Bidirectional hydraulic three-state switch boosting system and control method thereof

Abstract

The invention relates to a bidirectional hydraulic three-state switch boosting system and a control method thereof. The system comprises a basic boosting loop, a follow current branch and a load hydraulic cylinder H, the basic boosting loop comprises a liquid sensing element L1, a high-speed switch valve S1, a load end high-speed one-way valve C3, a load end energy accumulator A and an auxiliary liquid sensing element L2, an outlet of a pump source Ps is connected with the L1, an outlet of the L1 is simultaneously connected with the S1 and the C3, an outlet of the S1 is connected with an oil tank T, an outlet of the C3 is simultaneously connected with the A and the L2, an outlet of the L2 is connected with the H, and an outlet of the H is connected with the T; the follow current branch comprises a high-speed switch valve S2 and a high-speed one-way valve C2, an outlet of the L1 is connected with S2, an outlet of S2 is connected with C2, and an outlet of C2 is connected with an inlet of L1; and two ends of the high-speed one-way valve C1 are connected in parallel with two ends of the high-speed one-way valve C1, so that reverse motion control of the load hydraulic cylinder is realized. According to the system and the control method thereof, high-precision and high-dynamic control of the hydraulic load can be realized, and the working condition range of the transformer for controlling the hydraulic load is widened, so that the hydraulic switch boosting transformer can be better applied to industrial practice.

Description

Technical Field

[0001] The present invention relates to the technical field of hydraulic switching voltage transformation, and particularly relates to a bidirectional hydraulic three-state switch boosting system and a control method thereof. Background Art

[0002] The wide application of hydraulic proportional servo technology enables the hydraulic system to precisely control the movement of the hydraulic load actuator through the control of the valve opening. However, this control method is based on valve orifice throttling control. While controlling the load, a large amount of energy is consumed in the valve itself, and the voltage transformation efficiency is low. Different from throttling control, there is a concept of switching control in the field of power electronics, and its theoretical voltage transformation efficiency can reach 100%. The hydraulic switching voltage transformation technology is the implementation of the switching control technology in the field of power electronics in the hydraulic field, and can also reach 100% voltage transformation efficiency in theory, which is a potential new type of high-efficiency hydraulic voltage transformation method. The hydraulic switching voltage transformation technology adjusts the pressure and flow rate output from the pump source to the load by controlling the switching duty ratio of the high-speed switching valve of the control system, which conforms to the development trend of the digital hydraulic era. Therefore, the development of this technology can further promote the development of high-efficiency hydraulic voltage transformation and digital hydraulics.

[0003] The existing engineering application research on hydraulic switching transformer technology mainly focuses on the step-down system. Some research scholars have conducted research on control strategies for its engineering application. For example, Reference 1 (Kogler H, Scheidl R. Hydraulic switching control of resonant drives [C]. 2010.) proposed a control method combining flow feedforward and PD feedback of transformer flow error, which achieved the motion control of the load hydraulic cylinder. However, due to the fact that PD control does not consider the nonlinear problems of the transformer, the control accuracy is not good. Subsequently, the research team proposed a load flatness controller in Reference 2 (Kogler H, Scheidl R. Linear motion control with a low-power hydraulic switching converter - Part II: Flatness-based control [J]. Proceedings of the Institution of Mechanical Engineers, Part I: Journal of Systems and Control Engineering. 2015). Simulations and experiments showed that the hydraulic switching step-down system using this controller had good control effects when driving the hydraulic cylinder load. However, the above research has not yet involved the drive control of the hydraulic switching step-up system. In some hydraulic systems with many hydraulic actuators, only one or a few hydraulic actuators require a relatively high load pressure. At this time, a hydraulic switching step-up transformer can be used to maintain a relatively low pump source pressure and reduce the leakage flow of the pump source. In some other cases, due to the limitation of the pump source volume and size, a high pressure higher than the rated pressure cannot be provided. At this time, a hydraulic switching step-up transformer can also be used for voltage transformation. It has been found that due to the design of the reference circuit boost transformer circuit of the hydraulic switching step-up transformer, it inherits some characteristics of the circuit boost transformer. When the liquid inductance flow is continuous, the system linearized transfer function has a right-half plane zero point, resulting in the overall system having non-minimum phase characteristics, causing the system control bandwidth to be limited and it being impossible to achieve high-precision and high-dynamic motion control of the hydraulic load through high-gain feedback. Although the intermittent liquid inductance flow can alleviate this phenomenon, due to the limited ability to supply the load in the intermittent liquid inductance flow mode, it cannot be applied to some high-power scenarios. Therefore, it cannot be the best solution for the right-half plane zero point of the system. To avoid the performance limitations brought by the right-half plane zero point of the system to the system, as mentioned in Reference 3 (Ye Linqi, Zong Qun, Tian Bailing, etc. Review of tracking control for non-minimum phase systems [J]. Control Theory & Applications, 2017.), the only way is to change the system structure or avoid designing the system into a system containing a right-half plane zero point.However, the existing technology lacks research on the structure of the hydraulic switch boost system and also lacks research on the control method for the hydraulic switch boost system to control the actual load of the hydraulic cylinder. There are still difficulties in the high-precision and high-dynamic control of the hydraulic switch boost system. Summary of the Invention

[0004] The purpose of the present invention is to provide a bidirectional hydraulic three-state switch boost system and its control method. This system and its control method can achieve high-precision and high-dynamic control of the hydraulic load and broaden the working condition range of the transformer to control the hydraulic load, enabling the hydraulic switch boost transformer to be better applied to industrial practice.

[0005] To achieve the above purpose, the technical solution adopted by the present invention is: a bidirectional hydraulic three-state switch boost system, including a basic boost circuit, a freewheeling branch, and a load hydraulic cylinder H; the basic boost circuit includes a liquid inductance element L1, a high-speed switching valve S1, a load-side high-speed check valve C3, a load-side accumulator A, and an auxiliary liquid inductance L2. The outlet of the pump source Ps is connected to the liquid inductance element L1. The outlet of the liquid inductance element L1 is simultaneously connected to the high-speed switching valve S1 and the load-side high-speed check valve C3. The outlet of the high-speed switching valve S1 is connected to the oil tank T. The outlet of the load-side high-speed check valve C3 is simultaneously connected to the load-side accumulator A and the auxiliary liquid inductance L2 for stabilizing the flow rate of the load hydraulic cylinder H. The outlet of the auxiliary liquid inductance L2 is connected to the load hydraulic cylinder H, and the outlet of the load hydraulic cylinder H is connected to the oil tank T; the freewheeling branch includes a high-speed switching valve S2 and a high-speed check valve C2. The outlet of the liquid inductance element L1 is connected to the high-speed switching valve S2. The outlet of the high-speed switching valve S2 is connected to the high-speed check valve C2. The outlet of the high-speed check valve C2 is connected to the inlet of the liquid inductance element L1; a high-speed switching valve S3 is connected in parallel at both ends of the load-side high-speed check valve C3, and a high-speed check valve C1 is connected in parallel at both ends of the high-speed switching valve S1 for realizing the reverse movement control of the load hydraulic cylinder.

[0006] The present invention also provides a control method for the above bidirectional hydraulic three-state switch boost system. When the load hydraulic cylinder H is lifted to the set pressure by the boost system under the action of an external force to resist the external load, the movement direction of the load hydraulic cylinder H is opposite to the direction of the external load force, which is defined as the forward movement of the load hydraulic cylinder H; if the movement direction of the load hydraulic cylinder H is the same as the direction of the external load force under the drive of the boost system, it is defined as the reverse movement of the load hydraulic cylinder H.

[0007] Furthermore, during the forward movement of the load hydraulic cylinder H, the working process of a switching cycle of the bidirectional hydraulic three-state switch boost system is divided into three stages:

[0008] (1) In the first working stage, the high-speed switching valve S1 and the high-speed switching valve S2 are simultaneously turned on. The duty cycle of the high-speed switching valve S2 within each cycle is fixed, and its magnitude is determined by the duty cycle of the S2 that enables the system to have the maximum static output and the best transmission efficiency under the duty cycle of each high-speed switching valve S1. The maximum duty cycle of the high-speed switching valve S1 is less than the duty cycle of the high-speed switching valve S2. When the high-speed switching valve S1 and the high-speed switching valve S2 are simultaneously turned on, the system hydraulic oil first passes through the high-speed switching valve S1 branch. The flow rate of the liquid sensing element L1 increases under the action of the pressure difference before and after. At this time, the flow rate required for the movement of the load hydraulic cylinder H is provided by the accumulator A, and the auxiliary liquid inductor L2 plays a role in stabilizing the flow rate.

[0009] (2) In the second working stage, the high-speed switching valve S1 is closed, and the high-speed switching valve S2 remains turned on. At this time, the system hydraulic oil continues to flow through the high-speed switching valve S2. Due to the action of the self-liquid resistance of the high-speed switching valve S2, the high-speed one-way valve C2, and the liquid sensing element L1, the flow rate of the liquid sensing element L1 decreases. At this time, the flow rate required for the movement of the load hydraulic cylinder H is provided by the accumulator A, and the auxiliary liquid inductor L2 plays a role in stabilizing the flow rate.

[0010] (3) In the third working stage, the high-speed switching valves S1 and S2 are closed. At this time, the system hydraulic oil flows to the load branch through the high-speed one-way valve C3 at the load end, and the pressure and flow rate of the accumulator A are replenished to maintain the pressure and flow rate required for the movement of the load hydraulic cylinder H in the first two stages within the next switching cycle.

[0011] Furthermore, during the reverse movement of the load hydraulic cylinder H, the working process of a switching cycle of the bidirectional hydraulic three-state switching boost system is divided into two stages:

[0012] (1) In the first working stage, the high-speed switching valve S3 is turned on. The liquid sensing element L1 accelerates in the reverse direction under the action of the pressure difference between the load high pressure and the pump source pressure, and the reverse flow rate passing through it increases.

[0013] (2) In the second working stage, the high-speed switching valve S3 is closed, and the high-speed one-way valve C1 is turned on. The flow rate of the liquid sensing element L1 in the second working stage is compensated through the oil tank T. The entire system has a back pressure added to the oil tank to compensate for the conduction pressure drop of the high-speed one-way valve C1 and prevent cavitation in the system. During the whole process, due to the outflow of the oil in the accumulator A, the load hydraulic cylinder H moves in the reverse direction.

[0014] Furthermore, the entire system is divided into two parts for control. One part is the outer-loop system composed of accumulator A, auxiliary liquid inductor L2, and load hydraulic cylinder H, and the other part is the inner-loop system composed of the transformer part except the outer-loop system. The inner and outer-loop systems are respectively modeled, and the model-assisted active disturbance rejection control is applied to control the outer-loop system to obtain the desired liquid inductor flow rate of the inner-loop system. The sliding mode controller is applied to control the inner-loop system so that the average flow rate of the liquid inductor element L1 follows the desired liquid inductor flow rate.

[0015] Furthermore, the modeling of the outer-loop system includes the following steps:

[0016] Step S1: According to the force balance equation of the hydraulic cylinder, considering the piston mass of the hydraulic cylinder and the viscous friction coefficient of the hydraulic cylinder, the force balance equation of the load hydraulic cylinder H is established as:

[0017] P 1 A 1 -P t A 2 =ma+B m v+F l

[0018] In the formula, P 1 is the pressure of the rodless cavity of the load hydraulic cylinder H, A 1 is the cross-sectional area of the piston in the rodless cavity of the load hydraulic cylinder H, P t is the pressure of the rod cavity of the load hydraulic cylinder H, A 2 is the cross-sectional area of the piston in the rod cavity of the load hydraulic cylinder H, m is the equivalent mass of the piston of the load hydraulic cylinder H and the external load, B m is the viscous damping coefficient of the piston and the external load, F l is the external load force acting on the piston rod, a is the piston motion acceleration, and v is the piston motion speed; the pressure of the rodless cavity of the load hydraulic cylinder H is:

[0019]

[0020] Step S2: On the basis of not considering the leakage of the motor, according to the pressure-flow relationship at both ends of the hydraulic motor, the series relationship between the auxiliary liquid inductor L2 and the load hydraulic cylinder H, and the relationship between the input flow rate of the hydraulic cylinder and the motion speed of the hydraulic cylinder, the pressure difference at both ends of the auxiliary liquid inductor L2 is obtained as:

[0021]

[0022] In the formula, L 2 is the liquid inductance value of the auxiliary liquid inductor, R L2 is the liquid resistance value of the auxiliary liquid inductor, q 1 is the flow rate at the oil inlet of the hydraulic cylinder;

[0023] Step S3: According to the pressure relationship of the hydraulic series circuit, the pressure on accumulator A is obtained as:

[0024]

[0025] According to the pressure-flow relationship of the accumulator and the parallel relationship between the accumulator and the hydraulic cylinder, it can be obtained that:

[0026]

[0027] In the formula, C c is the equivalent liquid capacitance value of accumulator A at the operating point, and q l is the flow rate flowing into the entire outer loop system;

[0028] Step S4: Combine the formulas in Steps S1 - S3 to obtain the system model of the outer loop system as:

[0029]

[0030] In the formula, is the derivative of the tank pressure. Since the tank pressure is constant, this term can be ignored. x 1 , x 2 , x 3 are the displacement, velocity, and acceleration state variables of the hydraulic cylinder respectively.

[0031] Furthermore, the modeling of the inner loop system includes the following steps:

[0032] Step S1: In the forward motion mode, the mathematical models of each working stage of the transformer are established as follows:

[0033] Stage 1:

[0034]

[0035] Stage 2:

[0036]

[0037] Stage 3:

[0038]

[0039] In the formula, P in (t) is the pump source pressure, q m (t) is the average flow rate of the liquid inductance element L1 in one switching cycle of the high-speed switching valve, P t (t) is the tank pressure, R s1 , R s2 are the non-linear liquid resistances when the high-speed switching valves S1 and S2 are always on, R L1 is the linear liquid resistance when the liquid inductance element L1 is working, Δp ch2(t), Δp ch3 (t) is the pressure drop when the high-speed one-way valves C2 and C3 are working;

[0040] In the forward motion mode, the three working stages of the transformer are unified into a differential equation by the switching period averaging method:

[0041]

[0042] In the formula, d s1 is the duty cycle of the high-speed switching valve S1 in one switching period, d s2 is the freewheeling duty cycle of the liquid inductance element L1, d c3 is the conduction duty cycle of the high-speed one-way valve C3 in one switching period, d s2 = 1 - d s1 - d c3 ;

[0043] Step S2: In the reverse motion mode, the mathematical models of each working stage of the transformer are established as follows:

[0044] Stage 1:

[0045]

[0046] Stage 2:

[0047]

[0048] In the reverse motion mode, the two working stages of the transformer are unified into a differential equation by the switching period averaging method:

[0049]

[0050] In the formula, d s3 is the conduction duty cycle of the high-speed switching valve S3 in one switching period, d c1 is the duty cycle of the high-speed one-way valve C1 in one switching period, d c1 = 1 - d s3 .

[0051] Furthermore, the outer-loop model-assisted active disturbance rejection control includes an extended state observer and a model-assisted active disturbance rejection controller; among them, the extended state observer is:

[0052]

[0053] In the formula, z 1 , z 2 , z 3 are the state estimation values of x 1 , x 2 , x 3 respectively, z 4is in an expanded state and is used to estimate external disturbances, β 1 , β 2 , β 3 , β 4 is the observation gain;

[0054] The model-assisted active disturbance rejection controller is designed as:

[0055]

[0056] where x d is the desired hydraulic cylinder displacement signal, k 1 , k 2 , k 3 are the control gains of the model-assisted active disturbance rejection controller, and u is the desired control input of the outer-loop system.

[0057] Furthermore, the inner-loop sliding mode controller is designed as:

[0058] The inner-loop control error is: e q = q md - q m , where q md is the desired hydraulic fluid flow rate, the magnitude of which is obtained from the outer-loop desired control input. When the system is in the forward motion, q md = q ld / d c3 , when the system is in the reverse motion, q md = q ld / d s3 , q ld is the outer-loop desired control input u;

[0059] The sliding mode surface is selected as: s = e q , and the exponential reaching law is selected for the sliding mode reaching law;

[0060] The duty ratio control law of the high-speed switching valve S1 in the forward motion mode is:

[0061]

[0062] The duty ratio control law of the high-speed switching valve S3 in the reverse motion mode is:

[0063]

[0064] where k 01 , k 02 , ε 1 , ε 2 are the sliding mode controller parameters in the forward and reverse motion processes, and sat(s) is the saturation function to replace the original sign function of the sliding mode control.

[0065] Further, in the bidirectional hydraulic three-state switch boost system, a flow sensor is respectively arranged at the front end of the liquid inductance element L1 to collect the liquid inductance flow rate, a pressure sensor is arranged at the outlet of the accumulator A to collect the load system pressure, and a displacement sensor is arranged at the piston rod of the load hydraulic cylinder H to collect the movement displacement of the hydraulic cylinder. At the same time, the entire control system is equipped with a filter to filter these three collected signals respectively, and the filtered signals are fed back to the control system.

[0066] Compared with the prior art, the present invention has the following beneficial effects: The present invention provides a bidirectional hydraulic three-state switch boost system and its control method. By adding a freewheeling structure, the right-half plane zero problem in the basic circuit of the original boost system is successfully solved. The bidirectional movement of the load is realized by additionally adding a high-speed switch valve and a high-speed one-way valve in the circuit, which broadens the movement range of the load. On this basis, model-assisted active disturbance rejection control and nonlinear sliding mode robust control are respectively applied to control the load outer loop and the transformer inner loop, so that the transformer hydraulic cylinder load has strong anti-interference ability and the entire variable pressure system has good robustness. The bidirectional hydraulic three-state switch boost system and its control method proposed by the present invention can make the hydraulic switch boost transformer better applied to industrial practice and solve some scenario problems that require the application of hydraulic switch boost transformers. Brief Description of the Drawings

[0067] Figure 1 is a schematic structural diagram of the bidirectional hydraulic three-state switch boost system according to an embodiment of the present invention;

[0068] Figure 2 is a schematic diagram of three working stages of the bidirectional hydraulic three-state switch boost system during the forward movement process of the load according to an embodiment of the present invention;

[0069] Figure 3 is a schematic diagram of two working stages of the bidirectional hydraulic three-state switch boost system during the reverse movement process of the load according to an embodiment of the present invention;

[0070] Figure 4 is a schematic diagram of load outer loop modeling according to an embodiment of the present invention;

[0071] Figure 5 is an equivalent schematic diagram of the liquid capacitance, liquid inductance, and liquid resistance of each component in the transformer inner loop according to an embodiment of the present invention;

[0072] Figure 6 is a logic block diagram of the control method of the bidirectional hydraulic three-state switch boost system according to an embodiment of the present invention;

[0073] Figure 7 is a control effect diagram of the displacement of the load hydraulic cylinder without applying disturbance according to an embodiment of the present invention;

[0074] Figure 8It is the control effect diagram of the displacement of the load hydraulic cylinder when a sinusoidal time-varying external load disturbance is applied in the embodiment of the present invention. Specific Embodiments

[0075] The present invention will be further described below in conjunction with the accompanying drawings and embodiments.

[0076] It should be noted that the following detailed description is exemplary and is intended to provide further explanation of the present application. Unless otherwise specified, all technical and scientific terms used herein have the same meaning as commonly understood by those of ordinary skill in the technical field to which the present application belongs.

[0077] It should be noted that the terms used herein are only for describing specific embodiments and are not intended to limit the exemplary embodiments according to the present application. As used herein, unless the context clearly indicates otherwise, the singular forms are also intended to include the plural forms. In addition, it should be understood that when the terms "comprising" and / or "including" are used in this specification, they indicate the presence of features, steps, operations, devices, components, and / or combinations thereof.

[0078] This embodiment provides a bidirectional hydraulic three-state switch boost system, which includes a basic boost circuit capable of realizing the load boost function, a freewheeling branch, a load hydraulic cylinder, and a high-speed switch valve and a high-speed one-way valve for realizing the load reverse movement control. Among them, the basic boost circuit includes a liquid inertia element composed of a hydraulic motor and a flywheel, a liquid capacitance element - an accumulator, a high-frequency response high-speed switch valve and a high-frequency response one-way valve as the switching elements of the system, and the freewheeling branch includes a high-speed switch valve and a high-speed one-way valve.

[0079] Figure 1 It is the structural schematic diagram of the bidirectional hydraulic three-state switch boost system in this embodiment. As Figure 1As shown in the figure, the bidirectional hydraulic three-state switch boost system includes a basic boost circuit, a freewheeling branch, and a load hydraulic cylinder H. The basic boost circuit includes a hydraulic inductor L1, a high-speed switch valve S1, a load-side high-speed check valve C3, a load-side accumulator A, and an auxiliary hydraulic inductor L2. The outlet of the pump source Ps is connected to the hydraulic inductor L1. The outlet of the hydraulic inductor L1 is simultaneously connected to the high-speed switch valve S1 and the load-side high-speed check valve C3. The outlet of the high-speed switch valve S1 is connected to the oil tank T. The outlet of the load-side high-speed check valve C3 is simultaneously connected to the load-side accumulator A and the auxiliary hydraulic inductor L2 for stabilizing the flow rate of the load hydraulic cylinder H. When the system switching frequency is high enough or the load inertia is large enough, this auxiliary hydraulic inductor L2 can be omitted. The outlet of the auxiliary hydraulic inductor L2 is connected to the load hydraulic cylinder H, and the outlet of the load hydraulic cylinder H is connected to the oil tank T. The freewheeling branch is composed of a series connection of a high-speed switch valve S2 and a high-speed check valve C2, and its two ends are connected in parallel to the hydraulic inductor L1. Specifically, the outlet of the hydraulic inductor L1 is connected to the high-speed switch valve S2, the outlet of the high-speed switch valve S2 is connected to the high-speed check valve C2, and the outlet of the high-speed check valve C2 is connected to the inlet of the hydraulic inductor L1. The function of the freewheeling branch is to control the duty cycle of the main hydraulic inductor L1 supplying the load within one switching cycle of the high-speed switch valve, and is used to solve the right-half plane zero problem of the basic boost system. A high-speed switch valve S3 is connected in parallel at both ends of the load-side high-speed check valve C3, and a high-speed check valve C1 is connected in parallel at both ends of the high-speed switch valve S1, which is used to realize the reverse motion control of the load hydraulic cylinder.

[0080] Under the action of an external force, the load hydraulic cylinder H is boosted to the set pressure by the boost system to resist the external load, so that the movement direction of the load hydraulic cylinder H is opposite to the direction of the external load force, which is defined as the forward movement of the load hydraulic cylinder H; if the movement direction of the load hydraulic cylinder H is the same as the direction of the external load force under the drive of the boost system, it is defined as the reverse movement of the load hydraulic cylinder H.

[0081] Figure 2 It is a schematic diagram of the working stage of the bidirectional hydraulic three-state switch boost system during the forward movement of the hydraulic cylinder load in this embodiment. During the forward movement of the load hydraulic cylinder H, the working process of one switching cycle of the bidirectional hydraulic three-state switch boost system is divided into three stages:

[0082] (1) As shown in Figure 2As shown in (a), during the first working stage, the high-speed switching valve S1 and the high-speed switching valve S2 are simultaneously turned on. The duty cycle of the high-speed switching valve S2 within each period is fixed, and its magnitude is determined by the duty cycle of the high-speed switching valve S1 when the system has the maximum static output and the best transmission efficiency, which can be measured by any one of the analytical model, simulation, and experiment. The maximum duty cycle of the high-speed switching valve S1 is less than the duty cycle of the high-speed switching valve S2. When the two switching valves, the high-speed switching valve S1 and the high-speed switching valve S2, are simultaneously turned on, the system hydraulic oil first passes through the high-speed switching valve S1 branch. The flow rate of the hydraulic inductance element L1 increases under the action of the pressure difference before and after. At this time, the flow rate required for the movement of the load hydraulic cylinder H is provided by the accumulator A, and the auxiliary hydraulic inductance L2 plays a role in stabilizing the flow rate.

[0083] (2) As Figure 2 As shown in (b), during the second working stage, the high-speed switching valve S1 is closed, and the high-speed switching valve S2 is still turned on. At this time, the system hydraulic oil continues to flow through the high-speed switching valve S2. Due to the action of the self-hydraulic resistance of the high-speed switching valve S2, the high-speed check valve C2, and the hydraulic inductance element L1, the flow rate of the hydraulic inductance element L1 will slightly decrease. At this time, the flow rate required for the movement of the load hydraulic cylinder H is provided by the accumulator A, and the auxiliary hydraulic inductance L2 plays a role in stabilizing the flow rate.

[0084] (3) As Figure 2 As shown in (c), during the third working stage, the high-speed switching valves S1 and S2 are closed. At this time, the system hydraulic oil flows through the load-side high-speed check valve C3 to the load branch. The energy stored in the hydraulic inductance element L1 is converted into hydraulic energy to supply the load, and the pressure and flow rate of the accumulator A are replenished to maintain the pressure and flow rate required for the movement of the load hydraulic cylinder H in the first two stages within the next switching cycle. In the forward movement mode, only by adjusting the duty cycle of the high-speed switching valve S1 can the flow rate supplied by the transformer to the load be changed, and the movement control of the load hydraulic cylinder H can be realized.

[0085] Figure 3 It is a schematic diagram of the working stages of the bidirectional hydraulic three-state switch boost system during the reverse movement of the hydraulic cylinder load in this embodiment. During the reverse movement of the load hydraulic cylinder H, the working process of a switching cycle of the bidirectional hydraulic three-state switch boost system is divided into two stages:

[0086] (1) As Figure 3 As shown in (a), during the first working stage, the high-speed switching valve S3 is turned on. The hydraulic inductance element L1 accelerates in the reverse direction under the action of the pressure difference between the load high pressure and the pump source pressure, and the reverse flow rate passing through it increases.

[0087] (2) As Figure 3As shown in Figure (b), during the second working stage, the high-speed switching valve S3 is closed, and the high-speed check valve C1 is conducting. The oil tank T compensates for the flow rate of the liquid capacitance element L1 during the second working stage. A certain backpressure is applied to the entire system oil tank to compensate for the conduction pressure drop of the high-speed check valve C1 and prevent cavitation in the system. During this entire process, due to the outflow of the oil in the accumulator A, the load hydraulic cylinder H moves in the reverse direction.

[0088] The design of the control method for the entire variable pressure system first requires establishing the mathematical models of the system load outer loop and the transformer inner loop. The entire system is divided into two parts for control. One part is the outer loop system composed of the accumulator A (liquid capacitance element), the auxiliary liquid inductor L2, and the load hydraulic cylinder H. The other part is the inner loop system composed of the transformer part excluding the outer loop system. Modeling is performed on the inner and outer loop systems respectively. The model-assisted active disturbance rejection control is applied to control the outer loop system to obtain the desired liquid inductor flow rate of the inner loop system. The sliding mode controller is applied to control the inner loop system so that the average flow rate of the liquid inductor element L1 follows the desired liquid inductor flow rate.

[0089] As Figure 4 shown, the system load outer loop includes the accumulator A, the auxiliary liquid inductor L2, and the load hydraulic cylinder H. The modeling of the outer loop system includes the following steps:

[0090] Step S1: According to the force balance equation of the hydraulic cylinder, considering the mass of the hydraulic cylinder piston and the viscous friction coefficient of the hydraulic cylinder, the force balance equation of the load hydraulic cylinder H is established as:

[0091] P 1 A 1 -P t A 2 =ma + B m v + F l

[0092] In the formula, P 1 is the pressure in the rodless cavity of the load hydraulic cylinder H, A 1 is the cross-sectional area of the piston in the rodless cavity of the load hydraulic cylinder H, P t is the pressure in the rod cavity of the load hydraulic cylinder H, which is the oil tank pressure in this embodiment, A 2 is the cross-sectional area of the piston in the rod cavity of the load hydraulic cylinder H, m is the equivalent mass of the piston of the load hydraulic cylinder H and the external load, B m is the viscous damping coefficient of the piston and the external load, F l is the external load force acting on the piston rod, a is the piston movement acceleration, and v is the piston movement speed. The pressure in the rodless cavity of the load hydraulic cylinder H can be obtained as:

[0093]

[0094] Step S2: On the basis of not considering the leakage of the motor, according to the pressure-flow relationship at both ends of the hydraulic motor, the series relationship between the auxiliary liquid inductor L2 and the load hydraulic cylinder H, and the relationship between the input flow of the hydraulic cylinder and the moving speed of the hydraulic cylinder, the pressure difference at both ends of the auxiliary liquid inductor L2 can be obtained as follows:

[0095]

[0096] In the formula, L 2 is the inductance value of the auxiliary liquid inductor, R L2 is the resistance value of the auxiliary liquid inductor, q 1 is the flow rate at the oil inlet of the hydraulic cylinder.

[0097] Among them, the auxiliary liquid inductor value L 2 , and the resistance value R L2 can be determined by the input-output transfer function of the liquid inductor motor. The following is an example of calculating the inductance value. The equivalent liquid capacitance C c of the accumulator working point and the non-linear liquid resistance R s of the high-speed switching valve can both be obtained in this way. The calculation of the pressure drop of the high-speed check valve can adopt the throttling formula when it is fully opened, and can be approximately calculated as the product of the linear liquid resistance and the flow rate when it is not fully opened:

[0098] The following input-output relationship is obtained from the flow continuity equation and the force balance equation of the motor:

[0099]

[0100] In the formula, Q L is the average flow rate of the hydraulic motor; C tm is the total leakage coefficient of the hydraulic motor; P L is the pressure difference at both ends of the hydraulic motor; D m is the displacement of the hydraulic motor; ω m is the angular velocity of the hydraulic motor; V t is the total volume of the two chambers of the hydraulic motor; β e is the bulk modulus of elasticity of the oil; J t is the total inertia of the hydraulic motor and the load reduced to the motor shaft; B L is the viscous damping coefficient of the hydraulic motor and the load; G is the torsional stiffness of the load; θ m is the rotation angle of the hydraulic motor; T L is the external load torque acting on the hydraulic motor.

[0101] Taking the Laplace transform of the above two equations, we can get:

[0102]

[0103] Substituting the Laplace-transformed force balance equation into the flow continuity equation, we can get:

[0104]

[0105] The transfer function of the hydraulic motor pressure flow can be further obtained:

[0106]

[0107] Ignoring the effects of motor leakage and oil compressibility, the fluid sense value of the fluid sense motor is Liquid resistance

[0108] Step S3: According to the pressure relationship of the hydraulic series circuit, the pressure on the accumulator A can be obtained as:

[0109]

[0110] According to the pressure-flow relationship of the accumulator and the parallel relationship between the accumulator and the hydraulic cylinder, it can be obtained that:

[0111]

[0112] In the formula, C c is the equivalent liquid capacity of accumulator A at the working point, q l is the flow rate flowing into the entire outer ring system.

[0113] Step S4: By combining the formulas of steps S1-S3, the system model of the outer loop system can be obtained as follows:

[0114]

[0115] Where P t is the derivative of the tank pressure. The tank pressure is constant, so this term can be ignored. 1 ,x 2 ,x 3 They are the displacement, velocity and acceleration state variables of the hydraulic cylinder respectively.

[0116] Figure 5 This is an equivalent schematic diagram of the liquid capacity, liquid inductance, and liquid resistance of each component of the inner ring system. Figure 2 and Figure 3 Schematic diagram of each working stage of the system. The modeling of the inner loop system includes the following steps:

[0117] Step S1: In the forward motion mode, the mathematical model of each working stage of the transformer can be established as follows:

[0118] Phase 1:

[0119]

[0120] Phase 2:

[0121]

[0122] Stage 3:

[0123]

[0124] In the formula, P in (t) is the pump source pressure, q m (t) is the average flow rate of the liquid sensing element L1 in one switching cycle of the high-speed switching valve, P t (t) is the tank pressure, R s1 、R s2 are the non-linear liquid resistances when the high-speed switching valves S1 and S2 are always open, R L1 is the linear liquid resistance when the liquid sensing element L1 is working, Δp ch2 (t), Δp ch3 (t) are the pressure drops when the high-speed check valves C2 and C3 are working.

[0125] In the forward motion mode, the three working stages of the transformer are unified into a differential equation by the switching cycle averaging method:

[0126]

[0127] In the formula, d s1 is the duty cycle of the high-speed switching valve S1 in one switching cycle, d s2 is the freewheeling duty cycle of the liquid sensing element L1, d c3 is the conduction duty cycle of the high-speed check valve C3 in one switching cycle, d s2 =1 - d s1 - d c3 .

[0128] Verification that the hydraulic three-state switch boost system in the forward motion mode does not contain right-half plane zeros:

[0129] The mathematical relationship between the liquid sensing flow rate and the load outer loop is:

[0130]

[0131] Without considering the external load disturbance, regarding the external load as a constant load, the tank pressure P t is regarded as a constant value, and by performing Laplace transform and Taylor expansion on the equation, we can obtain:

[0132]

[0133] In the formula, is the expansion value of the variable x, representing the disturbance amount within the domain of the variable x. Substituting the pressure equation on the accumulator into the inner loop model in the forward motion mode, we can obtain:

[0134]

[0135] Since the pressure drops of the high-speed switching valve, high-speed check valve, and hydraulic inductor motor have no impact on whether the system has right-half plane zeros, for the convenience of calculation, they are ignored in the model. The model is linearized by Taylor expansion to obtain:

[0136]

[0137] In the formula, since F l , P t , d c3 are all constant values, they are not expanded. Ignoring the DC term and AC small signal terms of the second order and above, the Laplace transform of this formula is performed and the expanded value of the hydraulic inductor flow rate is substituted to obtain:

[0138]

[0139] The pump source pressure is regarded as constant. Therefore the input-output transfer function of the entire system can be obtained:

[0140]

[0141] From the input-output transfer function of the system, it can be seen that the duty ratio of the switching valve to the moving speed of the load hydraulic cylinder is a third-order system that does not contain right-half plane zeros.

[0142] Step S2: In the reverse motion mode, the mathematical models of each working stage of the transformer can be established as follows:

[0143] Stage 1:

[0144]

[0145] Stage 2:

[0146]

[0147] In the reverse motion mode, the two working stages of the transformer are unified into a differential equation by the switching period averaging method:

[0148]

[0149] In the formula, d s3 is the conduction duty ratio of the high-speed switching valve S3 in one switching period, d c1 is the duty ratio of the high-speed check valve C1 in one switching period, d c1 = 1 - d s3 .

[0150] The outer-loop model-assisted active disturbance rejection control contains an extended state observer (ESO) and a model-assisted active disturbance rejection controller (MADRC). The extended state observer can be designed as follows:

[0151]

[0152] where z 1 , z 2 , z 3 are the state estimation values of x 1 , x 2 , x 3 respectively, z 4 is the extended state, used to estimate external disturbances, and β 1 , β 2 , β 3 , β 4 are the observation gains. The part of the load system that has been modeled is added to the extended state observer, which is a feature of the model-assisted active disturbance rejection control. The observation gains can be selected using the bandwidth configuration method, and the specific bandwidth configuration method is as follows:

[0153]

[0154] ω o is the observation bandwidth. When ω o is configured larger, the observer observes the system state more accurately, and the control accuracy will also be improved accordingly. However, an overly large observation bandwidth will introduce unnecessary noise, so the bandwidth cannot be too large.

[0155] In this embodiment, a linear extended state observer is taken as an example, and a non-linear extended state observer can also be used. However, whether a linear or non-linear extended state observer is used, it falls within the protection scope of the present invention.

[0156] The model-assisted active disturbance rejection controller can be designed as follows:

[0157]

[0158] where x d is the desired hydraulic cylinder displacement signal, k 1 , k 2 , k 3 are the control gains of the model-assisted active disturbance rejection controller, u is the desired control input of the outer-loop system, and the part of the load system that has been modeled is also added to the model-assisted active disturbance rejection controller, which is also a feature of the model-assisted active disturbance rejection control.

[0159] The inner-loop sliding mode controller (SMC) can be designed as follows:

[0160] The inner-loop control error is: e q = q md - qm , where q md is the expected liquid flow rate, the magnitude of which is obtained from the expected control input of the outer loop. When the system is in the forward motion, q md = q ld / d c3 . When the system is in the reverse motion, q md = q ld / d s3 . q ld is the expected control input u of the outer loop. The relationship between the above expected flow rates is determined by the relationship between the average flow rate of the liquid sensing element L1 and the average flow rate of the load branch in one switching period of the system.

[0161] The sliding mode surface is selected as: s = e q , and the exponential reaching law is selected for the sliding mode reaching law.

[0162] The duty cycle control law of the high-speed switching valve S1 in the forward motion mode is:

[0163]

[0164] The duty cycle control law of the high-speed switching valve S3 in the reverse motion mode is:

[0165]

[0166] In the formula, k 01 , k 02 , ε 1 , ε 2 are the sliding mode controller parameters in the forward and reverse motion processes, and sat(s) is the saturation function to replace the original sign function of the sliding mode control.

[0167] In this embodiment, the entire control logic of the bidirectional hydraulic three-state switching boost system is as Figure 6 shown.

[0168] In the bidirectional hydraulic three-state switching boost system, a flow sensor is respectively arranged at the front end of the liquid sensing element L1 to collect the liquid sensing flow rate, a pressure sensor is arranged at the outlet of the accumulator A to collect the load system pressure, a displacement sensor is arranged at the piston rod of the load hydraulic cylinder H to collect the hydraulic cylinder movement displacement. At the same time, the entire control system is equipped with a filter to filter these three collected signals respectively, and the filtered signals are fed back to the control system.

[0169] In this embodiment, the parameter selection of each component of the hydraulic three-state switching boost system is as follows:

[0170]

[0171] In this embodiment, with an external force of 7000 N acting on the load hydraulic cylinder (at this time, the load pressure is about 100 bar, close to twice the pressure boost), it is required that the load hydraulic cylinder track a sine curve with a frequency of 0.3 Hz and an amplitude of 0.03 m. Consider whether the variable pressure system can drive the load to achieve a good tracking control accuracy under the action of the controller to verify the effectiveness of the proposed structure and the controller. In the present invention, the duty ratio of the high-speed switching valve S2 is taken as 0.75. Since there is a transition process when the high-speed switching valve switches, to avoid the duty ratio overlap when the high-speed switching valves S1 and S2 are closed, which may lead to a change in the duty ratio of the high-speed one-way valve C3 at the load end in the third stage of the forward mode, causing the reappearance of the zero point in the right half-plane of the system, the maximum duty ratio of the high-speed switching valve S1 is limited to 0.65. In the simulation, the two valves are opened simultaneously. From Figure 7 It can be seen that the load hydraulic cylinder tracks the sine desired curve well, and the maximum tracking error is 0.57 mm.

[0172] Under the same system parameters, a sine external load disturbance with a frequency of 0.4 Hz and an amplitude of 300 N is applied to the hydraulic cylinder to investigate the control ability of the model-assisted active disturbance rejection controller under the external load disturbance. From Figure 8 It can be seen that the load hydraulic cylinder still has good tracking accuracy. Although the maximum tracking error has increased, it is still about 1.4 mm, which is sufficient to verify the effectiveness of the designed structure and the controller.

[0173] The above is only a preferred embodiment of the present invention and is not a limitation of the present invention in other forms. Any person skilled in the art may use the disclosed technical content to make changes or modifications into equivalent embodiments with equivalent changes. However, any simple modification, equivalent change, and modification made to the above embodiments based on the technical essence of the present invention without departing from the technical solution content of the present invention still fall within the protection scope of the technical solution of the present invention.

Claims

1. A bidirectional hydraulic three-state switch boost system, characterized in that: It includes a basic boost circuit, a freewheeling branch and a load hydraulic cylinder H; the basic boost circuit includes a liquid sensing element L1, a high-speed switch valve S1, a load-end high-speed one-way valve C3, a load-end accumulator A and an auxiliary liquid sensor L2, the pump source Ps outlet is connected to the liquid sensing element L1, the liquid sensing element L1 outlet is simultaneously connected to the high-speed switch valve S1 and the load-end high-speed one-way valve C3, the high-speed switch valve S1 outlet is connected to the oil tank T, and the load-end high-speed one-way valve C3 outlet is simultaneously connected to the load-end accumulator A and the auxiliary liquid for stabilizing the flow of the load hydraulic cylinder H The outlet of the auxiliary liquid sensor L2 is connected to the load hydraulic cylinder H, and the outlet of the load hydraulic cylinder H is connected to the oil tank T; the freewheeling branch includes a high-speed switch valve S2 and a high-speed one-way valve C2, the outlet of the liquid sensor element L1 is connected to the high-speed switch valve S2, the outlet of the high-speed switch valve S2 is connected to the high-speed one-way valve C2, and the outlet of the high-speed one-way valve C2 is connected to the inlet of the liquid sensor element L1; the high-speed switch valve S3 is connected in parallel at both ends of the load-end high-speed one-way valve C3, and the high-speed one-way valve C1 is connected in parallel at both ends of the high-speed switch valve S1, which is used to realize the reverse motion control of the load hydraulic cylinder.

2. A control method for a bidirectional hydraulic three-state switch boost system according to claim 1, characterized in that: Under the action of external force, the load hydraulic cylinder H is raised to a set pressure by the booster system to resist the external load, so that the movement direction of the load hydraulic cylinder H is opposite to the direction of the external load force, which is defined as the forward movement of the load hydraulic cylinder H; if the movement direction of the load hydraulic cylinder H is the same as the direction of the external load force driven by the booster system, it is defined as the reverse movement of the load hydraulic cylinder H.

3. The control method of a bidirectional hydraulic three-state switch boost system according to claim 2 is characterized in that: During the forward movement of the load hydraulic cylinder H, the working process of a switching cycle of the bidirectional hydraulic three-state switch boost system is divided into three stages: (1) In the first working stage, the high-speed switch valve S1 and the high-speed switch valve S2 are turned on at the same time. The duty cycle of the high-speed switch valve S2 in each cycle is fixed, and its size is determined by the duty cycle of S2 under each high-speed switch valve S1 duty cycle, when the system has the maximum static output and the best transmission efficiency, and the maximum duty cycle of the high-speed switch valve S1 is smaller than the duty cycle of the high-speed switch valve S2; When the high-speed switch valve S1 and the high-speed switch valve S2 are turned on at the same time, the system oil first passes through the high-speed switch valve S1 branch, and the flow of the liquid sensor L1 increases under the action of the front and rear pressure difference. At this time, the flow required for the movement of the load hydraulic cylinder H is provided by the accumulator A, and the auxiliary liquid sensor L2 plays a role in stabilizing the flow; (2) In the second working stage, the high-speed switch valve S1 is closed, and the high-speed switch valve S2 is still open. At this time, the system hydraulic oil continues to flow through the high-speed switch valve S2. Due to the effect of the high-speed switch valve S2, the high-speed check valve C2 and the hydraulic resistance of the liquid sensing element L1 itself, the flow of the liquid sensing element L1 decreases. At this time, the flow required for the movement of the load hydraulic cylinder H is provided by the accumulator A, and the auxiliary liquid sensing element L2 plays a role in stabilizing the flow; (3) In the third working stage, the high-speed switching valves S1 and S2 are closed. At this time, the system hydraulic oil flows to the load branch through the load-end high-speed check valve C3, and the pressure flow of the accumulator A is replenished to maintain the pressure flow required for the movement of the load hydraulic cylinder H in the first two stages of the system in the next switching cycle.

4. The control method of a bidirectional hydraulic three-state switch boost system according to claim 2, characterized in that: During the reverse movement of the load hydraulic cylinder H, the working process of a switching cycle of the bidirectional hydraulic three-state switch boost system is divided into two stages: (1) In the first working stage, the high-speed switch valve S3 is turned on, and the liquid sensing element L1 is accelerated in the reverse direction under the action of the load high pressure and the pump source pressure difference, and the reverse flow rate increases; (2) In the second working stage, the high-speed switch valve S3 is closed, and the high-speed check valve C1 is turned on. The flow of the liquid sensing element L1 in the second working stage is compensated by the oil tank T. The entire system oil tank is added with back pressure to compensate for the conduction pressure drop of the high-speed check valve C1 to prevent cavitation in the system. During the whole process, due to the outflow of oil from the accumulator A, the load hydraulic cylinder H moves in the opposite direction.

5. The control method of a bidirectional hydraulic three-state switch boost system according to claim 2, characterized in that: The whole system is divided into two parts for control. One part is the outer loop system composed of accumulator A, auxiliary fluid inductor L2 and load hydraulic cylinder H, and the other part is the inner loop system composed of the transformer part outside the outer loop system. The inner and outer loop systems are modeled separately, and the outer loop system is controlled by model-assisted anti-disturbance control to obtain the expected fluid inductance flow of the inner loop system. The inner loop system is controlled by a sliding mode controller so that the average flow of the fluid inductor L1 follows the expected fluid inductance flow.

6. The control method of a bidirectional hydraulic three-state switch boost system according to claim 5, characterized in that: Modeling of the outer loop system includes the following steps: Step S1: According to the force balance equation of the hydraulic cylinder, the force balance equation of the load hydraulic cylinder H is established by taking into account the mass of the hydraulic cylinder piston and the viscous friction coefficient of the hydraulic cylinder: P1A1-P t A2=ma+B m v+F l Where P1 is the pressure of the rodless chamber of the load hydraulic cylinder H, A1 is the cross-sectional area of ​​the piston of the rodless chamber of the load hydraulic cylinder H, and P t is the pressure of the rod chamber of the load hydraulic cylinder H, A2 is the cross-sectional area of ​​the piston of the rod chamber of the load hydraulic cylinder H, m is the equivalent mass of the piston of the load hydraulic cylinder H and the external load, B m is the viscous damping coefficient between the piston and the external load, F l is the external load force acting on the piston rod, a is the piston acceleration, and v is the piston speed; the rodless chamber pressure of the load hydraulic cylinder H is: Step S2: Without considering the motor leakage, according to the pressure-flow relationship at both ends of the hydraulic motor, the series relationship between the auxiliary fluid sensor L2 and the load hydraulic cylinder H, and the relationship between the hydraulic cylinder input flow and the hydraulic cylinder movement speed, the pressure difference at both ends of the auxiliary fluid sensor L2 is obtained as: Where, L2 is the liquid sense value of auxiliary liquid sense, R L2 is the hydraulic resistance value of the auxiliary fluid sensor, q1 is the flow rate of the hydraulic cylinder oil inlet; Step S3: According to the pressure relationship of the hydraulic series circuit, the pressure on the accumulator A is obtained as: According to the pressure-flow relationship of the accumulator and the parallel relationship between the accumulator and the hydraulic cylinder, it can be obtained that: In the formula, C c is the equivalent liquid capacity of accumulator A at the working point, q l is the flow into the entire outer ring system; Step S4: Combining the formulas of steps S1-S3, the system model of the outer loop system is obtained as follows: In the formula, is the derivative of the tank pressure. The tank pressure is constant, so this term can be ignored. x1, x2, and x3 are the displacement, velocity, and acceleration state variables of the hydraulic cylinder, respectively.

7. The control method of a bidirectional hydraulic three-state switch boost system according to claim 5, characterized in that: The modeling of the inner loop system includes the following steps: Step S1: In the forward motion mode, the mathematical model of each working stage of the transformer is established as follows: Phase 1: Phase 2: Phase 3: Where P in (t) is the pump source pressure, q m (t) is the average flow rate of the liquid sensing element L1 in one switching cycle of the high-speed switching valve, P t (t) is the tank pressure, R s1 , R s2 is the nonlinear fluid resistance when the high-speed switch valves S1 and S2 are normally open, R L1 is the linear liquid resistance of the liquid sensing element L1 when it is working, Δp ch2 (t), Δp ch3 (t) is the pressure drop when high-speed check valves C2 and C3 are working; In the forward motion mode, the three working stages of the transformer are unified into one differential equation by the switching cycle average method: Where, d s1 is the duty cycle of a switching cycle of the high-speed switching valve S1, d s2 is the continuous current duty cycle of the liquid sensing element L1, d c3 is the on-duty ratio of a switching cycle of the high-speed one-way valve C3, d s2 =1-d s1 -d c3 ; Step S2: In the reverse motion mode, the mathematical model of each working stage of the transformer is established as follows: Phase 1: Phase 2: In the reverse motion mode, the two working phases of the transformer are unified into one differential equation by the switching cycle average method: Where, d s3 is the on-duty ratio of a switching cycle of the high-speed switching valve S3, d c1 is the duty cycle of a switching cycle of the high-speed one-way valve C1, d c1 =1-d s3 .

8. The control method of a bidirectional hydraulic three-state switch boost system according to claim 5, characterized in that: The outer loop model-assisted active disturbance rejection control includes an extended state observer and a model-assisted active disturbance rejection controller; the extended state observer is: Where z1, z2, z3 are the state estimates of x1, x2, x3 respectively, z4 is the expansion state used to estimate external disturbances, β1, β2, β3, β4 are observation gains; The model-assisted active disturbance rejection controller is designed as: In the formula, x d is the expected hydraulic cylinder displacement signal, k1, k2, k3 are the control gains of the model-assisted ADRC, and u is the expected control input of the outer loop system.

9. The control method of a bidirectional hydraulic three-state switch boost system according to claim 5, characterized in that: The inner loop sliding mode controller is designed as: The inner loop control error is: e q =q md -q m , where q md is the expected liquid flow rate, whose size is obtained by the expected control input of the outer loop. When the system is in forward motion, q md =q ld / d c3 , when the system is in reverse motion, q md =q ld / d s3 ,q ld is the outer loop desired control input u; The sliding surface is selected as: s = e q , the sliding mode reaching law is selected as the exponential reaching law; The duty cycle control law of the high-speed switching valve S1 in the forward motion mode is: The duty cycle control law of the high-speed switching valve S3 in the reverse motion mode is: In the formula, k 01 ,k 02 ,ε1,ε2 are the sliding mode controller parameters for the forward and reverse motion processes, and sat(s) is the saturation function that replaces the original sign function of the sliding mode control.

10. The control method of a bidirectional hydraulic three-state switch boost system according to claim 5, characterized in that: In the bidirectional hydraulic three-state switch boost system, a flow sensor is set at the front end of the liquid sensing element L1 to collect the liquid sensing flow, a pressure sensor is set at the outlet of the accumulator A to collect the load system pressure, and a displacement sensor is set at the piston rod of the load hydraulic cylinder H to collect the movement displacement of the hydraulic cylinder. At the same time, the entire control system is equipped with a filter to filter the three collected signals respectively and feed the filtered signals back to the control system.

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