Control system and control method of load port independent control valve based on pilot hydraulic control

Abstract

The invention belongs to the technical field of hydraulic actuator control, and particularly relates to a control system and a control method of a load port independent control valve based on pilot hydraulic control. The control system of the present invention comprises: a hydraulic actuator; an oil tank and a load port independent control valve; a plurality of flow sensors and pressure sensors; the control module is used for controlling the internal on-off of the load port independent control valve according to a control strategy; the load port independent control valve comprises a valve body, a main pressure reducing valve installed on the valve body, a first pressure compensation pressure reducing module, a second pressure compensation pressure reducing module and a shuttle valve. First-stage pressure reduction is conducted on hydraulic oil through the main pressure reducing valve, second-stage pressure reduction or pressure compensation is conducted on an oil inlet and an oil outlet of the hydraulic actuator through the two independent pressure compensation pressure reducing modules, and therefore the problems that in the working process of a traditional hydraulic system, when the load of the actuator changes, control modes need to be frequently switched, and the working efficiency is high are solved. And meanwhile, the structure of the valve body is simplified.

Description

Technical Field

[0001] The present invention belongs to the technical field of hydraulic actuator control, and in particular relates to a control system and a control method for a load port independent control valve based on pilot hydraulic control. Background Art

[0002] As the core power transmission method in modern industry, hydraulic transmission technology plays an irreplaceable role in engineering machinery, mining equipment, aerospace, and other fields. However, with the continuous improvement of industrial automation levels and increasingly stringent energy conservation and environmental protection requirements, the technical limitations of traditional hydraulic systems in complex operating conditions are gradually becoming apparent. These are mainly manifested in problems such as large energy loss, low control accuracy, and slow response speed. These problems are particularly prominent under conditions with frequent load changes.

[0003] Traditional hydraulic valve control systems generally utilize a single spool displacement adjustment mechanism. This one-size-fits-all control approach fails to independently optimize the inlet and outlet throttle ports based on actual operating conditions, resulting in throttling energy losses of up to 30%-40%. Using a single spool displacement mechanism to adjust the throttling area of ​​the inlet and outlet ports prevents independent control of the ports based on actual operating conditions, leading to excessive throttling losses and the inability to achieve flow regeneration. Furthermore, traditional systems require frequent switching of control modes in response to load changes, increasing control complexity and the risk of failure. Summary of the Invention

[0004] The present invention provides a control system and control method for a load-port independent control valve based on pilot hydraulic control, so as to solve the problems that in the operation of traditional hydraulic systems, the actuator needs to frequently switch the control mode when the load changes, which leads to complex system control and high failure risk. It also improves the energy saving and control performance of the system and simplifies the valve body structure.

[0005] The technical solution adopted by the present invention to solve its technical problem is:

[0006] As a first aspect, the present invention provides a control system for a load port independent control valve based on pilot hydraulic control, comprising:

[0007] Hydraulic actuators;

[0008] An oil tank and a load port independent control valve; the oil tank is in fluid communication with the load port independent control valve via a hydraulic pump, and the output port of the load port independent control valve is in fluid communication with the hydraulic actuator;

[0009] A plurality of flow sensors and pressure sensors are distributed inside the load port independent control valve;

[0010] a control module for controlling the internal on-off of the load port independent control valve according to a control strategy; the control module includes a rod chamber flow PID control unit, a rod chamber pressure PID control unit, a rodless chamber flow PID control unit, and a rodless chamber pressure PID control unit of the hydraulic actuator;

[0011] The load port independent control valve includes a valve body, a main pressure-reducing valve installed on the valve body, a first pressure-compensating pressure-reducing module, a second pressure-compensating pressure-reducing module, and a shuttle valve arranged on the output oil circuits of the first pressure-compensating pressure-reducing module and the second pressure-compensating pressure-reducing module; the output oil ports of the first pressure-compensating pressure-reducing module and the second pressure-compensating pressure-reducing module are both connected to the oil inlet and oil outlet oil circuits of the hydraulic actuator; the oil outlet of the shuttle valve is connected to the Ls end of the hydraulic pump.

[0012] Specifically, the valve body includes a main oil inlet circuit P, a main oil return circuit T, an Ls port for working pressure feedback, and a first working oil port A and a second working oil port B;

[0013] The oil inlet P of the main pressure reducing valve Z , the pressure compensation oil inlet P of the first pressure compensation and pressure reducing module Y1 , the pressure compensation oil inlet P of the second pressure compensation pressure reducing module Y2 The oil return port T of the main pressure reducing valve is connected to the main oil inlet oil circuit P. Z , the oil return port T of the first pressure compensation and pressure reducing module Y1 , the oil return port T of the second pressure compensation and pressure reducing module Y2 Both are connected to the main oil return line T liquid line;

[0014] The first working oil port A of the main pressure reducing valve Z Respectively connected to the pressure reducing oil inlet P of the first pressure compensating pressure reducing module J1 and the pressure reducing oil inlet P of the second pressure compensating pressure reducing module J2 The first working oil port A of the first pressure compensation and pressure reducing module Y1 The first working oil port A of the valve body is connected to the oil circuit, and the first working oil port A of the second pressure compensation and pressure reducing module is connected to the oil circuit. Y2 The second working oil port B of the valve body is connected to the oil circuit; the oil outlet of the shuttle valve is connected to the Ls port liquid circuit of the working pressure feedback.

[0015] Specifically, the first pressure-compensating and reducing pressure module includes a first pressure-compensating valve, a first proportional reducing valve and a first two-position three-way main valve core; the oil inlet of the first pressure-compensating valve is connected to the main oil inlet oil circuit of the valve body, and the oil outlet of the first pressure-compensating valve is connected to the oil inlet liquid circuit of the first two-position three-way main valve core; the oil inlet of the first proportional reducing valve is connected to the oil outlet liquid circuit of the main reducing valve, and the oil outlet of the first proportional reducing valve is connected to the PP end liquid circuit of the first two-position three-way main valve core; the oil return port of the first proportional reducing valve and the oil return port of the first two-position three-way main valve core are both connected to the main return oil circuit T of the valve body.

[0016] Specifically, the second pressure-compensating and reducing pressure module includes a second pressure-compensating valve, a second proportional reducing valve and a second two-position three-way main valve core; the oil inlet of the second pressure-compensating valve is connected to the main oil inlet oil circuit of the valve body, and the oil outlet of the second pressure-compensating valve is connected to the oil inlet liquid circuit of the second two-position three-way main valve core; the oil inlet of the second proportional reducing valve is connected to the oil outlet liquid circuit of the main reducing valve, and the oil outlet of the second proportional reducing valve is connected to the PP end liquid circuit of the second two-position three-way main valve core; the oil return port of the second proportional reducing valve and the oil return port of the second two-position three-way main valve core are both connected to the main return oil circuit T of the valve body.

[0017] As a second aspect, the present application provides a control method for a control system of a load port independent control valve based on pilot hydraulic control, comprising the following steps:

[0018] Establish connections between the load port independent control valve, the oil tank, the hydraulic actuator, and several pressure sensors and flow sensors; the first working oil port A of the load port independent control valve is in communication with the rodless chamber fluid path of the hydraulic actuator, and the oil inlet port B of the load port independent control valve is in communication with the rod chamber fluid path of the hydraulic actuator;

[0019] Based on the external load force F applied to the hydraulic actuator i , establish the force balance equation of the hydraulic actuator; use the force balance equation of the hydraulic actuator and the logic judgment function to identify the positive and negative loads and determine the working mode;

[0020] Corresponding to different working modes, a flow equation for the flow out of the rod chamber of the hydraulic actuator is established based on the valve port pressure flow equation;

[0021] The displacements of the first and second two-position three-way main valve cores in the corresponding working mode are obtained through the hydraulic actuator force balance equation and the valve port pressure and flow equation.

[0022] The control module controls the displacement of the first two-position three-way main valve core and the second two-position three-way main valve core to control the inlet and outlet flow of the hydraulic controller.

[0023] Specifically, the operating mode includes an impedance operating mode and an overrunning load operating mode;

[0024] In the impedance working mode, F i >0, the control of the hydraulic actuator is as follows:

[0025] Regulate the rodless cavity flow PID control unit to control the flow of the rodless cavity of the hydraulic actuator;

[0026] Set the pressure PID control unit output corresponding to the rodless cavity of the hydraulic actuator to be invalid;

[0027] By setting the target flow, using the valve port flow-pressure differential formula and PID closed-loop control, the flow of the rodless chamber of the hydraulic actuator is actively controlled;

[0028] The loss of the pressure difference before and after the first proportional pressure reducing valve is mechanically compensated by the first pressure compensating valve, and the pressure difference before and after the first proportional pressure reducing valve is set to a constant value, so that the flow change and the valve opening present a linear relationship;

[0029] Set the rod cavity flow PID control unit to be invalid and the pressure PID control unit output corresponding to the rod cavity to be valid, and perform pressure control on the rod cavity;

[0030] By comparing the set target pressure difference with the actual pressure difference, a pressure closed-loop control is formed to actively control the pressure in the rod cavity, adjust the opening change of the second proportional pressure reducing valve port, and change the front and rear pressure difference of the second proportional pressure reducing valve port.

[0031] Specifically, under the overrunning load condition, F i <0, the control of the hydraulic actuator is as follows:

[0032] When the load changes from positive to negative, the rod chamber of the hydraulic actuator switches from pressure control to flow control, and the flow of the rod chamber is actively controlled by using the valve port flow-pressure difference formula and PID closed-loop control;

[0033] The loss of pressure difference before and after the second proportional pressure reducing valve is mechanically compensated by the second pressure compensating valve. The pressure difference before and after the second proportional pressure reducing valve is set to a fixed value, so that the flow change and the valve opening present a linear relationship.

[0034] Set the rodless cavity flow PID control unit to be invalid and the rodless cavity pressure PID control unit output corresponding to the rodless cavity to be valid, and perform pressure control on the rod cavity.

[0035] Specifically, the use of the hydraulic actuator force balance equation and the logic judgment function to identify the positive and negative loads and determine the working mode specifically includes:

[0036] Using the logic function H(F i)、G(F i ), outputs 0 or 1 value according to the load direction to control whether the corresponding flow control PID and pressure control PID are valid; when valid, the corresponding flow and pressure are actively controlled;

[0037] The logic function is:

[0038]

[0039] The force balance equation of the hydraulic actuator is: p1A1-p2A2=F l ;

[0040] Among them, p1 represents the rodless cavity pressure of the hydraulic actuator, A1 represents the rodless cavity area of ​​the actuator, p2 represents the rod cavity pressure of the hydraulic actuator, A2 represents the rod cavity area of ​​the actuator, F i Indicates the load force of the hydraulic actuator.

[0041] Specifically, the flow equation out of the rod chamber of the hydraulic actuator is:

[0042] Where, Q2 represents the rod cavity flow, c d represents the hydraulic system flow coefficient, ω represents the area gradient, x2 represents the throttle valve core displacement, and ρ represents the fluid density;

[0043] Then the throttle valve core displacement x2 formula is:

[0044]

[0045] Where, Q2 represents the rod cavity flow, c d Indicates the flow coefficient of the hydraulic system, ω indicates the area gradient, p1 indicates the rodless cavity pressure of the hydraulic actuator, A1 indicates the rodless cavity area of ​​the actuator, p2 indicates the rod cavity pressure of the hydraulic actuator, A2 indicates the rod cavity area of ​​the actuator, F i Indicates the load force of the hydraulic actuator.

[0046] Specifically, when the speed V of the hydraulic actuator is not fixed, the target pressure difference is always calculated by the target pressure. When the pressure control is performed in the rodless chamber, a minimum pressure difference value Δp is set. 1min ; When the rod cavity is used for pressure control, a minimum pressure difference value Δp is set 2min ;

[0047] The calculation formula of Δp1min is:

[0048]

[0049] The calculation formula of Δp2min is:

[0050]

[0051] In the formula, ps represents the oil supply pressure of the system, that is, the input pressure of the main pressure reducing valve; p0 represents the return oil pressure or tank pressure of the system; p 1min Indicates the minimum working pressure of the rodless cavity; p 2min Indicates the minimum working pressure of the rod chamber.

[0052] The beneficial effects of the control system and control method of the load port independent control valve based on pilot hydraulic control of the present invention are:

[0053] The independent load port control system of the present invention utilizes a specially designed independent load port control valve. A main pressure reducing valve provides primary pressure reduction for hydraulic oil, and two independent pressure compensating and reducing modules provide secondary pressure reduction or compensation at the inlet and outlet of the hydraulic actuator. The independent load port control system employs a control module to calculate a flow feedback strategy for precise actuator speed control. Multiple sensors detect the pressure differential across the proportional pressure reducing valve in the independent load port control valve, and an appropriate flow coefficient is selected to estimate the actual flow at the valve port. The actual flow is used as negative feedback and subtracted from the input target flow, thereby achieving closed-loop flow control of the proportional pressure reducing valve. A pressure feedback control strategy is also employed. The pressure differential across the main proportional pressure reducing valve is measured in real time and compared with the target pressure differential in a closed-loop manner. The pressure differential is then varied by controlling the valve opening of the proportional pressure reducing valve. The innovative use of independent inlet / outlet throttle control technology and intelligent flow regeneration solves the problem of complex control and high failure risk caused by frequent actuator control mode switching during load changes during operation of conventional hydraulic systems. This improves the system's energy efficiency and control performance while simplifying the valve structure.

[0054] The present invention establishes a connection relationship between a load port independent control valve, an oil tank, a hydraulic actuator, and a plurality of pressure sensors and flow sensors, based on the external load force F i, establish a force balance equation for the hydraulic actuator; use the hydraulic actuator force balance equation and the logic judgment function to identify the positive and negative loads, determine the working mode, and corresponding to different working modes, establish the rod chamber flow equation out of the hydraulic actuator according to the valve port pressure flow equation, and obtain the displacement of the first two-position three-way main valve core and the second two-position three-way main valve core under the corresponding working mode through the hydraulic actuator force balance equation and the valve port pressure flow equation; control the displacement of the first two-position three-way main valve core and the second two-position three-way main valve core through the control module to control the inlet and outlet flow of the hydraulic controller. The present invention adopts the control strategy of load force direction prediction and two-chamber pressure switching to study the feasibility of switching the working mode at the moment of load direction change, thereby solving the problems of complex control and application limitations in the existing technology. The control strategy of load force direction prediction of the present application can reduce the number of mode switching times, can maintain the continuity of the actuator operation to the greatest extent, ensure higher control accuracy and higher dynamic response, and promote the development of hydraulic transmission technology towards high efficiency, precision and intelligence. BRIEF DESCRIPTION OF THE DRAWINGS

[0055] The present invention will be further described in detail below with reference to the accompanying drawings and specific embodiments.

[0056] Figure 1 1 is a connection principle diagram of a control system in an embodiment of the present invention.

[0057] Figure 2 1 is an internal schematic diagram of a load port independent control valve according to an embodiment of the present invention.

[0058] Figure 3 This is a control principle diagram of a hydraulic actuator in an embodiment of the present invention, in which the first working oil port A is an oil inlet and the second working oil port B is an oil return port.

[0059] Figure 4 This is a control principle diagram of a hydraulic actuator in an embodiment of the present invention, in which the first working oil port A is an oil return port and the second working oil port B is an oil inlet port.

[0060] Figure 5 It is a flow chart of the control method steps of the control system in an embodiment of the present invention.

[0061] Figure 6 This is a flow chart of controlling the hydraulic actuator in the impedance working mode according to an embodiment of the present invention.

[0062] Figure 7 Schematic diagram of the impedance load condition control strategy of the load port independent control system in an embodiment of the present invention.

[0063] Figure 8 Schematic diagram of the overriding load condition control strategy of the load port independent control system in an embodiment of the present invention.

[0064] Figure 9 This is a control block diagram based on load force direction prediction and two-chamber pressure switching in an embodiment of the present invention.

[0065] In the figure: 1. Hydraulic actuator, 2. Oil tank, 3. Load port independent control valve, 31. Valve body, 32. Main pressure reducing valve, 33. First pressure compensating and pressure reducing module, 331. First pressure compensating valve, 332. First proportional pressure reducing valve, 333. First two-position three-way main valve core, 34. Second pressure compensating and pressure reducing module, 341. Second pressure compensating valve, 342. Second proportional pressure reducing valve, 343. Second two-position three-way main valve core, 35. Shuttle valve, 5. Hydraulic pump. DETAILED DESCRIPTION

[0066] The present invention will now be described in further detail with reference to the accompanying drawings, which are simplified schematic diagrams that illustrate the basic structure of the present invention in a schematic manner.

[0067] like Figure 1-Figure 4 The specific embodiment of the control system of the load port independent control valve 3 based on pilot hydraulic control of the present invention shown includes a hydraulic actuator 1, an oil tank 2 and a load port independent control valve 3, a plurality of flow sensors and pressure sensors and a control module, wherein the oil tank 2 is connected to the load port independent control valve 3 through a hydraulic pump 5, the output port of the load port independent control valve 3 is connected to the hydraulic actuator 1, and a plurality of flow sensors and pressure sensors are distributed inside the load port independent control valve 3. The control module controls the internal on and off of the load port independent control valve 3 according to the control strategy; the control module includes a rod chamber flow PID control unit, a rod chamber pressure PID control unit, a rodless chamber flow PID control unit and a rodless chamber pressure PID control unit of the hydraulic actuator 1.

[0068] Specifically, the load port independent control valve 3 includes a valve body 31, a main pressure reducing valve 32 installed on the valve body 31, a first pressure compensating pressure reducing module 33, a second pressure compensating pressure reducing module 34, and a shuttle valve 35 provided on the output oil circuits of the first pressure compensating pressure reducing module 33 and the second pressure compensating pressure reducing module 34; the output oil ports of the first pressure compensating pressure reducing module 33 and the second pressure compensating pressure reducing module 34 are both connected to the oil inlet and oil outlet oil circuits of the hydraulic actuator 1. The pressure sensors in this embodiment are respectively placed on the inlet and outlet of the main valve, the pump outlet and the oil return line of the oil tank to detect the working pressures p1 and p2 of the two chambers of the actuator and the system oil supply pressure p p and oil return back pressure p T ; The flow sensor in this embodiment is placed at the inlet and outlet of the main valve to directly detect the flow entering / out of the actuator.

[0069] It is further explained that the valve body 31 in this embodiment includes a main oil inlet circuit P, a main oil return circuit T, an Ls port for working pressure feedback, and a first working oil port A and a second working oil port B. The oil inlet port P of the main pressure reducing valve 32 Z , the pressure compensation oil inlet P of the first pressure compensation and pressure reducing module 33 Y1 , the pressure compensation oil inlet P of the second pressure compensation and pressure reducing module 34 Y2 The oil return port T of the main pressure reducing valve 32 is connected to the main oil inlet oil circuit P. Z , the oil return port T of the first pressure compensation and pressure reducing module 33 Y1 , the oil return port T of the second pressure compensating and reducing module 34 Y2 The first working oil port A of the main pressure reducing valve 32 is connected to the main oil return line T. Z Respectively connected to the pressure reducing oil inlet P of the first pressure compensating pressure reducing module 33 J1 and the decompression oil inlet P of the second pressure compensating and decompression module 34 J2 The first working oil port A of the first pressure compensating pressure reducing module 33 Y1 The first working oil port A of the valve body 31 is connected to the first working oil port A of the second pressure compensation and pressure reducing module 34. Y2 The second working oil port B of the valve body 31 is connected to an oil circuit, and the oil outlet of the shuttle valve 35 is connected to the Ls port of the valve body 31 for feedback of the working pressure.

[0070] The main pressure-reducing valve 32 reduces the pressure of the oil output from the hydraulic pump 5 in a single stage, minimizing pressure fluctuations and transferring the pressure to two proportional pressure-reducing valves: the first proportional pressure-reducing valve 332 and the second proportional pressure-reducing valve 342. These proportional pressure-reducing valves control the displacement of the main valve core by controlling the relationship between pressure and springs. Two pressure-compensating valves (the first pressure-compensating valve 331 and the second pressure-compensating valve 341) maintain a constant pressure differential between the inlet and outlet of the two-position, three-way main valve core, thereby ensuring a linear relationship between the opening degree of the two-position, three-way main valve core and the flow rate. The shuttle valve 35 feeds back the highest operating pressure in the actuator's two chambers to the Ls terminal of the hydraulic pump 5, ensuring that the hydraulic pump 5 delivers the required flow rate to the system. In this embodiment, the hydraulic pump 5 is a load-sensing pump.

[0071] The load port independent control system in this embodiment employs a specially designed load port independent control valve 3. A main pressure reducing valve 32 is used to reduce hydraulic oil pressure in a primary stage. Two independent pressure compensating and reducing modules are then used to provide secondary pressure reduction or compensation at the oil inlet and outlet of the hydraulic actuator 1. The load port independent control system employs a control module to calculate a flow feedback strategy to precisely control the speed of the hydraulic actuator 1. Multiple sensors are used to detect the pressure differential across the proportional pressure reducing valve in the load port independent control valve 3. An appropriate flow coefficient is selected to estimate the actual flow rate at the valve port. The actual flow rate is used as negative feedback and subtracted from the input target flow rate, thereby achieving closed-loop flow control of the proportional pressure reducing valve. A pressure feedback control strategy is also employed. The pressure differential across the main proportional pressure reducing valve is detected in real time and compared with the target pressure differential in a closed-loop manner. The pressure differential across the valve port is then varied by controlling the valve opening of the proportional pressure reducing valve. The innovative use of independent inlet / outlet throttle control technology and intelligent flow regeneration solves the problem of complex system control and high failure risk caused by the frequent switching of control modes required by actuators when load changes during operation in conventional hydraulic systems. This improves the system's energy efficiency and control performance while simplifying the structure of the valve body 31.

[0072] like Figure 2 As shown, the first pressure-compensating and reducing pressure module 33 in this embodiment includes a first pressure-compensating valve 331, a first proportional reducing valve 332 and a first two-position three-way main valve core 333; the oil inlet of the first pressure-compensating valve 331 is fluidly connected to the main oil inlet oil circuit of the valve body 31, and the oil outlet of the first pressure-compensating valve 331 is fluidly connected to the oil inlet of the first two-position three-way main valve core 333; the oil inlet of the first proportional reducing valve 332 is fluidly connected to the oil outlet of the main reducing valve 32, and the oil outlet of the first proportional reducing valve 332 is fluidly connected to the PP end of the first two-position three-way main valve core 333; the oil return port of the first proportional reducing valve 332 and the oil return port of the first two-position three-way main valve core 333 are both fluidly connected to the main return oil circuit T of the valve body 31.

[0073] Specifically, the second pressure-compensating and reducing pressure module 34 in this embodiment includes a second pressure-compensating valve 341, a second proportional reducing valve 342 and a second two-position three-way main valve core 343; the oil inlet of the second pressure-compensating valve 341 is fluidly connected to the main oil inlet oil circuit of the valve body 31, and the oil outlet of the second pressure-compensating valve 341 is fluidly connected to the oil inlet of the second two-position three-way main valve core 343; the oil inlet of the second proportional reducing valve 342 is fluidly connected to the oil outlet of the main reducing valve 32, and the oil outlet of the second proportional reducing valve 342 is fluidly connected to the PP end of the second two-position three-way main valve core 343; the oil return port of the second proportional reducing valve 342 and the oil return port of the second two-position three-way main valve core 343 are both fluidly connected to the main return oil circuit T of the valve body 31.

[0074] In a specific application, when oil flows into the first working oil port A and oil returns from the second working oil port B, the main reducing valve 32 reduces the pump source pressure by one level and transmits the pressure to the first proportional reducing valve 332; the first proportional reducing valve 3321 transmits the second level reducing pressure to the first two-position three-way main valve core 333, compresses the spring of the first two-position three-way main valve core 333 to generate displacement, so that the high-pressure flow flows into the rodless cavity of the hydraulic cylinder, and the first pressure compensation valve 331 compensates the inlet and outlet oil ports of the first two-position three-way main valve core 333; the low-pressure oil in the rod cavity of the hydraulic cylinder directly returns oil through the second two-position three-way main valve core 343. During this process, the second pressure compensation valve 341 and the second proportional reducing valve 342 do not work; the shuttle valve 35 introduces the high pressure of the rodless cavity of the hydraulic cylinder into the Ls end of the valve group. Its working principle is as follows Figure 3 shown.

[0075] In a specific application, when oil flows into the second working oil port B and oil returns from the first working oil port A, the main reducing valve 32 reduces the pump source pressure by one level and transmits the pressure to the second proportional reducing valve 342; the second proportional reducing valve 342 transmits the two-level reducing pressure to the second two-position three-way main valve core 343, compresses the spring of the second two-position three-way main valve core 343 to generate displacement, so that the high-pressure flow flows into the rod chamber of the hydraulic cylinder, and the second pressure compensation valve 341 compensates the inlet and outlet oil ports of the second two-position three-way main valve core 343; the low-pressure oil in the rodless chamber of the hydraulic cylinder directly returns to the oil through the first two-position three-way main valve core 333. During this process, the first pressure compensation valve 331 and the second proportional reducing valve 342 do not work; the shuttle valve 35 introduces the high pressure of the rod chamber of the hydraulic cylinder into the Ls end of the valve group, and its working principle is as follows Figure 4 shown.

[0076] The load-port independent control system in this embodiment uses a calculated flow feedback strategy to precisely control actuator speed. Multiple sensors detect the pressure differential across the control valve, select an appropriate flow coefficient, and estimate the actual flow rate at the valve port. This actual flow rate is used as negative feedback and then subtracted from the input target flow rate, thereby achieving closed-loop flow control of the proportional pressure reducing valve. Actuator pressure control typically utilizes a pressure feedback control strategy. This strategy measures the pressure differential across the proportional pressure reducing valve in real time, compares it with the target pressure differential in a closed-loop manner, and modulates the pressure differential by controlling the valve opening.

[0077] like Figure 5 As shown, the control method based on the above-mentioned load port independent control system includes the following steps:

[0078] S1: Establish connections between the load port independent control valve 3, the oil tank 2, the hydraulic actuator 1, and several pressure sensors and flow sensors; the first working oil port A of the load port independent control valve 3 is connected to the rodless chamber fluid path of the hydraulic actuator 1, and the oil inlet B of the load port independent control valve 3 is connected to the rod chamber fluid path of the hydraulic actuator 1;

[0079] S2: Based on the external load force F applied to the hydraulic actuator 1 i , establish the force balance equation of the hydraulic actuator 1; use the force balance equation of the hydraulic actuator 1 and the logic judgment function to identify the positive and negative loads and determine the working mode;

[0080] S3: Corresponding to different working modes, according to the valve port pressure flow equation, establish the rod chamber flow equation of the hydraulic actuator 1;

[0081] S4: Calculate the displacements of the first two-position three-way main valve core 333 and the second two-position three-way main valve core 343 in the corresponding working mode through the force balance equation of the hydraulic actuator 1 and the valve port pressure and flow equation;

[0082] S5: The control module controls the displacement of the first two-position three-way main valve core 333 and the second two-position three-way main valve core 343 to control the inlet and outlet flow of the hydraulic controller.

[0083] The working modes in this embodiment include an impedance working mode and an overrunning load working mode. In the impedance extension working mode of the hydraulic actuator 1, the first proportional pressure reducing valve 332 is used to control the speed of the hydraulic actuator 1, and the second proportional pressure reducing valve 342 is used to control the back pressure of the hydraulic actuator 1 to reduce energy consumption. The block diagram of the pressure and flow composite control strategy is shown in FIG. Figure 6 As shown, when in impedance working mode, F i >0, the control of hydraulic actuator 1 is as follows:

[0084] A1: Regulates the rodless cavity flow PID control unit to control the flow of the rodless cavity of hydraulic actuator 1;

[0085] A2: Set the pressure PID control unit output corresponding to the rodless chamber of hydraulic actuator 1 to be invalid;

[0086] A3: By setting the target flow rate, the valve port flow-pressure differential formula and PID closed-loop control are used to actively control the flow rate of the rodless chamber of hydraulic actuator 1;

[0087] A4: The pressure difference loss across the first proportional pressure reducing valve 332 is mechanically compensated by the first pressure compensating valve 331. The pressure difference across the first proportional pressure reducing valve 332 is set to a constant value, so that the flow rate change and valve opening exhibit a linear relationship.

[0088] A5: Set the rod cavity flow PID control unit to be invalid and the pressure PID control unit output corresponding to the rod cavity to be valid, and perform pressure control on the rod cavity;

[0089] A6: By comparing the set target pressure differential with the actual pressure differential, a pressure closed-loop control is formed to actively control the pressure in the rod chamber, adjust the opening of the second proportional pressure reducing valve 342 port, and change the front and rear pressure differential of the second proportional pressure reducing valve 342 port.

[0090] Under the overload condition of the hydraulic actuator 1, the first proportional pressure reducing valve 332 is used to control the pressure of the oil inlet chamber to keep it at a low pressure and prevent cavitation. The second proportional pressure reducing valve 342 controls the actuator speed. The block diagram of the pressure and flow composite control strategy is shown in the figure. Figure 4 As shown, when in overload condition, F i <0, the control of hydraulic actuator 1 is as follows:

[0091] When the load changes from positive to negative, the rod chamber of the hydraulic actuator 1 switches from pressure control to flow control, and the flow of the rod chamber is actively controlled by using the valve port flow-pressure difference formula and PID closed-loop control; the loss of the pressure difference before and after the second proportional reducing valve 342 is mechanically compensated by the second pressure compensation valve 341, and the pressure difference before and after the second proportional reducing valve 342 is set to a constant value, so that the flow change and the valve port opening show a linear relationship, and the rodless chamber flow PID control unit is set to invalid and the rodless chamber pressure PID control unit output corresponding to the rodless chamber is valid, so as to perform pressure control on the rod chamber.

[0092] In view of the variability of engineering machinery loads in the prior art, a single mode cannot meet the requirements of actual working conditions, and it is necessary to switch modes according to load changes. For working conditions with a load spectrum, the mapping relationship between the cylinder stroke and the load force needs to be saved in the controller, so that the control strategy can be switched in advance according to the load changes. That is, the load mode must be known in advance to switch the impedance and overrun control strategies, which will increase the complexity of control and the limitations of application. In order to solve this problem, the present application adopts a control strategy of load force direction prediction and two-chamber pressure switching to study the feasibility of switching the working mode at the moment of load direction change. That is, the present invention establishes a connection relationship between the load port independent control valve 3, the oil tank 2, the hydraulic actuator 1, and a number of pressure sensors and flow sensors, based on the external load force F i, establish a force balance equation for the hydraulic actuator 1; use the force balance equation of the hydraulic actuator 1 and the logic judgment function to identify the positive and negative loads, determine the working mode, and corresponding to different working modes, establish the flow equation of the rod chamber outflowing the hydraulic actuator 1 according to the valve port pressure flow equation, and obtain the displacement of the first two-position three-way main valve core 333 and the second two-position three-way main valve core 343 under the corresponding working mode through the force balance equation of the hydraulic actuator 1 and the valve port pressure flow equation; control the displacement of the first two-position three-way main valve core 333 and the second two-position three-way main valve core 343 through the control module to control the inlet and outlet flow of the hydraulic controller. The present invention adopts the control strategy of load force direction prediction and two-cavity pressure switching to study the feasibility of switching the working mode at the moment of load direction change, thereby solving the problems of complex control and application limitations in the existing technology. The present application ensures higher control accuracy and higher dynamic response, and promotes the development of hydraulic transmission technology towards high efficiency, precision and intelligence.

[0093] Specifically, if Figures 7 to 9 As shown, when the load Fi>0, that is, the load is positive, the proportional pressure reducing valve flow PID control unit connected to the rodless cavity is regulated to control the flow of its rodless cavity; and the corresponding pressure PID control unit output is invalid, and the control strategy is the impedance working mode at this time; by setting the target flow, the valve port flow-pressure difference formula and PID closed-loop control are used to realize active control of the flow of the rodless cavity; in the flow control process, the loss of the pressure difference before and after the first proportional pressure reducing valve 332 is mechanically compensated by the first pressure compensation valve 331 to ensure that the pressure difference before and after the first proportional pressure reducing valve 332 is a constant value, so that the flow change and the valve port opening present a linear relationship; at the same time, the rod cavity flow PID controller is invalid and the pressure PID output is valid, that is, the rod cavity is pressure controlled; by setting the target pressure difference, a pressure closed-loop control is formed by comparing the actual pressure difference, so as to realize active control of the pressure of the rod cavity, and regulate the opening change of the second proportional pressure reducing valve 342 to change the pressure difference of the proportional pressure reducing valve port;

[0094] When the load Fi is less than 0, that is, the load is negative, the flow control chamber and the pressure control chamber are switched; when the load changes from positive to negative, the pressure in the rodless chamber will drop sharply at the switching moment. According to the valve port flow-pressure differential formula, the valve port opening of the rodless chamber proportional pressure reducing valve will become smaller, making the oil replenishment speed less than the volume expansion speed of the rodless chamber, which will cause cavitation; when the load force direction changes, the logic function controls the rodless chamber to switch to the pressure control mode. To prevent the cavitation phenomenon, the valve port opening of the first proportional pressure reducing valve 332 of the rodless chamber is adjusted to increase, so as to replenish oil to the rodless chamber, reduce the pressure difference before and after the valve, increase the pressure of the rodless chamber, and compensate for the sudden drop caused by the reverse load force; the rod chamber switches to flow control, and the valve port flow-pressure differential formula and PID closed-loop control are used to realize active control of the flow of the rod chamber; during the flow control process, the pressure difference loss before and after the second proportional pressure reducing valve 342 is mechanically compensated by the pressure compensation valve to ensure that the pressure difference before and after the second proportional pressure reducing valve 342 is a constant value, so that the flow change and the valve port opening show a linear relationship.

[0095] In this embodiment, the force balance equation of the hydraulic actuator 1 and the logic judgment function are used to identify the positive and negative loads, and the working mode is determined specifically including:

[0096] Using the logic function H(F i )、G(F i ), outputs 0 or 1 value according to the load direction to control whether the corresponding flow control PID and pressure control PID are valid; when valid, the corresponding flow and pressure are actively controlled;

[0097] The logic function is:

[0098]

[0099] The force balance equation of hydraulic actuator 1 is: p1A1-p2A2=F l ;

[0100] Where p1 represents the rodless cavity pressure of the hydraulic actuator 1, A1 represents the rodless cavity area of ​​the actuator, p2 represents the rod cavity pressure of the hydraulic actuator 1, A2 represents the rod cavity area of ​​the actuator, and F i Indicates the load force of the hydraulic actuator 1.

[0101] Specifically, in this embodiment, when F i >0, indicating that the load is positive, at this time H(F i )=1 and G(F i )=0; This means that the rodless chamber performs flow control and the rod chamber performs pressure control. In this embodiment, when F i <0, indicating that the load is negative. At this time, H(F i )=0 and G(F i)=1; This means that the rodless chamber performs pressure control and the rod chamber performs flow control.

[0102] Specifically, the flow equation of the rod chamber outflowing from the hydraulic actuator 1 is:

[0103] Where, Q2 represents the rod cavity flow, c d represents the flow coefficient of the hydraulic system, ω represents the area gradient, x2 represents the throttle valve core displacement, and ρ represents the fluid density.

[0104] By establishing the force balance equation of the hydraulic actuator 1 and the valve port pressure and flow equation, the formula for the throttle valve core displacement x2 is obtained as follows:

[0105]

[0106] Where, Q2 represents the rod cavity flow, c d represents the hydraulic system flow coefficient, ω represents the area gradient, p1 represents the rodless cavity pressure of hydraulic actuator 1, A1 represents the rodless cavity area of ​​the actuator, p2 represents the rod cavity pressure of hydraulic actuator 1, A2 represents the rod cavity area of ​​the actuator, F i Indicates the load force of the hydraulic actuator 1.

[0107] When the speed V of the hydraulic actuator 1 is not fixed, the target pressure difference is always calculated by the target pressure. When the pressure control is performed in the rodless chamber, a minimum pressure difference value Δp is set. 1min ; When the rod cavity is used for pressure control, a minimum pressure difference value Δp is set 2min ;

[0108] Δp 1min The calculation formula is:

[0109]

[0110] The calculation formula of Δp2min is:

[0111]

[0112] Where p s Indicates the oil supply pressure of the system, that is, the input pressure of the main pressure reducing valve 32; p0 indicates the return oil pressure of the system or the pressure of the oil tank 2; p 1min Indicates the minimum working pressure of the rodless cavity; p 2min Indicates the minimum working pressure of the rod cavity. s Indicates the highest pressure in the hydraulic system, which provides power to the hydraulic actuator 1 such as a hydraulic cylinder or a hydraulic motor; in a load-port independent control system, p sis the input pressure of the main pressure reducing valve 32, which is used to drive the hydraulic actuator 1 after being reduced in pressure; p0 represents the system return oil pressure or the pressure of the oil tank 2; this is the lowest pressure in the hydraulic system, usually close to atmospheric pressure; in the hydraulic system, the return oil pressure is the pressure in the return oil chamber of the hydraulic actuator 1, which determines the pressure level of the hydraulic oil when it returns to the oil tank 2; p0 can also be a value slightly higher than atmospheric pressure to prevent cavitation in the return oil pipeline.

[0113] The detailed explanation is that the load force is calculated based on the force balance equation of the hydraulic actuator 1, and the direction is determined. The flow control chamber and the pressure control chamber are switched instantly when the load direction changes by logical judgment. There is no need to obtain the load force in real time. When the speed and steering are not fixed, the target pressure difference is always calculated by the target pressure. When the pressure control is performed in the rodless chamber, a minimum pressure difference value is set to ensure that the pressure of the rodless chamber is not too low, thereby avoiding cavitation or other adverse effects caused by insufficient pressure; the minimum target pressure difference of the rodless chamber is set to Δp 1min When the speed and direction are not fixed, the target pressure difference is always calculated by the target pressure. When the rod cavity is pressure controlled, a minimum pressure difference value is set to ensure that the pressure of the rod cavity is not too low, thereby avoiding cavitation or other adverse effects caused by insufficient pressure; the minimum target pressure difference of the rod cavity is set to Δp 2min When F i >0, which is the impedance working condition. At this time, the rodless cavity is flow controlled and the rod cavity is pressure controlled. Because the area of ​​the rod cavity is a fixed value, there must be throttling loss at the oil return valve port. Keeping p2 at a small value can reduce throttling loss. Setting p 2min =1.5MPa, the corresponding p1 will also take a smaller value, thereby improving energy saving.

[0114] When F i <0, which is the overload condition. The rodless chamber pressure p1 is obtained from the force balance equation of the hydraulic actuator 1:

[0115]

[0116] Where, p1 represents the rodless cavity pressure of hydraulic actuator 1, A1 represents the rodless cavity area of ​​hydraulic actuator, p2 represents the rod cavity pressure of hydraulic actuator 1, A2 represents the rod cavity area of ​​hydraulic actuator, F i Indicates the load force of the hydraulic actuator 1.

[0117] Under overload conditions, the rodless cavity is pressure controlled and the rod cavity is flow controlled. i +A2p2>0, at this time, there will be no cavitation in the rodless cavity of the oil rod, and the pressure of p1 is determined by the load F i and pressure p2; when F i+A2p2<0, at this time, cavitation will occur in the rodless cavity of the oil rod. In order to maintain p1>0MPa and avoid cavitation, set p 1min =1.5MPa, the corresponding p2 will also take a smaller value, thus achieving energy saving.

[0118] When v<0, the hydraulic actuator 1 is in the overload condition. Under the overload condition, the rodless chamber of the hydraulic actuator 1 performs pressure control and the rod chamber performs flow control. At this time, the target pressure difference Δp 1min is set to a smaller value to avoid cavitation in the rodless cavity and maintain the energy efficiency of the system; at the same time, the flow control of the rod cavity ensures that the speed of the hydraulic actuator 1 is properly adjusted; when v ≥ 0, the hydraulic actuator 1 is in the impedance working condition; under the impedance working condition, the rodless cavity of the hydraulic actuator 1 performs flow control and the rod cavity performs pressure control; at this time, the target pressure difference Δp 2min is set to a smaller value to reduce throttling loss and improve energy saving; at the same time, the flow control of the rodless chamber ensures that the speed of the hydraulic actuator 1 is properly adjusted.

[0119] In this example, when the hydraulic actuator 1 is in the impedance state (v ≥ 0), the pressure p2 in the rod chamber will be controlled at p 2min When the hydraulic actuator 1 is in the overload condition (v<0), the pressure p1 of the rodless chamber will be controlled at p 1min Compared with the traditional four-mode switching, the load force direction prediction control strategy of the present application can reduce the number of mode switching, maintain the continuity of the operation of the hydraulic actuator 1 to the greatest extent, ensure high control accuracy and high dynamic response, and promote the development of hydraulic transmission technology towards high efficiency, precision and intelligence.

[0120] It should be understood that the specific embodiments described above are only used to explain the present invention and are not intended to limit the present invention. Obvious changes or modifications derived from the spirit of the present invention are still within the scope of protection of the present invention.

Claims

1. A control system for a load port independent control valve based on pilot hydraulic control, characterized in that: include: Hydraulic actuators; An oil tank and a load port independent control valve; the oil tank is in fluid communication with the load port independent control valve via a hydraulic pump, and the output port of the load port independent control valve is in fluid communication with the hydraulic actuator; A plurality of flow sensors and pressure sensors are distributed inside the load port independent control valve; a control module for controlling the internal on-off of the load port independent control valve according to a control strategy; the control module includes a rod chamber flow PID control unit, a rod chamber pressure PID control unit, a rodless chamber flow PID control unit, and a rodless chamber pressure PID control unit of the hydraulic actuator; The load port independent control valve includes a valve body, a main pressure-reducing valve installed on the valve body, a first pressure-compensating pressure-reducing module, a second pressure-compensating pressure-reducing module, and a shuttle valve arranged on the output oil circuits of the first pressure-compensating pressure-reducing module and the second pressure-compensating pressure-reducing module; the output oil ports of the first pressure-compensating pressure-reducing module and the second pressure-compensating pressure-reducing module are both connected to the oil inlet and oil outlet oil circuits of the hydraulic actuator; the oil outlet of the shuttle valve is connected to the Ls end of the hydraulic pump.

2. A control system for a load port independent control valve based on pilot hydraulic control according to claim 1, characterized in that: The valve body includes a main oil inlet circuit P, a main oil return circuit T, an Ls port for working pressure feedback, and a first working oil port A and a second working oil port B; The oil inlet P of the main pressure reducing valve Z , the pressure compensation oil inlet P of the first pressure compensation and pressure reducing module Y1 , the pressure compensation oil inlet P of the second pressure compensation pressure reducing module Y2 The oil return port T of the main pressure reducing valve is connected to the main oil inlet oil circuit P. Z , the oil return port T of the first pressure compensation and pressure reducing module Y1 , the oil return port T of the second pressure compensation and pressure reducing module Y2 Both are connected to the main oil return line T liquid line; The first working oil port A of the main pressure reducing valve Z Respectively connected to the pressure reducing oil inlet P of the first pressure compensating pressure reducing module J1 and the pressure reducing oil inlet P of the second pressure compensating pressure reducing module J2 The first working oil port A of the first pressure compensation and pressure reducing module Y1 The first working oil port A of the valve body is connected to the oil circuit, and the first working oil port A of the second pressure compensation and pressure reducing module is connected to the oil circuit. Y2 The second working oil port B of the valve body is connected to the oil circuit; the oil outlet of the shuttle valve is connected to the Ls port liquid circuit of the working pressure feedback.

3. The control system of the load port independent control valve based on pilot hydraulic control according to claim 2, characterized in that: The first pressure-compensating and reducing pressure module includes a first pressure-compensating valve, a first proportional reducing valve and a first two-position three-way main valve core; the oil inlet of the first pressure-compensating valve is fluidly connected to the main oil inlet oil circuit of the valve body, and the oil outlet of the first pressure-compensating valve is fluidly connected to the oil inlet of the first two-position three-way main valve core; the oil inlet of the first proportional reducing valve is fluidly connected to the oil outlet of the main reducing valve, and the oil outlet of the first proportional reducing valve is fluidly connected to the PP end of the first two-position three-way main valve core; the oil return port of the first proportional reducing valve and the oil return port of the first two-position three-way main valve core are both fluidly connected to the main return oil circuit T of the valve body.

4. The control system of the load port independent control valve based on pilot hydraulic control according to claim 2, characterized in that: The second pressure-compensating and reducing pressure module includes a second pressure-compensating valve, a second proportional reducing valve and a second two-position three-way main valve core; the oil inlet of the second pressure-compensating valve is connected to the main oil inlet oil circuit of the valve body, and the oil outlet of the second pressure-compensating valve is connected to the oil inlet liquid circuit of the second two-position three-way main valve core; the oil inlet of the second proportional reducing valve is connected to the oil outlet liquid circuit of the main reducing valve, and the oil outlet of the second proportional reducing valve is connected to the PP end liquid circuit of the second two-position three-way main valve core; the oil return port of the second proportional reducing valve and the oil return port of the second two-position three-way main valve core are both connected to the main oil return oil circuit T of the valve body.

5. A control method for a control system of a load port independent control valve based on pilot hydraulic control according to any one of claims 2 to 4, characterized in that: The following steps are involved: Establish connections between the load port independent control valve, the oil tank, the hydraulic actuator, and several pressure sensors and flow sensors; the first working oil port A of the load port independent control valve is in communication with the rodless chamber fluid path of the hydraulic actuator, and the oil inlet port B of the load port independent control valve is in communication with the rod chamber fluid path of the hydraulic actuator; Based on the external load force F applied to the hydraulic actuator i , establish the force balance equation of the hydraulic actuator; use the force balance equation of the hydraulic actuator and the logic judgment function to identify the positive and negative loads and determine the working mode; Corresponding to different working modes, a flow equation for the flow out of the rod chamber of the hydraulic actuator is established based on the valve port pressure flow equation; The displacements of the first and second two-position three-way main valve cores in the corresponding working mode are obtained through the hydraulic actuator force balance equation and the valve port pressure and flow equation. The control module controls the displacement of the first two-position three-way main valve core and the second two-position three-way main valve core to control the inlet and outlet flow of the hydraulic controller.

6. The control method of a control system for a load port independent control valve based on pilot hydraulic control according to claim 5, characterized in that: The working mode includes an impedance working mode and a load exceeding working mode; In the impedance working mode, F i >0, the control of the hydraulic actuator is as follows: Regulate the rodless cavity flow PID control unit to control the flow of the rodless cavity of the hydraulic actuator; Set the pressure PID control unit output corresponding to the rodless cavity of the hydraulic actuator to be invalid; By setting the target flow, using the valve port flow-pressure differential formula and PID closed-loop control, the flow of the rodless chamber of the hydraulic actuator is actively controlled; The loss of the pressure difference before and after the first proportional pressure reducing valve is mechanically compensated by the first pressure compensating valve, and the pressure difference before and after the first proportional pressure reducing valve is set to a constant value, so that the flow change and the valve opening present a linear relationship; Set the rod cavity flow PID control unit to be invalid and the pressure PID control unit output corresponding to the rod cavity to be valid, and perform pressure control on the rod cavity; By comparing the set target pressure difference with the actual pressure difference, a pressure closed-loop control is formed to actively control the pressure in the rod cavity, adjust the opening change of the second proportional pressure reducing valve port, and change the front and rear pressure difference of the second proportional pressure reducing valve port.

7. The control method of a control system for a load port independent control valve based on pilot hydraulic control according to claim 5, characterized in that: Under the overrunning load condition, F i <0, the control of the hydraulic actuator is as follows: When the load changes from positive to negative, the rod chamber of the hydraulic actuator switches from pressure control to flow control, and the flow of the rod chamber is actively controlled by using the valve port flow-pressure difference formula and PID closed-loop control; The loss of pressure difference before and after the second proportional pressure reducing valve is mechanically compensated by the second pressure compensating valve. The pressure difference before and after the second proportional pressure reducing valve is set to a fixed value, so that the flow change and the valve opening present a linear relationship. Set the rodless cavity flow PID control unit to be invalid and the rodless cavity pressure PID control unit output corresponding to the rodless cavity to be valid, and perform pressure control on the rod cavity.

8. The control method of a control system of a load port independent control valve based on pilot hydraulic control according to claim 5, characterized in that: The method of using the hydraulic actuator force balance equation and the logic judgment function to identify the positive and negative loads and determine the working mode specifically includes: Using the logic function H(F i )、G(F i ), outputs 0 or 1 value according to the load direction to control whether the corresponding flow control PID and pressure control PID are valid; when valid, the corresponding flow and pressure are actively controlled; The logic function is: The force balance equation of the hydraulic actuator is: p1A1-p2A2=F l ; Among them, p1 represents the rodless cavity pressure of the hydraulic actuator, A1 represents the rodless cavity area of ​​the actuator, p2 represents the rod cavity pressure of the hydraulic actuator, A2 represents the rod cavity area of ​​the actuator, F i Indicates the load force of the hydraulic actuator.

9. The control method of a control system of a load port independent control valve based on pilot hydraulic control according to claim 8, characterized in that: The flow equation out of the rod chamber of the hydraulic actuator is: Where, Q2 represents the rod cavity flow, c d represents the hydraulic system flow coefficient, ω represents the area gradient, x2 represents the throttle valve core displacement, and ρ represents the fluid density; Then the throttle valve core displacement x2 formula is: Where, Q2 represents the rod cavity flow, c d Indicates the flow coefficient of the hydraulic system, ω indicates the area gradient, p1 indicates the rodless cavity pressure of the hydraulic actuator, A1 indicates the rodless cavity area of ​​the actuator, p2 indicates the rod cavity pressure of the hydraulic actuator, A2 indicates the rod cavity area of ​​the actuator, F i Indicates the load force of the hydraulic actuator.

10. A control method for a control system of a load port independent control valve based on pilot hydraulic control according to any one of claims 5 to 9, characterized in that: When the speed V of the hydraulic actuator is not fixed, the target pressure difference is always calculated by the target pressure. When the pressure is controlled in the rodless chamber, a minimum pressure difference value Δp is set. 1min ; When the rod cavity is used for pressure control, a minimum pressure difference value Δp is set 2min ; The Δp 1min The calculation formula is: The Δp 2min The calculation formula is: Where p s Indicates the oil supply pressure of the system, that is, the input pressure of the main pressure reducing valve; p0 indicates the return oil pressure or tank pressure of the system; p 1min Indicates the minimum working pressure of the rodless cavity; p 2min Indicates the minimum working pressure of the rod chamber.

Images