k i , intermediate variable Intermediate variables Step 1-2: To facilitate the design of the controller, define the state variables and convert the mathematical model of the electro-hydraulic proportional servo valve position axis control system into a state space equation, as follows: Define the state variables: Among them, the intermediate variable x1 = y, the intermediate variable The intermediate variable x3 = (AP1-AP2) / m, then the equation (2) is converted into a state space equation: Formula (7), represents the first-order derivative of x1, represents the first-order derivative of x2, represents the first-order derivative of x3, the unknown dynamics of the system Intermediate variables Intermediate variables Intermediate variables System unknown dynamics T stands for transpose; To facilitate the design of the controller, the following assumptions are made: Assumption 1: The system is expected to track the position command x d It is second-order continuous, and the system expects that the position command, velocity command and acceleration command are all bounded; Assumption 2: The unknown dynamics D1 and D2 of the system satisfy: In formula (8), δ1 and δ2 are both unknown positive constants; Go to step 2; Step 2, based on the mathematical model of the electro-hydraulic proportional servo valve position axis control system, a nonlinear robust position axis control controller with output preset performance is designed, as follows: Step 2-1, define the error ξ1 = z1 / ρ1(t), where z1 = x1-x d represents the system tracking error, x d is the system's expected tracking position command, ρ1(t) represents the preset performance safety function. In order to facilitate the system state x1 to track the expected position command x as accurately as possible under the designed controller drive, d , and let z1 always satisfy |z1|<ρ1(t), and ensure that the tracking error ξ1 tends to 0; The tracking error z1 meets the preset range: -ρ1(t)<z1<ρ1(t) (9), In formula (9), the preset performance safety function ρ1(t)=(ρ0-ρ ∞ ) -ct +ρ ∞ , where ρ0 is a positive constant and ρ ∞ represents a constant that is always positive, c represents a constant that is always positive, and satisfies the condition ρ0>ρ ∞ ; To facilitate the design of the controller, the following nonlinear filter is designed: Formula (10), filter gain τ1>0, filter error ε1=α 1f -α1,σ1(t) represents a function that is always positive and satisfies Where ν represents the integral variable, represents a positive constant, represents the first-order derivative of α1, Represents α 1f The first derivative of The upper bound of l1>0; Taking the derivative of the tracking error ξ1, we get: Among them, ξ2 represents the error; Represents x d The first derivative of ; represents the first-order derivative of the preset performance safety function ρ1(t); Select the Lyapunov function We can get: Design the virtual control α1 as: Formula (13), gain k1>0, then Step 2-2: Define error ξ2 = x2 - α 1f , where α 1f Represents the filtered signal of α1, α1 represents the virtual control of x2. To ensure that the tracking error ξ1 tends to 0, it is necessary to ensure that the error ξ2 tends to 0; the details are as follows: Design the following nonlinear filter: Formula (16), filter gain τ2>0, filter error ε2=α 2f -α2,σ2(t) represents a function that is always positive and satisfies Where ν represents the integral variable, represents a positive constant, represents the first-order derivative of α2, Represents α 2f The first derivative of The upper bound of l2>0; Taking the derivative of ξ2, we get: Select the Lyapunov function We can get: Design the virtual control α2 as: Formula (20), gain k2>0, α 2a represents the model-based compensation term, α 2s represents the robust term, α 2s1 represents the linear robust term, α 2s2 represents a nonlinear robust term, σ3(t) represents a function that is always positive and satisfies Where ν represents the integral variable, Represents a constant that is always positive, the upper bound of D1 is l3>0; Substituting formula (18) into formula (17), we get: Step 2-3: Define error ξ3 = x3-α 2f , where α 2f represents the filtered signal of α2, α2 represents the virtual control of x3. To ensure that the error ξ2 tends to 0, the error ξ3 needs to tend to 0, as follows: Taking the derivative of ξ3, we get: Select the Lyapunov function have to: According to formula (21), the control input of the valve core, that is, the nonlinear robust position axis control controller u considering the output preset performance safety, is: Formula (22), gain k3>0, u a represents the model-based compensation term, u s represents the robust term, u s1 represents the linear robust term, u s2 represents the nonlinear robust term, σ4(t) represents a function that is always positive and satisfies Where ν represents the integral variable, Represents a constant that is always positive, the upper bound of D2 is l4>0; Substituting formula (22) into formula (21), we get: Go to step 3; Step 3, use Lyapunov stability theory to prove the stability of the nonlinear robust position axis control controller, and obtain the result that the system tracking error is asymptotically stable.

2. The method for position axis control of an electro-hydraulic proportional servo valve considering output performance safety according to claim 1, characterized in that: The stability of the nonlinear robust position axis control controller is proved by using Lyapunov stability theory as described in step 3, and the result that the system tracking error is asymptotically stable is obtained, as follows: The Lyapunov function is defined as follows: The stability is proved by using Lyapunov stability theory, and the result that the system tracking error is asymptotically stable is obtained.

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