Electro-hydraulic execution system with oil return cavity pressure regulation function and control method of electro-hydraulic execution system

Abstract

The invention provides an electro-hydraulic execution system with an oil return cavity pressure adjusting function and a control method thereof. The control method comprises the steps that the reference compensation pressure of an oil return cavity is determined by adopting an expansion state observer principle; the reference pressure of the oil return cavity is determined based on the reference compensation pressure and the preset pressure; based on the pressure of the pump control cavity, the pressure of the oil return cavity, the outlet pressure of the metering pump and the pressure of the oil tank, the front-back pressure difference of the valve body is determined; determining the preliminary control voltage of the proportional valve based on the front-back pressure difference of the valve body and the flow of the proportional valve by adopting a valve body flow formula; the final control voltage of the proportional valve is determined by combining the PID feedback device and the preliminary control voltage of the proportional valve; and controlling the proportional valve based on the proportional valve final control voltage. According to the control method, through oil return cavity reference pressure and combination of flow feed-forward based on a valve model and PID feedback, main disturbance is compensated through feed-forward control, other disturbance is restrained through closed-loop feedback, and the reverse elastic load disturbance restraining performance of a system is improved.

Description

Technical Field

[0001] This application relates to the field of electro-hydraulic servo control technology, and in particular to an electro-hydraulic actuator system with return oil chamber pressure regulation and its control method. Background Technology

[0002] Permanent magnet synchronous motor (PMSM) drives are used across various industrial sectors, including electromechanical actuators and electrohydraulic actuators. Compared to electromechanical actuators, electrohydraulic actuators can apply greater driving forces and absorb shocks and collisions, offering significant advantages in terms of high power density, reliability, and robustness.

[0003] Independent electro-hydraulic actuators offer higher power-to-weight ratios, better maintainability, and greater robustness. However, electro-hydraulic servo systems are typically nonlinear systems, exhibiting numerous nonlinear characteristics and model uncertainties. Furthermore, when the hydraulic cylinder actuator is subjected to a reverse elastic load disturbance, if the actuator lacks sufficient back pressure and return oil damping, the uncertain reverse elastic load disturbance may trigger cavitation, air intake, actuator vibration, or even motion control instability, leading to a degraded system performance. Summary of the Invention

[0004] This application provides an electro-hydraulic actuator system with return oil chamber pressure regulation and its control method to solve the technical problem of poor performance of existing electro-hydraulic systems in suppressing reverse elastic load disturbances.

[0005] The control method for an electro-hydraulic actuator system with return oil chamber pressure regulation provided in this application is applied to such an system. The electro-hydraulic actuator system includes an oil tank, a pump-controlled direct drive subsystem, a valve body, an elastic load, and a hydraulic cylinder. The pump-controlled direct drive subsystem includes a permanent magnet synchronous motor and a fixed displacement pump. The valve body includes a proportional valve and a switching valve. The proportional valve includes a first proportional valve and a second proportional valve. The switching valve includes a first switching valve and a second switching valve. The fixed displacement pump is connected to the oil tank, the permanent magnet synchronous motor, the first switching valve, and the second switching valve. The piston of the hydraulic cylinder is connected to the elastic load. The first and second proportional valves are connected to the return oil chamber and the pump-controlled chamber of the hydraulic cylinder, respectively. The first and second switching valves are connected to the first and second proportional valves, respectively. The first proportional valve, the second proportional valve, the first switching valve, and the second switching valve form an H-bridge structure. The control method includes: determining the reference compensation pressure of the return oil chamber using the principle of an expansion state observer. Based on reference compensation pressure and the preset pressure of the return oil chamber Determine the reference pressure of the return oil chamber. Based on the pressure of the pump control chamber Pressure of the return oil chamber outlet pressure of the metering pump and the pressure of the fuel tank Determine the pressure difference across the valve body. Among them, the pressure difference across the valve body The operating modes of the electro-hydraulic actuator system are related to its operation. The system includes a first operating mode and a second operating mode. In the first operating mode, the first switching valve is open and the second switching valve is closed; in the second operating mode, the first switching valve is closed and the second switching valve is open. A valve body flow formula is used, based on the pressure difference across the valve body. and the flow rate of the proportional valve Determine the initial control voltage of the proportional valve The initial control voltage is achieved by combining a PID feedback circuit and a proportional valve. Determine the final control voltage of the proportional valve. Reference pressure of the oil return chamber The reference input for the PID feedback; the final control voltage based on the proportional valve. Control proportional valve.

[0006] In some feasible implementations, the extended state observer principle is used to determine the reference compensation pressure. This includes: establishing the state-space equations of the electro-hydraulic actuator; and using the state-space equations combined with the extended state observer principle to determine the estimated elastic load disturbance of the electro-hydraulic actuator. After the elastic load is disturbed, the pressure in the pump control chamber... Unchanged, elastic load force Under the condition that remains unchanged, combined with the estimated elastic load disturbance Determine reference compensation pressure .

[0007] In some feasible implementations, the state-space equations of the electro-hydraulic actuator system are established, including: the state-space equations are as follows: ;in, Let be the state variables in the state-space equation. for The derivative; These are the output variables of the state-space equations; The input vector; ; The system matrix is ​​the state-space equation. The control matrix of the state-space equation; This is the output matrix of the state-space equations; Let be the perturbation matrix of the state-space equation. To observe elastic load disturbances, for The derivative; , ; ; is the piston effective area of the pump chamber of the hydraulic cylinder, is the mass of the elastic load, is the viscous damping coefficient, is the elastic coefficient of the elastic load; is the piston rod displacement, is the piston rod velocity; is the second linear extended state observer coefficient, is the third linear extended state observer coefficient, is the fourth linear extended state observer coefficient.

[0008] In some possible implementation manners, the estimated elastic load disturbance of the electro-hydraulic execution system is determined by adopting a state space equation combined with an extended state observer principle , and the method comprises: controlling a state space model based on an extended state observer principle to obtain an extended state equation; wherein the extended state equation comprises: ; is the observed piston rod displacement; is the derivative of ; is the observed piston rod velocity; is the derivative of ; is the observed disturbance; is the derivative of ; is the first gain coefficient; is the second gain coefficient; is the third gain coefficient; is the first linear extended state observer coefficient; , is the piston effective area of the return oil chamber of the hydraulic cylinder; the elastic load disturbance is estimated in real time based on the extended state equation, and the estimated elastic load disturbance is obtained.

[0009] In some possible implementation manners, the elastic load disturbance is estimated in real time based on the extended state equation, and the estimated elastic load disturbance is obtained, and the method comprises: placing poles of the extended state equation at the same position to obtain a characteristic polynomial ; wherein the characteristic polynomial is: ; I is a unit matrix, is the bandwidth of the LESO, and s is a complex variable in a complex frequency domain; the estimated elastic load disturbance is obtained based on the characteristic polynomial Second gain coefficient and the third gain coefficient Determine the estimated elastic load disturbance ; where the first gain coefficient Second gain coefficient and the third gain coefficient satisfy Estimate elastic load disturbance satisfy: ; The coefficients of the fifth linearly extended observer. .

[0010] Among some feasible implementation methods, based on reference compensation pressure and the preset pressure of the return oil chamber Determine the reference pressure of the return oil chamber. ,include: ;in, This is the disturbance coefficient for elastic loads. .

[0011] In some feasible implementations, based on pump control chamber pressure Pressure of the return oil chamber outlet pressure of the metering pump and tank pressure Determine the pressure difference across the valve body. This includes: in the first operating mode, the pressure difference across the valve body. Including the pressure difference across the first switching valve Pressure difference across the second proportional valve Pressure difference across the first switching valve Pressure difference across the second proportional valve satisfy: In the second operating mode, the pressure difference across the valve body... Including the pressure difference across the second switching valve Pressure difference across the first proportional valve Pressure difference across the second switching valve Pressure difference across the second proportional valve satisfy: .

[0012] In some feasible implementations, a valve body flow formula is used, based on the pressure difference across the valve body. and proportional valve flow Determine the initial control voltage of the proportional valve This includes: the initial control voltage of the proportional valve. satisfy: Where i is 3 or 4; S is the maximum cross-sectional area of ​​the throttling orifice of the proportional valve. It is the normalization coefficient for the valve control voltage to the valve core position; The maximum flow coefficient; It refers to the density of the oil.

[0013] In some feasible implementations, a combination of PID feedback and a proportional valve is used to initially control the voltage. Determine the final control voltage of the proportional valve This includes: the final control voltage of the proportional valve. satisfy: ; in, = , To control the deviation signal, This is the proportionality coefficient. The integral time constant is... The differential time constants; t0 and t i These represent different times.

[0014] The control method for an electro-hydraulic actuator system with return chamber pressure regulation provided in the first aspect of this application utilizes a pump-valve composite system combining a variable speed pump control system and valve control technology, designing the pump control system and valve control system separately. To ensure system control accuracy, the pump control strategy adopts an adaptive model predictive pressure control (FFRLS-AMPC) strategy with a forgetting factor and recursive least squares, fully leveraging the energy-saving advantages of variable speed pump control and the improved dynamic response of adaptive model predictive control. The return chamber valve control pressure strategy consists of a pressure planner and a pressure controller. The pressure planner outputs the return chamber reference pressure through a small preset pressure and a linearly extended state observer (LESO). The pressure controller combines flow feedforward based on the valve model with proportional-integral-derivative (PID) feedback, providing both feedforward control to compensate for major disturbances and closed-loop feedback to suppress other disturbances, thereby improving the system's performance in suppressing reverse elastic load disturbances.

[0015] The second aspect of this application provides an electro-hydraulic actuator system with return oil chamber pressure regulation, comprising: a first determining module configured to determine a reference compensation pressure of the return oil chamber using the principle of an expansion state observer. The second determining module is configured to be based on reference compensation pressure. and the preset pressure of the return oil chamber Determine the reference pressure of the return oil chamber. The third determining module is configured to be based on the pressure of the pump control chamber. Pressure of the return oil chamber outlet pressure of the metering pump and the pressure of the fuel tank Determine the pressure difference across the valve body. Among them, the pressure difference across the valve body In relation to the working mode of the electro-hydraulic execution system, the working mode of the electro-hydraulic execution system includes a first working mode and a second working mode, in the first working mode, the first switch valve is opened and the second switch valve is closed; in the second working mode, the first switch valve is closed and the second switch valve is opened; the fourth determination module is configured to use a valve body flow formula to determine the reference pressure of the oil return cavity based on the pressure difference before and after the valve body and the flow of the proportional valve determine the preliminary control voltage of the proportional valve ; the fifth determination module is configured to combine the PID feedback and the preliminary control voltage of the proportional valve determine the final control voltage of the proportional valve ; the reference pressure of the oil return cavity is the reference input of the PID feedback; the control module is configured to control the proportional valve based on the final control voltage of the proportional valve.

[0016] The second aspect of the present application provides an electro-hydraulic execution system with oil return cavity pressure regulation, which adopts the control method of the electro-hydraulic execution system with oil return cavity pressure regulation provided in the first aspect, and has the beneficial technical effects of the first aspect, which will not be repeated here. BRIEF DESCRIPTION OF DRAWINGS

[0017] In order to more clearly illustrate the technical solutions of the present application, the drawings needed in the embodiments will be briefly introduced below, and obviously, other drawings can also be obtained by those skilled in the art without creative labor on the basis of these drawings.

[0018] Figure 1 is a structural schematic diagram of an electro-hydraulic execution system with oil return cavity pressure regulation provided by the embodiments of the present application; Figure 2 is a general control block diagram of an electro-hydraulic execution system with oil return cavity pressure regulation provided by the embodiments of the present application; Figure 3 is a block schematic diagram of an electro-hydraulic execution system with oil return cavity pressure regulation provided by the embodiments of the present application; Figure 4 is a valve control pressure block diagram of an oil return cavity provided by the embodiments of the present application; Figure 5 is a flowchart of a control method of an electro-hydraulic execution system with oil return cavity pressure regulation provided by the embodiments of the present application; Figure 6 is a linear extended observer principle diagram; Figure 7 is a flow coefficient and Reynolds number relationship curve; Figure 8 ​is a structural block diagram of an electro-hydraulic execution system with oil return cavity pressure regulation provided by the embodiment of the present application.

[0019] Illustration mark: 1-elastic load; 2-hydraulic cylinder; 21-oil return cavity; 22-pump control cavity; 3-proportional valve; V3-first proportional valve; V4-second proportional valve; 4-on-off valve; first on-off valve V1; second on-off valve V2; 5-constant displacement pump; 6-oil tank; 7-permanent magnet synchronous motor; 100-electro-hydraulic execution system; 101-first determining module; 102-second determining module; 103-third determining module; 104-fourth determining module; 105-fifth determining module; 106-control module. DETAILED DESCRIPTION

[0020] The technical solutions in the embodiments of the present application will be clearly described below with reference to the drawings in the embodiments of the present application. Obviously, the described embodiments are part of the embodiments of the present application, rather than all the embodiments. Based on the embodiments of the present application, other embodiments obtained by those skilled in the art without creative work fall within the protection scope of the present application.

[0021] Hereinafter, the terms "first", "second", and the like are only used for description purposes, and cannot be understood as indicating or implying relative importance or implicitly indicating the number of the indicated technical features. Therefore, the features defined with "first", "second", and the like can explicitly or implicitly include one or more features. In the description of the present application, unless otherwise specified, the meaning of "multiple" is two or more.

[0022] In addition, in the present application, the orientation terms such as "upper", "lower", "inner", "outer", and the like are defined with respect to the orientation of the components shown in the drawings. It should be understood that these directional terms are relative concepts, which are used for relative description and clarification, and can be changed accordingly according to the change of the orientation of the components placed in the drawings.

[0023] The reverse elastic load disturbance refers to the disturbance elastic load force in the same direction as the running speed of the hydraulic cylinder, which is usually caused by the inertia force or speed same direction external force when the actuator is quickly stopped. Since the traditional variable speed pump control system uses an on-off valve to realize the reversing function, the oil return cavity only relies on the sliding friction force to provide damping, which leads to poor performance of the system in suppressing the reverse elastic load disturbance. The uncertain reverse elastic load will cause the pump control tracking performance to decline.

[0024] In the existing hydraulic configuration scheme, although various hydraulic principles have been proposed to deal with the reverse elastic load disturbance of the hydraulic actuator. However, few can achieve energy-saving characteristics comparable to variable speed pump control systems and high-precision actuator trajectory tracking performance. Although the variable speed pump control system can adopt adaptive model predictive control, the servo motor-quantitative pump can accurately provide the ideal flow, but due to the design of the open hydraulic circuit, the actuator only relies on sliding friction as damping, so that the system cannot track the ideal trajectory under the condition of reverse elastic load disturbance. That is, in the case of fast braking or reverse elastic load of the actuator, the above system cannot guarantee high-precision trajectory tracking performance, so it is necessary to improve the traditional variable speed pump control cylinder hydraulic circuit in the face of control requirements under the condition of reverse elastic load disturbance, and design an optimized control algorithm to achieve high-precision and high-energy-efficiency trajectory tracking performance.

[0025] In view of the poor reverse elastic load disturbance suppression capability of the variable speed pump control electro-hydraulic actuator, the embodiments of the present application provide a control method of an electro-hydraulic execution system with a return oil cavity pressure regulation, which adopts a pump-valve composite control system method combining variable speed pump control and valve control technology. In the pump control strategy, the hydraulic cylinder pump control cavity realizes pressure control by adjusting the variable speed pump flow. In order to improve the pressure control accuracy and response speed, the pressure adaptive model predictive control (FFRLS-AMPC) adjustment strategy with forgetting factor recursive least squares method is adopted. By updating the cavity volume in real time and using the recursive least squares online estimation of elastic load parameter change with forgetting factor, the parameter robustness of the model predictive control is improved; the motor torque control provides high bandwidth and reduces torque ripple through the generalized double vector model predictive control (GTV-MPCC-L) based on the Luenberger observer.

[0026] In the valve control strategy: in the hardware aspect of the valve control, the three-position four-way on-off valve in the open variable speed pump control system is replaced by an H-bridge structure composed of four two-position two-way valves, including the first on-off valve V1 and the second on-off valve V2, which are two-position two-way on-off valves to realize the reversing function, and the first proportional valve V3 and the second proportional valve V4, which are two-position two-way proportional valves to realize the return oil cavity pressure control, to build a pump-valve composite control system. On the premise of considering the high energy efficiency of the pump control system, the valve control technology is used to improve the reverse elastic load disturbance suppression performance of the variable speed pump control system. In the algorithm aspect of the valve control, the return oil cavity pressure is controlled by adjusting the opening of the first proportional valve V3 and the second proportional valve V4, the valve control algorithm compensates the reference pressure through the extended state observer (LESO), and the control strategy is combined with the model predictive flow feedforward and the proportional-integral-derivative (PID) feedback to improve the reverse elastic load disturbance suppression performance, guarantee the pump control cavity pressure control performance, and improve the system reverse elastic load disturbance suppression performance.

[0027] Figure 1Figure 1 is a structural schematic diagram of an electro-hydraulic execution system with oil return cavity pressure regulation provided by an embodiment of the present application.

[0028] Referring to Figure 1 As shown in Figure 1, the electro-hydraulic execution system with oil return cavity pressure regulation provided by the embodiment of the present application comprises an elastic load 1, a hydraulic cylinder 2, a pump control direct drive subsystem, a valve body, and an oil tank 6. The pump control direct drive subsystem comprises a bidirectional constant-displacement pump 5 and a permanent magnet synchronous motor 7. The valve body comprises a proportional valve 3 and an on-off valve 4. The proportional valve 3 comprises a first proportional valve V3 and a second proportional valve V4. The on-off valve 4 comprises a first on-off valve V1 and a second on-off valve V2. The constant-displacement pump 5 is connected to the oil tank 6, the permanent magnet synchronous motor 7, the first on-off valve V1, and the second on-off valve V2, respectively. The piston of the hydraulic cylinder 2 is connected to the elastic load 1. The first proportional valve V3 and the second proportional valve V4 are connected to the oil return cavity 21 and the pump control cavity 22 of the constant-displacement pump 5, respectively. The first on-off valve V1 and the second on-off valve V2 are connected to the first proportional valve V3 and the second proportional valve V4, respectively. Specifically, one end of the constant-displacement pump 5 is connected to the first on-off valve V1 and the second on-off valve V2, respectively. The other end of the constant-displacement pump 5 is connected to the oil tank 6. The permanent magnet synchronous motor 7 is connected to the bidirectional constant-displacement pump 5. The permanent magnet synchronous motor 7 and the bidirectional constant-displacement pump 5 serve as a main variable pump source. The first on-off valve V1 and the second on-off valve V2 realize a reversing function. The first proportional valve V3 and the second proportional valve V4 are used for oil return pressure control. The first on-off valve V1 is further connected to the second on-off valve V2. The first proportional valve V3 is further connected to the second proportional valve V4. The first on-off valve V1, the second on-off valve V2, the first proportional valve V3, and the second proportional valve V4 form an H-bridge structure.

[0029] Figure 2 Figure 2 is a general control block diagram of an electro-hydraulic execution system with oil return cavity pressure regulation provided by an embodiment of the present application.

[0030] Referring to Figure 2 As shown in Figure 2, a pump control pressure algorithm design is provided. Pump control pressure adopts a Forgetting Factor Recursive Least Square-Adaptive Model Predictive Control (FFRLS-AMPC) strategy. Pump control cavity pressure control obtains an optimal motor torque control amount through adaptive model predictive control, improves pressure dynamic response, adopts online identification of elastic load elastic coefficient, dynamically adjusts prediction model parameters, and improves robustness of adaptive model predictive control. Torque control of the permanent magnet synchronous motor 7 adopts a Generalized Two-Vector Model Predictive Current Control (GTV-MPCC) with a Luenberger observer to realize high-frequency response, reduce torque fluctuation, and accurately and quickly control the flow rate of the permanent magnet synchronous motor 7 driving the constant-displacement pump 5 to the pump control cavity 22 of the hydraulic cylinder 2, thereby realizing pump control cavity pressure control.

[0031] Figure 3 is a block diagram of an electro-hydraulic actuating system with back oil chamber pressure regulation provided by an embodiment of the present application. In the diagram, Figure 3 corresponds Figure 2 to the back oil chamber pressure proportional valve control part.

[0032] In combination Figures 1 to 3 , as shown, the system dynamic performance can be improved by adjusting the back pressure, and the principle of the proposed back oil chamber pressure planner and controller is analyzed. By adjusting the opening of the first proportional valve V3 and the second proportional valve V4, the best back oil damping can be obtained, so that the pressure of the back oil chamber of the pump valve composite system can produce appropriate throttling damping, and the ability to suppress reverse elastic load disturbance of the pump valve composite system can be improved. When the system is stable, the system is kept at a small preset pressure value , maintaining a certain ability to suppress reverse elastic load disturbance, while avoiding cavitation and air suction in the back oil chamber under reverse elastic load. In addition, the back oil pressure can also be controlled not to be too high, so as to avoid excessive energy loss due to back oil throttling.

[0033] Figure 4 is a back oil chamber valve control pressure block diagram provided by an embodiment of the present application.

[0034] Referring to Figure 4 , the valve control strategy can be composed of a pressure planner and a pressure controller. The pressure planner observes the elastic load disturbance through a linear extended state observer (LESO) to dynamically adjust the back oil chamber pressure reference, suppresses the elastic load disturbance, provides appropriate throttling damping, and improves the anti-disturbance performance of the pump control chamber. When there is no elastic load disturbance, the pressure planner outputs a small preset pressure , ensuring that the back oil chamber has a certain back oil damping, and when the disturbance occurs, the compensation reference pressure is superimposed to increase the back oil damping and improve the performance of suppressing reverse elastic load disturbance. The pressure controller combines PID feedback closed-loop control based on valve model prediction flow feedforward control. The advantage of this control method is that it has feedforward control to compensate for major disturbances, and closed-loop feedback to eliminate errors caused by inaccurate valve models, thereby improving the control accuracy of the back oil chamber pressure.

[0035] The specific control block diagram of the valve control pressure subsystem is shown in Figure 4 , in which is the disturbance elastic load compensation reference pressure, the elastic load disturbance is estimated by a linear extended observer , the compensation term is obtained by balancing the hydraulic cylinder force , and the back oil chamber pressure reference signal is obtained . When there is no elastic load disturbance, the reference is a small preset back oil chamber pressure , and when there is an elastic load disturbance, the reference increases Because the valve model used by the flow feedforward does not consider the influence of model mismatch, there is an error. By introducing a PID controller for feedback closed-loop control, the actual return oil chamber pressure signal measured by a pressure sensor and the pump control chamber outlet pressure reference signal are used to adjust the proportional valve spool control voltage, so as to realize closed-loop control of the return oil chamber outlet pressure.

[0036] Figure 5 is a flowchart of a control method of an electro-hydraulic execution system with return oil chamber pressure regulation provided by an embodiment of the present application.

[0037] Referring to FIG. 1, Figure 5 the control method can be implemented by the following steps S100 to S600.

[0038] Step S100: determining a reference compensation pressure of the return oil chamber by using an extended state observer principle .

[0039] When an elastic load disturbance occurs, in order to make the pump control chamber pressure keep track of stability and improve the anti-disturbance performance, the valve opening connected to the return oil chamber needs to be controlled to adjust the return oil chamber pressure, provide appropriate throttling damping, and eliminate the disturbance. Because the elastic load is elastic and has large stiffness, when the displacement of the hydraulic cylinder changes, the elastic load and the return oil chamber pressure jointly balance the disturbance. When there is no compensation reference pressure, the return oil chamber reference pressure is equal to a smaller preset pressure. The disturbance can be eliminated and the given reference pressure can be reached by flow feedforward based on the valve model and PID feedback control. When the hydraulic cylinder is subjected to a constant reverse elastic load disturbance, according to the hydraulic cylinder force balance formula, because the return oil chamber pressure is reduced to a smaller preset pressure by PID closed-loop control, the elastic load will bear more force to balance the system stability. When the elastic load disturbance is removed, according to the hydraulic cylinder force balance formula, the elastic load restores the deformation to reduce the output force, and at the same time, the return oil chamber pressure will be further reduced, resulting in that the return oil chamber pressure cannot be maintained at a smaller preset pressure. In order to ensure that the system has a certain throttling damping when the disturbance occurs, and the return oil chamber pressure is a smaller preset pressure when the system is stable, an extended state observer is introduced to observe the elastic load disturbance force, and a compensation reference pressure value is calculated by the hydraulic cylinder force balance.

[0040] Specifically, step S100 can include steps S110 to S130.

[0041] Step S110: establishing a state space equation of the electro-hydraulic execution system.

[0042] In order to facilitate the description of the linear extended state observer principle, a second-order system is first established: ; wherein, is the piston rod displacement, isderivative of , is the piston rod velocity, derivative of , is the effective area of the piston in the hydraulic cylinder's return chamber; is the effective area of the piston in the hydraulic cylinder's pump chamber; m is the mass of the elastic load; is the viscous damping coefficient; is the elastic coefficient of the elastic load; p1 is the pressure in the pump chamber after the elastic load disturbance; p2 is the pressure in the return chamber; f L is the elastic load disturbance.

[0043] wherein, , a 1n is the first linear extended observer coefficient, a 2n is the second linear extended observer coefficient, a 3n is the third linear extended observer coefficient, a 4n is the fourth linear extended observer coefficient, a 5n is the fifth linear extended observer coefficient.

[0044] The above formula is arranged as follows: ; Let , be written in the form of state space equation as follows: ; wherein, is the state variable of the state space equation, derivative of ; is the output variable of the state space equation; is the input vector; ; A is the system matrix of the state space equation, B is the control matrix of the state space equation; C is the output matrix of the state space equation; E is the disturbance matrix of the state space equation, f is the observed elastic load disturbance, derivative of f; , ; ; A2 is the effective area of the piston in the hydraulic cylinder's pump chamber, m is the mass of the elastic load, b v is the viscous damping coefficient, k s is the elastic coefficient of the elastic load.

[0045] Step S120: determining the estimated elastic load disturbance of the electro-hydraulic execution system by using the state space equation combined with the extended state observer principle .

[0046] Specifically, step S120 can include step S121 and step S122.

[0047] Step S121: Control the state space model based on the extended state observer principle to obtain an extended state equation.

[0048] Figure 6 is a linear extended observer principle diagram.

[0049] Referring to Figure 6 , the LESO is designed according to the state space equation as follows: ; Wherein, , and are the observation value and gain vector of the observer respectively. is the first observation value, is the second observation value.

[0050] Let The current extended state equation is obtained as follows: ; z1 is the observed piston rod displacement; is the derivative of z1; z2 is the observed piston rod speed; is the derivative of z2; z3 is the observed disturbance; is the derivative of z3; is the first gain coefficient; is the second gain coefficient; is the third gain coefficient; a 1n is the first linear extended observer coefficient.

[0051] Step S122: Real-time estimation of the elastic load disturbance f L based on the extended state equation to obtain the estimated elastic load disturbance .

[0052] Specifically, step S122 can include step S1221 and step S1222.

[0053] Step S1221: Place the poles of the extended state equation at the same position to obtain the characteristic polynomial .

[0054] In order to ensure stability and facilitate parameter setting, according to Lyapunov stability criterion, the poles of the characteristic equation of the observer are placed at the same position , so the characteristic polynomial is: ;

[0055] where I is an identity matrix, is the bandwidth of LESO, s is a complex variable in the complex frequency domain.

[0056] Step S1222: determining the estimated elastic load disturbance based on the characteristic polynomial , the first gain coefficient , the second gain coefficient and the third gain coefficient .

[0057] In order to stabilize the observer, the roots of the characteristic equation need to be distributed on the negative half axis, so the LESO gain coefficients can be designed as: ; According to the known disturbance term, the estimated elastic load disturbance can be obtained as: .

[0058] Step S130: determining the reference compensation pressure of the return oil chamber based on the estimated elastic load disturbance . .

[0059] Assuming that the pump control chamber pressure is unchanged, the elastic load force is unchanged after the elastic load disturbance occurs, the hydraulic cylinder force balance formula before and after the elastic load disturbance is: .

[0060] Combined with the hydraulic cylinder force balance formula, the return oil chamber compensation pressure is obtained as: .

[0061] Step S200: determining the return oil chamber reference pressure based on the reference compensation pressure and the preset pressure of the return oil chamber.

[0062] In step S200, K is the elastic load disturbance coefficient, and let , when the disturbance is eliminated in the steady state, the return oil chamber pressure is the smaller preset pressure , and the return oil chamber reference pressure is designed as: .

[0063] After the reference pressure of the return oil chamber is determined, the proportional valve is controlled by the flow feedforward controller and the PID feedback controller.​

[0064] Step S300: Based on the pressure of the pump control chamber Pressure of the return oil chamber outlet pressure of the metering pump and the pressure of the fuel tank Determine the pressure difference across the valve body. .

[0065] The valve body includes a proportional valve and an on / off valve, and the pressure difference across the valve body... The electro-hydraulic actuator has two operating modes: a first operating mode and a second operating mode. In the first operating mode, the first switching valve is open and the second switching valve is closed. In the second operating mode, the first switching valve is closed and the second switching valve is open.

[0066] In this step, since the flow feedforward control based on the valve model requires a valve mathematical model, the mathematical model analysis of the two-position two-way proportional valve used is performed first.

[0067] Oil is compressible; the compressibility of the cavity is achieved through the equivalent elastic modulus β. e Calculations. The flow state of oil is classified into two types: laminar flow and turbulent flow. The boundary between the two is based on the critical Reynolds number. The specific state of the flow field is determined by calculating the Reynolds number. The determination is made using the following formula: ; Where q is the flow rate (m 3 / s); S v For the flow area (m²) 2 ); The kinematic viscosity is (m² / s).

[0068] Hydraulic diameter The calculation formula is: ; S is the maximum cross-sectional area (m²) of the throttling orifice of the proportional valve. 2 ).

[0069] Number of streams The calculation formula is: .

[0070] in, This represents the pressure difference across the valve port.

[0071] Figure 7 It is a curve showing the relationship between flow coefficient and Reynolds number.

[0072] See Figure 7 As shown, the flow coefficient of the throttling orifice is not constant, but varies with the Reynolds number Re. When the Reynolds number is small, the flow coefficient drops rapidly with the decrease of the Reynolds number, and in this stage, its variation trend can be approximated as an asymptote. When the Reynolds number is large, the flow coefficient tends to a constant value, which is the maximum flow coefficient . Therefore, the flow coefficient in the entire Reynolds number range is calculated as: ; wherein, is the maximum flow coefficient, which is generally set to 0.7, and the flow coefficient is constant, the flow of the proportional valve in the electro-hydraulic execution system depends on the valve port amplitude x v and the pressure difference before and after the valve port , the valve port amplitude x v depends on the control electric signal, and the pressure difference before and after the valve port depends on the elastic load pressure and the inlet pressure.

[0073] The valve port amplitude of the proportional valve is x v , and according to the flow formula of the throttle orifice, the first proportional valve V3 and the second proportional valve V4 controlled by the oil return cavity pressure are expressed as:

[0074] ; wherein, is the valve core position control voltage (V); S is the maximum cross-sectional area of the throttle hole of the valve port of the proportional valve (m 2 ); is the normalized coefficient of the valve control voltage to the valve core position (1 / V); is the flow coefficient; is the oil density (kg / m 3 ); represents the pressure difference before and after the proportional valve (Pa).

[0075] The pressure difference before and after the valve varies according to different working models. In the first working model, the variable speed pump is connected to the oil return cavity, and the pressure difference before and after the first on-off valve and the pressure difference before and after the second proportional valve satisfy the following formulas respectively: ; In the second working model, the variable speed pump is connected to the pump control cavity, and the pressure difference before and after the second on-off valve and the pressure difference before and after the first proportional valve satisfy the following formulas respectively:

[0076] wherein, Pc is the pressure (MPa) of the oil return cavity, Pd is the outlet pressure (MPa) of the metering pump; Pp is the pressure (MPa) of the pump control cavity; Pb is the pressure (MPa) of the oil tank.

[0077] Thus, in the present application, first, a compensation pressure generated by an elastic load disturbance is observed by a linear extended state observer (LESO) , and a smaller preset pressure is combined to obtain a compensated reference pressure of the oil return cavity, and second, an oil return cavity pressure controller is designed to adjust the valve control voltage of the first proportional valve V3 and the second proportional valve V4, so that they track the reference pressure and balance the disturbance elastic load.

[0078] Step S400: The preliminary control voltage of the proportional valve is determined based on the pressure difference before and after the valve body and the proportional valve flow using the valve body flow formula.

[0079] The preliminary control voltage of the proportional valve satisfies: ; wherein i is 3, 4.

[0080] Specifically, according to the mathematical model of the two-position two-way proportional valve, in the first working mode, the variable displacement pump is connected to the oil return cavity, and the second proportional valve V4 controls the flow from the oil return cavity to the oil tank. According to the proportional valve flow equation, the main variables affecting the proportional valve are the pressure difference before and after the proportional valve , the flow through the proportional valve , and the valve control voltage , wherein can be obtained according to the piston rod speed . Since the flow coefficient is not measurable, the valve control voltage is calculated according to the maximum flow coefficient.

[0081] The pressure difference before and after the oil return cavity valve V4 can be expressed as: ; By substituting the valve flow and the pressure difference before and after the valve into the valve flow equation, the control voltage of the valve is obtained as: .

[0082] ​​Step S500: preliminary control voltage of proportional valve combined with PID feedback determining final control voltage of proportional valve reference pressure of oil return cavity reference input of PID feedback.

[0083] Since the valve model used by flow feedforward control does not consider model mismatch and parameter variation, there is a control error. By adding a PID feedback controller to adjust the opening of the valve, the error is eliminated.

[0084] PID control has the advantage of simple calculation process, without the need to accurately know the mathematical model of the controlled object, it can realize effective control. Its control principle is to combine the error between the actual behavior of the controlled object and the target through proportional, integral and differential operations, and then act on the controlled object.

[0085] The expression of the PID controller is: .

[0086] wherein, = , is the control deviation signal, is the proportional coefficient, is the integral time constant, is the differential time constant; t0 and t i are different times.

[0087] Through flow feedforward and PID feedback control, the valve body voltage control signal output by the final oil return cavity pressure controller is: .

[0088] Specifically, in the first working mode, the final oil return cavity pressure controller outputs the second proportional valve voltage as: .

[0089] In the second working mode, the final oil return cavity pressure controller outputs the first proportional valve voltage as: .

[0090] Step S600: control proportional valve based on final control voltage of proportional valve .

[0091] After obtaining the final control voltage of the proportional valve , the working of the proportional valve can be controlled through the final control voltage of the proportional valve .

[0092] The variable speed pump control direct drive system has the problem of poor ability to suppress reverse elastic load disturbance, and the control performance of the pump control chamber is reduced. The control method of the electro-hydraulic execution system with oil return chamber pressure regulation provided in the embodiments of the present application separates the pump control system and the valve control system by combining the pump control system with the variable speed pump control system and the valve control technology to form a pump-valve composite system. In order to ensure the control accuracy of the system, the pump control strategy adopts the forgetting factor recursive least square adaptive model predictive pressure control (FFRLS-AMPC) strategy, which fully utilizes the energy saving advantage of variable speed pump control speed regulation and the dynamic response improvement of adaptive model predictive control; the oil return chamber valve control pressure strategy is composed of a pressure planner and a pressure controller, wherein the pressure planner outputs the oil return chamber reference pressure through a small preset pressure and a linear extended state observer (LESO), and the pressure controller is combined with the flow feedforward based on the valve model and the proportional-integral-derivative (PID) feedback, which not only has feedforward control to compensate for the main disturbance, but also has closed-loop feedback to suppress other disturbances, thereby improving the performance of the system in suppressing reverse elastic load disturbance.

[0093] Corresponding to the embodiments of the control method of the electro-hydraulic execution system with oil return chamber pressure regulation, the present application also provides embodiments of the electro-hydraulic execution system with oil return chamber pressure regulation.

[0094] Figure 8 is a structural block diagram of the electro-hydraulic execution system with oil return chamber pressure regulation provided in the embodiments of the present application.

[0095] Referring to Figure 8 , the electro-hydraulic execution system provided in the embodiments of the present application can include a first determination module 101, a second determination module 102, a third determination module 103, a fourth determination module 104, a fifth determination module 105, and a control module 106.

[0096] The first determination module 101 is configured to determine the reference compensation pressure of the oil return chamber based on the pump control chamber pressure .

[0097] That is, the first determination module 101 is configured to execute the step S100 provided in the above embodiments.

[0098] The second determination module 102 is configured to determine the oil return chamber reference pressure based on the reference compensation pressure and the preset pressure .

[0099] That is, the second determination module 102 is configured to execute the step S200 provided in the above embodiments.

[0100] The third determination module 103 is configured to determine the pump control chamber pressure based on the pump control chamber pressure outlet pressure of the metering pump and tank pressure Determine the pressure difference across the valve body. Among them, the pressure difference across the valve body The electro-hydraulic actuator has two operating modes: a first operating mode and a second operating mode. In the first operating mode, the first switching valve is open and the second switching valve is closed. In the second operating mode, the first switching valve is closed and the second switching valve is open.

[0101] In other words, the third determining module 103 is used to execute step S300 provided in the above embodiments.

[0102] The fourth determining module 104 is configured to use the valve body flow formula, based on the pressure difference across the valve body. and proportional valve flow Determine the initial control voltage of the proportional valve .

[0103] In other words, the fourth determining module 104 is used to execute step S400 provided in the above embodiments.

[0104] The fifth determining module 105 is configured to initially control the voltage by combining a PID feedback unit and a proportional valve. Determine the final control voltage of the proportional valve Reference pressure of the oil return chamber This is the reference input for the PID feedback circuit.

[0105] In other words, the fifth determining module 105 is used to execute step S500 provided in the above embodiments.

[0106] Control module 106 is configured to control the final voltage of the proportional valve. Control proportional valve.

[0107] In other words, the sixth determining module is used to execute step S600 provided in the above embodiments.

[0108] It should be noted that, upon considering the specification and practicing the application disclosed herein, those skilled in the art will readily conceive of other embodiments of this application. This application is intended to cover any variations, uses, or adaptations of this application that follow the general principles of this application and include common knowledge or customary techniques in the art not disclosed herein.

[0109] It should be understood that this application is not limited to the precise structure described above and shown in the accompanying drawings, and various modifications and changes can be made without departing from its scope. The true scope is indicated by this application.

Claims

1. A control method for an electro-hydraulic actuator system with return oil chamber pressure regulation, characterized in that, An electro-hydraulic actuator system with return oil chamber pressure regulation is applied. The system includes an oil tank, a pump-controlled direct drive subsystem, a valve body, an elastic load, and a hydraulic cylinder. The pump-controlled direct drive subsystem includes a permanent magnet synchronous motor and a fixed displacement pump. The valve body includes a proportional valve and a switching valve. The proportional valve includes a first proportional valve and a second proportional valve. The switching valve includes a first switching valve and a second switching valve. The fixed displacement pump is connected to the oil tank, the permanent magnet synchronous motor, the first switching valve, and the second switching valve. The piston of the hydraulic cylinder is connected to the elastic load. The first and second proportional valves are connected to the return oil chamber and the pump-controlled chamber of the hydraulic cylinder, respectively. The first and second switching valves are connected to the first and second proportional valves, respectively. The first proportional valve, the second proportional valve, the first switching valve, and the second switching valve form an H-bridge structure. The control method includes: The reference compensation pressure of the return oil chamber is determined using the principle of expansion state observer. ; Based on the reference compensation pressure and the preset pressure of the oil return chamber Determine the reference pressure of the oil return chamber. ; Based on the pressure of the pump control chamber The pressure of the oil return chamber The outlet pressure of the metering pump and the pressure of the oil tank Determine the pressure difference across the valve body. The pressure difference across the valve body. Related to the operating mode of the electro-hydraulic actuator, the operating mode of the electro-hydraulic actuator includes a first operating mode and a second operating mode. In the first operating mode, the first switching valve is open and the second switching valve is closed; in the second operating mode, the first switching valve is closed and the second switching valve is open. The valve body flow rate formula is used, based on the pressure difference across the valve body. and the flow rate of the proportional valve Determine the initial control voltage of the proportional valve. ; The initial control voltage of the PID feedback unit and the proportional valve Determine the final control voltage of the proportional valve. The reference pressure of the oil return chamber This serves as the reference input for the PID feedback unit; Based on the final control voltage of the proportional valve Control the proportional valve.

2. The control method according to claim 1, characterized in that, The reference compensation pressure is determined using the principle of the expansion state observer. ,include: Establish the state-space equations of the electro-hydraulic actuation system; The estimated elastic load disturbance of the electro-hydraulic actuator is determined by combining the state-space equation with the principle of the extended state observer. ; After the elastic load is disturbed, the pressure in the pump control chamber Unchanged, the elastic load force Under the condition that remains unchanged, combined with the estimated elastic load disturbance Determine reference compensation pressure .

3. The control method according to claim 1, characterized in that, Establishing the state-space equations of the electro-hydraulic actuation system includes: The state-space equation is: ;in, Let be the state variables of the state-space equation. for The derivative; The output variables of the state-space equation; The input vector; ; The system matrix of the state-space equation is... The control matrix of the state-space equation; The output matrix of the state-space equation; Let be the perturbation matrix of the state-space equation. To observe elastic load disturbances, for The derivative; , ; ; The effective area of ​​the piston in the pump control chamber of the hydraulic cylinder is [missing information]. The mass of the elastic load, The viscous damping coefficient is... The elasticity coefficient is the elastic load. It is the displacement of the piston rod. It is the piston rod speed; For the coefficients of the second linearly extended observer, For the coefficients of the third linearly extended observer, For the fourth linearly extended observer coefficients.

4. The control method according to claim 3, characterized in that, The estimated elastic load disturbance of the electro-hydraulic actuator is determined by combining the state-space equation with the principle of the extended state observer. ,include: Based on the principle of extended state observer, the state space model is controlled to obtain the extended state equation; The extended state equations include: ; It involves observing the displacement of the piston rod; for The derivative; It involves observing the piston rod speed; for The derivative; It is an observed disturbance; for The derivative; The first gain coefficient; This is the second gain coefficient; This is the third gain coefficient; These are the coefficients of the first linearly extended observer; , The effective area of ​​the piston in the return oil chamber of the hydraulic cylinder; Based on the extended state equation, the elastic load disturbance is... Real-time estimation is performed to obtain the estimated elastic load disturbance. .

5. The control method according to claim 4, characterized in that, Based on the extended state equation, the elastic load disturbance... Real-time estimation is performed to obtain the estimated elastic load disturbance. ,include: Place the poles of the extended state equation in the same position. The characteristic polynomial is obtained. ; wherein, the characteristic polynomial for: I is the identity matrix. Let be the bandwidth of LESO, and s be a complex variable in the complex frequency domain; Based on the characteristic polynomial The first gain coefficient The second gain coefficient and the third gain coefficient Determine the estimated elastic load disturbance Wherein, the first gain coefficient The second gain coefficient and the third gain coefficient satisfy The estimated elastic load disturbance satisfy: ; The coefficients of the fifth linearly extended observer. .

6. The control method according to claim 5, characterized in that, Based on the reference compensation pressure and the preset pressure of the oil return chamber Determine the reference pressure of the oil return chamber. ,include: ;in, This is the disturbance coefficient for elastic load. .

7. The control method according to claim 6, characterized in that, Based on the pump control chamber pressure The pressure of the oil return chamber The outlet pressure of the metering pump and the oil tank pressure Determine the pressure difference across the valve body. ,include: In the first operating mode, the pressure difference across the valve body Including the pressure difference across the first switching valve Pressure difference across the second proportional valve The pressure difference across the first switching valve and the pressure difference across the second proportional valve satisfy: ； In the second operating mode, the pressure difference across the valve body Including the pressure difference across the second switching valve Pressure difference across the first proportional valve The pressure difference across the second switching valve and the pressure difference across the second proportional valve satisfy: 。 8. The control method according to claim 7, characterized in that, The valve body flow rate formula is used, based on the pressure difference across the valve body. and proportional valve flow Determine the initial control voltage of the proportional valve. ,include: The initial control voltage of the proportional valve satisfy: Where i is 3 or 4; S is the maximum cross-sectional area of ​​the throttling orifice of the proportional valve. It is the normalization coefficient for the valve control voltage to the valve core position; The maximum flow coefficient; It refers to the density of the oil.

9. The control method according to claim 8, characterized in that, The initial control voltage is achieved by combining the PID feedback unit and the proportional valve. Determine the final control voltage of the proportional valve ,include: The proportional valve ultimately controls the voltage. satisfy: ； in, = , To control the deviation signal, This is the proportionality coefficient. The integral time constant is... The differential time constants; t0 and t i These represent different times.

10. An electro-hydraulic actuator with return oil chamber pressure regulation, characterized in that, The control method of the electro-hydraulic actuator system with return oil chamber pressure regulation as described in any one of claims 1-9, wherein the electro-hydraulic actuator system with return oil chamber pressure regulation comprises: The first determining module is configured to determine the reference compensation pressure of the return oil chamber using the principle of the expansion state observer. ; The second determining module is configured to base its determination on the reference compensation pressure. and the preset pressure of the oil return chamber Determine the reference pressure of the oil return chamber. ; The third determining module is configured to determine the pressure based on the pump control chamber. The pressure of the oil return chamber The outlet pressure of the metering pump and the pressure of the oil tank Determine the pressure difference across the valve body. The pressure difference across the valve body. Related to the operating mode of the electro-hydraulic actuator, the operating mode of the electro-hydraulic actuator includes a first operating mode and a second operating mode. In the first operating mode, the first switching valve is open and the second switching valve is closed; in the second operating mode, the first switching valve is closed and the second switching valve is open. The fourth determining module is configured to use the valve body flow formula, based on the pressure difference across the valve body. and the flow rate of the proportional valve Determine the initial control voltage of the proportional valve. ; The fifth determining module is configured to combine the PID feedback and the initial control voltage of the proportional valve. Determine the final control voltage of the proportional valve. The reference pressure of the oil return chamber This serves as the reference input for the PID feedback unit; The control module is configured to operate based on the final control voltage of the proportional valve. Control the proportional valve.

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