Hydraulic loading system based on heterogeneous valve group cooperative driving and control method

Abstract

The invention relates to the technical field of hydraulic power control and engineering testing, in particular to a hydraulic loading system based on heterogeneous valve group cooperative driving and a control method, and the hydraulic loading system comprises a double-acting pressure cylinder, an execution oil cylinder, an oil source, an oil return tank, a mixed valve control driving module, an unloading valve and a controller; a large-area driving cavity and a small-area high-pressure output cavity are formed in the double-acting pressure cylinder, and a piston of the double-acting pressure cylinder is movably arranged in the large-area driving cavity and the small-area high-pressure output cavity; an oil inlet of the mixing valve control driving module is connected with an oil source, an oil return opening of the mixing valve control driving module is connected with an oil return box, and meanwhile a working oil opening of the mixing valve control driving module is connected with a large-area driving cavity of the double-acting pressure cylinder. The unloading valve is arranged between the oil inlet of the execution oil cylinder and the oil return tank; and the controller is electrically connected with the unloading valve of the mixing valve control driving module and is used for controlling the working state. The device has the advantage of being suitable for heavy-load and high-dynamic test scenes such as large-scale wind power transmission chain loading and the like.

Description

Technical Field

[0001] This technical solution relates to the field of hydraulic power control and engineering testing technology, specifically a hydraulic loading system and control method based on the coordinated drive of heterogeneous valve groups. Background Technology

[0002] As the capacity of individual wind turbine units continues to increase, the structural dimensions and load-bearing capacity of their drivetrains are also growing, leading to a trend towards larger and heavier-load-bearing ground-based drivetrain test benches. To realistically reproduce the actual operating conditions of the turbines on the test bench, various coupled loads, such as torque, axial force, radial force, and bending moment, need to be applied to the main shaft, gearbox, and related connecting components. Currently, the non-torque loads on large wind turbine units are rapidly increasing, significantly raising the peak load and required flow rate of a single loading channel, making traditional hydraulic loading devices insufficient to meet testing requirements.

[0003] Existing ground test benches mostly use hydraulic cylinders as the loading actuators, with the output force determined by the system pressure and the effective piston area. To improve loading capacity, the rated system pressure is usually increased or the cylinder diameter is enlarged. However, when the test bench enters MW (megawatt) level and above heavy-duty conditions, the above methods have significant limitations: 1. Increased system stress leads to a significant increase in costs and risks; 2. Increasing the cylinder diameter leads to a dramatic increase in cavity volume, which reduces the system response speed and makes it impossible to reproduce the rapidly changing loads of the wind field; 3. Large-diameter servo valves have low frequency response and large dead zone, making it difficult to accurately reconstruct high-frequency dynamic loads.

[0004] Therefore, hydraulic loading systems face a technical contradiction between balancing high-flow-rate supply and high-frequency precision control. A single servo valve solution cannot simultaneously meet the engineering requirements of high response, high precision, and low cost. At the same time, the wind power industry has higher requirements for the ability to reproduce test conditions, needing to achieve high-frequency dynamic loading under high load and high flow conditions, making traditional solutions unsuitable.

[0005] Therefore, there is an urgent need for a hydraulic loading system that can provide efficient and reliable technical support for large-scale wind power transmission chain test benches. Summary of the Invention

[0006] The purpose of this technical solution is to provide a hydraulic loading system and control method based on heterogeneous valve group cooperative drive. By using a hybrid valve control structure and the area grading characteristics of the double-acting booster cylinder, a high-pressure, high-flow pump source is constructed to cooperatively drive the end-acting cylinder, thus solving the problem that existing hydraulic loading systems cannot be adapted to large-scale wind power transmission chain test benches.

[0007] The purpose of this technical solution is achieved as follows: A hydraulic loading system based on heterogeneous valve group cooperative drive includes: a double-acting booster cylinder, an actuator cylinder, an oil source, a return oil tank, a hybrid valve-controlled drive module, an unloading valve, and a controller; The double-acting booster cylinder has a large-area driving chamber and a small-area high-pressure output chamber. The piston of the double-acting booster cylinder moves within the large-area driving chamber and the small-area high-pressure output chamber. The oil inlet of the hybrid valve-controlled drive module is connected to the oil source, and the oil return port is connected to the oil return tank. At the same time, the working oil port of the hybrid valve-controlled drive module is connected to the large-area driving chamber of the double-acting booster cylinder. The unloading valve is located between the oil inlet of the actuator cylinder and the oil return tank. The controller is electrically connected to the unloading valve of the hybrid valve-controlled drive module and is used to control the working state.

[0008] Preferably, the hybrid valve-controlled drive module includes a servo valve and a switching valve; the servo valve and the switching valve adopt parallel oil circuits, the oil inlets of the servo valve and the switching valve are both connected to an oil source, the oil return ports of the servo valve and the switching valve are both connected to an oil return tank, and the working oil ports of the two are combined and connected to the drive chamber of the double-acting booster cylinder.

[0009] Preferably, the switching valve is a large-diameter two-position four-way solenoid valve or a three-position four-way solenoid valve, the number of switching valves is several, and the switching valves are arranged in parallel to linearly expand the maximum flow loading capacity of the system. The servo valve is a small-diameter, high-frequency servo valve used for fine adjustment and dynamic compensation of the double-acting booster cylinder.

[0010] Preferably, the unloading valve is a normally closed two-position two-way solenoid valve; during loading, the unloading valve remains closed to maintain pipeline pressure; during unloading, the unloading valve opens, connecting the circuit between the actuator cylinder and the return oil tank, eliminating pipeline back pressure to allow the return spring in the actuator cylinder to push the piston back.

[0011] Preferably, the piston of the double-acting booster cylinder is positioned within the large-area driving chamber and the small-area high-voltage output chamber, thereby dividing the large-area driving chamber into a first chamber and a second chamber, and dividing the small-area high-voltage output chamber into a third chamber and a fourth chamber. Preferably, a piston is provided inside the large-area driving cavity, which divides the large-area driving cavity into a first cavity and a second cavity. A first plunger and a second plunger are respectively provided on both sides of the piston, and the first plunger and the second plunger extend into the small-area high-voltage output cavity. The small-area high-voltage output cavity corresponding to the first plunger is the third cavity, and the small-area high-voltage output cavity corresponding to the second plunger is the fourth cavity. The double-acting booster cylinder also includes a hydraulic rectifier bridge, which includes: a first check valve installed between the oil outlet of the first chamber and the actuator cylinder; a second check valve installed between the oil outlet of the third chamber and the actuator cylinder; a third check valve installed between the oil outlet of the fourth chamber and the actuator cylinder; and a fourth check valve installed between the oil outlet of the second chamber and the actuator cylinder. Each check valve works together to achieve unidirectional high-pressure oil output when the piston of the double-acting booster cylinder reciprocates. The oil outlet of the third chamber is connected to a first high-pressure oil outlet channel and a first oil replenishment channel; a first check valve is provided on the path of the first oil replenishment channel, which allows oil to enter the third chamber; a second check valve is provided on the path of the first high-pressure oil outlet channel, which allows oil to leave the third chamber and enter the actuator cylinder. The oil outlet of the fourth chamber is connected to a second high-pressure oil outlet channel and a second oil replenishment channel; a third check valve is provided on the path of the second high-pressure oil outlet channel, which allows oil to leave the fourth chamber and enter the actuator cylinder; a fourth check valve is provided on the path of the second oil replenishment channel, which allows oil to enter the fourth chamber. A first oil driving channel is provided on the side of the first cavity adjacent to the third cavity. Both the first oil driving channel and the first oil replenishment channel are connected to the first output port of the servo valve and the third output port of the switching valve. A second oil drive channel is provided on the side of the second cavity adjacent to the fourth cavity. Both the second oil drive channel and the second oil replenishment channel are connected to the second output port of the servo valve and the fourth output port of the switching valve.

[0012] A control method for a hydraulic loading system based on heterogeneous valve group cooperative drive, applied to the hydraulic loading system based on heterogeneous valve group cooperative drive as described above, includes the following steps: Step S1: The controller divides the working mode into precision adjustment mode and high flow rate fast-acting mode according to the flow requirements of the target loading spectrum; Step S2: When in precision adjustment mode, the controller controls the unloading valve to close and the switching valve to close, only driving the servo valve to work; the servo valve outputs a small flow to drive the double-acting booster cylinder micro-motion, which, after area-level amplification, controls the actuator cylinder to achieve high-precision loading; Step S3: When in high-flow-rate high-speed mode, the controller controls the unloading valve to close and the switching valve to open; the switching valve quickly fills the double-acting booster cylinder with oil to establish the base load, and at the same time the servo valve works in coordination to dynamically compensate for the pressure fluctuations caused by the intervention of the switching valve, and controls the actuator to achieve high-flow-rate stable loading. Step S4: When entering the unloading return mode, the controller closes the servo valve and the switching valve, and controls the unloading valve to open; the oil in the actuator cylinder is discharged back to the oil tank through the unloading valve under the action of the return spring, realizing the system reset.

[0013] Preferably, in step S3, the double-acting booster cylinder operates as a hydraulic transformer, utilizing the area ratio of the large-area drive chamber to the small-area output chamber to convert the low-pressure, high-flow input from the switching valve into the high-pressure oil flow required to drive the actuator cylinder; the hydraulic rectifier bridge plays a rectifying role in this process, preventing the high-pressure oil from flowing back into the booster cylinder.

[0014] Preferably, in step S4, the opening of the unloading valve provides a bypass pressure relief channel, which overcomes the reverse shut-off characteristic of the one-way valve group inside the double-acting booster cylinder, enabling the actuator to overcome pipeline resistance and complete the return stroke by relying on its own spring force.

[0015] The key and beneficial technical effects of this technical solution compared to existing technologies are: This technical solution utilizes a hybrid valve-controlled drive module to regulate the hydraulic fluid, enabling the system to achieve high-flow loading without relying on large-diameter servo valves, thus reducing system costs. Simultaneously, the hybrid valve-controlled drive module also features fine-tuning capabilities, significantly improving the system's dynamic performance and scalability, making the system suitable for heavy-load, high-dynamic testing scenarios such as large-scale wind turbine drivetrain loading. Attached Figure Description

[0016] Figure 1 This is a schematic diagram of the hydraulic loading system of the present invention.

[0017] Figure 2 This is a flowchart of the hydraulic loading process of the present invention.

[0018] Figure 3 This is a schematic diagram of the wind turbine transmission chain test bench of the present invention.

[0019] Reference numerals in the attached diagram: 1. Double-acting booster cylinder; 10. Actuating cylinder; 11. First chamber; 12. Second chamber; 13. Third chamber; 14. Fourth chamber; 15. First check valve; 16. Second check valve; 17. Third check valve; 18. Fourth check valve; 19. Unloading valve; 2. Oil source; 3. Return oil tank; 4. Controller; 5. Servo valve; 6. Switch valve; 7. First pressure sensor; 8. Second pressure sensor; 9. Displacement sensor. Detailed Implementation

[0020] The specific implementation of this technical solution will be further described in detail below with reference to the accompanying drawings.

[0021] like Figures 1-3As shown, a hydraulic loading system based on heterogeneous valve group synergistic drive mainly comprises: a double-acting booster cylinder 1, an actuator cylinder 10, a servo valve 5, a switching valve 6, an oil source 2, a return oil tank 3, a pressure sensor 7 and a pressure sensor 8, a displacement sensor 9, an unloading valve 19, and a controller 4. In terms of hardware connections, the inlets of both the servo valve 5 and the switching valve 6 are connected to the same oil source 2, and their return ports are both connected to the return oil tank 3. Their working ports are connected in parallel and converged, then jointly connected to the large-area drive chamber (i.e., the first chamber 11 and the second chamber 12) of the double-acting booster cylinder 1. This design constitutes a heterogeneous valve group common-source parallel drive structure, ensuring that the high-frequency fine adjustment capability of the servo valve and the large flow capacity of the switching valve can synergistically act on the same power source.

[0022] The double-acting booster cylinder 1 serves as the pre-power conversion unit of this system. Its internal design employs a graded area structure, comprising a large-area drive chamber and a small-area high-pressure output chamber (i.e., the third chamber 13 and the fourth chamber 14). This structure forms a hydraulic booster pump system, enabling small pressure changes at the input end to be converted into high-pressure oil flow at the output end. To coordinate with the reciprocating motion of the double-acting booster cylinder 1 and achieve unidirectional output, the internal oil circuit of the booster cylinder integrates a hydraulic rectifier bridge consisting of a first check valve 15 installed between the oil outlet of the first chamber and the actuator cylinder, a second check valve 16 installed between the oil outlet of the third chamber and the actuator cylinder, a third check valve 17 installed between the oil outlet of the fourth chamber and the actuator cylinder, and a fourth check valve 18 installed between the oil outlet of the second chamber and the actuator cylinder.

[0023] Regardless of whether the piston of the double-acting booster cylinder 1 moves to the left or right, the squeezed oil will flow through the one-way valve 15 or 16 to the high-pressure output pipeline and be supplied unidirectionally to the end actuator cylinder 10.

[0024] The hydraulic cylinder 10, as the final mechanical actuator of the system, adopts a single-acting spring-reset structure. Its rodless chamber inlet receives high-pressure oil from the double-acting booster cylinder 1, which, under hydraulic pressure, pushes the piston rod out to apply a load. To achieve unloading and reset of the system, an unloading valve 19 is installed between the oil inlet of the hydraulic cylinder 10 and the return oil tank 3. It should be noted that, in order to enable the attached... Figure 1 The diagram is simple and clear. The electrical control circuit between the controller 4 and the unloading valve 19 is not specifically shown in the diagram. However, in this embodiment, the unloading valve 19 is a normally closed two-position two-way solenoid valve. Its electromagnet is electrically connected to the controller 4 and is controlled to open or close by the controller 4.

[0025] The aforementioned pressure sensor 7 is installed in the large-area drive chamber oil circuit of the double-acting booster cylinder (i.e., the drive chamber oil circuit of the first chamber 11 and the second chamber 12 of the double-acting booster cylinder after the servo valve 5 and the switching valve 6 are connected in parallel and merged), directly acquiring the real-time pressure signal of the drive chamber. Its main function is to monitor the pressure in the drive chamber of the double-acting booster cylinder, providing input pressure feedback for the coordinated control of the servo valve 5 and the switching valve 6; assisting in judging the working status of the booster cylinder, avoiding overload or abnormal fluctuations in the drive chamber pressure, and ensuring system safety.

[0026] Pressure sensor 8 is installed in the oil inlet circuit of the actuator cylinder 10 (i.e., the high-pressure oil circuit of the rodless chamber of the actuator cylinder after rectification by the first one-way valve 15 and the second one-way valve 16 from the high-pressure output chamber of the double-acting booster cylinder), directly acquiring the pressure signal of the actuator cylinder's oil inlet circuit. Its main function is to directly reflect the pressure in the actuator cylinder's oil inlet circuit, calculating the real-time loading force using the formula F=p×A (F is the loading force, p is the oil inlet pressure, and A is the effective piston area). As the core feedback signal for force closed-loop control, it is used to calibrate loading accuracy, compensate for system errors, and ensure that the loading force is highly consistent with the target spectrum.

[0027] Displacement sensor 9 is installed between the piston rod and the cylinder body (or at the end of the piston rod) of actuator cylinder 10 to collect displacement signals of the piston rod, reflecting the stroke and movement state of actuator cylinder. Its main function is to monitor the displacement of the actuator cylinder piston rod, reflecting the loading stroke and movement speed; it works with pressure sensor 8 to achieve force-displacement hybrid control, preventing overtravel during high-flow-rate rapid loading and ensuring displacement resolution during precision adjustment; it is also used for system status diagnosis, such as determining whether the actuator cylinder is stuck or whether the reset is in place.

[0028] The signal output terminals of all three are electrically connected to controller 4, forming a feedback link for multi-parameter closed-loop control.

[0029] A piston is provided inside the large-area driving cavity, which divides the large-area driving cavity into a first cavity 11 and a second cavity 12. A first plunger and a second plunger are respectively provided on both sides of the piston, and the first plunger and the second plunger extend into the small-area high-voltage output cavity. The small-area high-voltage output cavity corresponding to the first plunger is the third cavity 13, and the small-area high-voltage output cavity corresponding to the second plunger is the fourth cavity 14. The double-acting booster cylinder 1 also includes a hydraulic rectifier bridge, which includes: a first check valve 15 installed between the oil outlet of the first chamber 11 and the actuator cylinder 10; a second check valve 16 installed between the oil outlet of the third chamber 13 and the actuator cylinder 10; a third check valve 17 installed between the oil outlet of the fourth chamber 14 and the actuator cylinder 10; and a fourth check valve 18 installed between the oil outlet of the second chamber 12 and the actuator cylinder 10. Each check valve works together to achieve unidirectional high-pressure oil output when the piston of the double-acting booster cylinder 1 reciprocates. The oil outlet of the third chamber 13 is connected to a first high-pressure oil outlet channel and a first oil replenishment channel; a first one-way valve 15 is provided on the path of the first oil replenishment channel, which allows oil to enter the third chamber 13; a second one-way valve 16 is provided on the path of the first high-pressure oil outlet channel, which allows oil to leave the third chamber 13 and enter the actuator cylinder 10. The oil outlet of the fourth chamber 14 is connected to a second high-pressure oil outlet channel and a second oil replenishment channel; the third check valve 17 is provided on the path of the second high-pressure oil outlet channel, which allows oil to leave the fourth chamber and enter the actuator cylinder 10; the fourth check valve 18 is provided on the path of the second oil replenishment channel, which allows oil to enter the fourth chamber 14. A first oil driving channel A3 is provided on the side of the first cavity 11 adjacent to the third cavity 13. The first oil driving channel and the first oil replenishment channel are both connected to the first output port A1 of the servo valve 5 and the third output port A2 of the switching valve 6. A second oil drive channel B3 is provided on the side of the second cavity 12 adjacent to the fourth cavity 14. Both the second oil drive channel and the second oil replenishment channel are connected to the second output port B1 of the servo valve 5 and the fourth output port B2 of the switching valve 6.

[0030] The system's workflow and control strategy are as follows: Controller 4 dynamically schedules the status of each valve group based on the changing trend of the target load spectrum. When the system is in a low flow demand or precision adjustment phase, controller 4 keeps the switching valve 6 closed, only driving the servo valve 5. At this time, the small flow of hydraulic oil output by the servo valve 5 drives the double-acting booster cylinder 1 to micro-motion. After area-level amplification, it is converted into a high-resolution displacement or force output of the actuator cylinder 10. This mode utilizes the high-frequency response characteristics of the servo valve to achieve accurate reproduction of minute loads.

[0031] When the system needs to quickly establish a base load or is in a high-flow-rate rapid-motion phase, controller 4 controls the switching valve 6 to open rapidly. Utilizing its large-diameter characteristic, this valve forcefully fills the drive chamber of the double-acting booster cylinder 1 with oil, driving the piston of the double-acting booster cylinder 1 to move rapidly and pumping a large flow of oil into the actuator cylinder 10. Simultaneously, servo valve 5 works in coordination, providing real-time dynamic compensation for the flow rate jump or pressure fluctuation caused by the intervention of switching valve 6, ensuring that the loading curve output by actuator cylinder 10 is smooth and shock-free. This collaborative mechanism perfectly resolves the contradiction between high-flow-rate loading and high-precision control.

[0032] Specifically, hydraulic oil is supplied from oil source 2 into switching valve 6. The hydraulic oil flows out from the fourth output port B2 of switching valve 6. Part of it flows into the fourth chamber 14 after passing through the fourth check valve 18, and the other part flows into the second chamber 12 through the second oil drive channel B3. Due to the large-flow instantaneous inflow, it can push the piston to move quickly. The hydraulic oil in the first chamber 11 is discharged from the first oil drive channel A3 and enters the actuator cylinder 10 through the first check valve 15 and the second check valve 16. The oil in the third chamber 13 also enters the actuator cylinder 10 through the second check valve 16. When the piston moves from right to left to the set position, switching valve 6 switches the oil circuit, and the oil is output from the third output port A2. Part of it flows into the fourth chamber 14 after passing through the first check valve 15. In the third chamber 13, another portion flows into the first chamber 11 through the first oil drive channel A3. Due to the large flow rate, it continues to drive the piston to move rapidly. The hydraulic oil in the second chamber 12 is discharged from the second oil drive channel B3 and enters the actuator cylinder 10 through the fourth check valve 18 and the third check valve 17. The oil in the fourth chamber 14 also enters the actuator cylinder 10 through the third check valve 17. The servo valve 5 performs dynamic compensation. When the fourth output port B2 of the switching valve 6 supplies oil, the first output port A1 of the servo valve 5 opens to supply oil, smoothing the piston movement. After the piston moves smoothly, the oil supply stops. When the piston moves in the opposite direction, the second output port B1 of the servo valve 5 opens to supply oil, and after the piston moves smoothly, the oil supply stops.

[0033] When fine-tuning is required, hydraulic oil is supplied from oil source 2 into servo valve 5. The hydraulic oil flows out from the first output port A1 of servo valve 5. Part of it flows into the third chamber 13 after passing through the first check valve 15, and the other part flows into the first chamber 11 from the first hydraulic drive channel A3, pushing the piston to move to the right. Under the piston area amplification effect, the hydraulic oil in the fourth chamber 14 is squeezed out and flows into the actuator cylinder 10 after passing through the third check valve 17. At the same time, the oil in the second chamber 12 flows out from the second hydraulic drive channel B3 with the piston movement, and flows into the actuator cylinder 10 after passing through the fourth check valve 18 and the third check valve 17 in sequence, realizing high-precision feeding with small flow rate.

[0034] During the rapid execution phase (when the working pressure of the actuator cylinder 10 is relatively low), the switching valve 6 opens to provide a large flow drive: when the fourth output port B2 of the switching valve 6 supplies oil, the large flow of oil enters the second chamber 12 and the fourth chamber 14 in two separate paths, pushing the piston to move rapidly to the left. The oil in the first chamber 11 and the third chamber 13 flows into the actuator cylinder 10 through the corresponding check valve groups. The servo valve 5 simultaneously and briefly opens the first output port A1 to smooth the piston start-up impact, and closes after the piston movement stabilizes. When the switching valve 6 switches to supply oil to the third output port A2, the large flow of oil enters the first chamber 11 and the third chamber 13, pushing the piston to move rapidly to the right. The oil in the second chamber 12 and the fourth chamber 14 flows into the actuator cylinder 10 through the corresponding check valve groups. The servo valve 5 simultaneously and briefly opens the second output port B1 to achieve smooth reversal.

[0035] During the high-pressure execution phase, the oil flowing out of the second hydraulic drive channel B3 can only flow into the actuator cylinder 10 through the fourth check valve 18 and the third check valve 17, or supply oil to the second chamber 12 with the piston movement. It cannot return oil directly to the oil tank T through the second output port B1 or the fourth output port B2. The system oil return is only achieved through the return port T1 of the servo valve 5 or the return port T2 of the switching valve 6. At the same time, the booster continuously uses the piston area difference to provide high-pressure boosting output to the actuator cylinder 10.

[0036] When the test ends or a return stroke is required, the system enters the unloading process. Controller 4 issues a command to stop the oil supply action of servo valve 5 and switching valve 6, and simultaneously controls the unloading valve 19 to open. At this time, the oil inlet chamber of actuator cylinder 10 is connected to the return oil tank 3. The reset spring inside actuator cylinder 10 releases its elastic potential energy, pushing the piston rod to retract. The oil in the cylinder is quickly discharged back to the return oil tank 3 through the opened unloading valve 19, realizing the safe reset and zeroing of the system.

[0037] Furthermore, the 6-branch switching valve in this system supports a modular expansion design. Depending on the tonnage requirements of the test bench, multiple switching valves can be connected in parallel in the oil circuit. The controller 4 determines the number of switching valves to open based on real-time flow requirements, thereby linearly expanding the system's maximum loading speed. This design eliminates the need for the servo valve 5 to handle the main flow, allowing for the selection of a small-diameter, high-performance servo valve with a smaller dead zone and higher bandwidth, further enhancing the dynamic performance and control stability of the entire loading system.

[0038] The specific steps are as follows: Step S1: The controller divides the working mode into precision adjustment mode and high flow rate fast-acting mode according to the flow requirements of the target loading spectrum; Step S2: When in precision adjustment mode, the controller controls the unloading valve to close and the switching valve to close, only driving the servo valve to work; the servo valve outputs a small flow to drive the double-acting booster cylinder micro-motion, which, after area-level amplification, controls the actuator cylinder to achieve high-precision loading; Step S3: When in high-flow-rate high-speed mode, the controller controls the unloading valve to close and the switching valve to open; the switching valve quickly fills the double-acting booster cylinder with oil to establish the base load, and at the same time the servo valve works in coordination to dynamically compensate for the pressure fluctuations caused by the intervention of the switching valve, and controls the actuator to achieve high-flow-rate stable loading. Step S4: When entering the unloading return mode, the controller closes the servo valve and the switching valve, and controls the unloading valve to open; the oil in the actuator cylinder is discharged back to the oil tank through the unloading valve under the action of the return spring, realizing the system reset.

[0039] In step S3, the double-acting booster cylinder operates as a hydraulic transformer, utilizing the area ratio of the large-area drive chamber to the small-area output chamber to convert the low-pressure, high-flow input from the switching valve into the high-pressure oil flow required to drive the actuator cylinder; the hydraulic rectifier bridge plays a rectifying role in this process, preventing the high-pressure oil from flowing back into the booster cylinder.

[0040] In step S4, the opening of the unloading valve provides a bypass pressure relief channel, which overcomes the reverse shut-off characteristic of the one-way valve group inside the double-acting booster cylinder, enabling the actuator to overcome pipeline resistance and complete the return stroke by relying on its own spring force.

[0041] The foregoing has shown and described the basic principles, main features, and advantages of this technical solution. Those skilled in the art should understand that this technical solution is not limited to the above embodiments. The embodiments and descriptions in the specification are merely illustrative of the principles of this technical solution. Various changes and modifications can be made to this technical solution without departing from its spirit and scope, and all such changes and modifications fall within the scope of the claimed technical solution. The scope of protection of this technical solution is defined by the appended claims and their equivalents.

[0042] It should be noted that the structures, proportions, and sizes depicted in the accompanying drawings are solely for illustrative purposes and to aid those skilled in the art in understanding and reading the content disclosed herein. They are not intended to limit the implementation of this technical solution and therefore have no substantial technical significance. Any modifications to the structure, changes in proportions, or adjustments to size, provided they do not affect the effectiveness or purpose of this technical solution, should still fall within the scope of the technical content disclosed herein. Furthermore, the terms used in this specification, such as "upper," "lower," "left," "right," "middle," and "one," are merely for clarity and not intended to limit the scope of implementation of this technical solution. Changes or adjustments to their relative relationships, without substantially altering the technical content, should also be considered within the scope of the technical solution's implementation.

[0043] It should also be noted that when a component is referred to as being "fixed to" or "set on" another component, it can be directly on the other component or may be connected to an intermediary component. When a component is referred to as being "connected to" another component, it can be directly connected to the other component or indirectly connected to the other component through an intermediary component.

[0044] Furthermore, the use of terms such as "first" and "second" in this application is for descriptive purposes only and should not be construed as indicating or implying their relative importance or implicitly specifying the number of technical features indicated. Therefore, a feature defined as "first" or "second" may explicitly or implicitly include at least one of that feature. Additionally, the technical solutions of the various embodiments can be combined with each other, but only on the basis of being achievable by those skilled in the art. When the combination of technical solutions is contradictory or impossible to implement, such a combination of technical solutions should be considered non-existent and not within the scope of protection claimed in this application.

Claims

1. A hydraulic loading system based on the coordinated drive of heterogeneous valve groups, characterized in that, include: The system comprises a double-acting booster cylinder (1), an actuator cylinder (10), an oil source (2), a return oil tank (3), a hybrid valve-controlled drive module, an unloading valve (19), and a controller (4). The double-acting booster cylinder (1) has a large-area drive chamber and a small-area high-pressure output chamber inside. The piston of the double-acting booster cylinder (1) moves within the large-area drive chamber and the small-area high-pressure output chamber. The oil inlet of the hybrid valve-controlled drive module is connected to the oil source (2), and the oil return port is connected to the return oil tank (3). At the same time, the working oil port of the hybrid valve-controlled drive module is connected to the large-area drive chamber of the double-acting booster cylinder (1). The unloading valve (19) is located between the oil inlet of the actuator cylinder (10) and the return oil tank (3). The controller (4) is electrically connected to the hybrid valve-controlled drive module and the unloading valve (19) and is used to control the working state.

2. The hydraulic loading system based on heterogeneous valve group cooperative drive according to claim 1, characterized in that: The hybrid valve-controlled drive module has: a servo valve (5) and a switching valve (6); the servo valve (5) and the switching valve (6) adopt parallel oil circuits, the oil inlets of the servo valve (5) and the switching valve (6) are both connected to the oil source (2), the oil return ports of the servo valve (5) and the switching valve (6) are both connected to the oil return tank (3), and the working oil ports of the two are connected together to the drive chamber of the double-acting booster cylinder (1) after merging.

3. The hydraulic loading system based on heterogeneous valve group cooperative drive according to claim 1, characterized in that: The switching valve (6) is a large-diameter two-position four-way solenoid valve. There are several switching valves (6), and the switching valves (6) are connected in parallel to linearly expand the maximum flow loading capacity of the system. The servo valve (5) is a small-diameter, high-frequency servo valve used for fine adjustment and dynamic compensation of the double-acting booster cylinder (1).

4. The hydraulic loading system based on heterogeneous valve group cooperative drive according to claim 1, characterized in that: The unloading valve (19) is a normally closed two-position two-way solenoid valve. During the loading process, the unloading valve (19) remains closed to maintain the pipeline pressure. During the unloading process, the unloading valve (19) opens to connect the circuit between the actuator cylinder (10) and the return oil tank (3), eliminating the pipeline back pressure so that the return spring in the actuator cylinder (10) can push the piston back.

5. The hydraulic loading system based on heterogeneous valve group cooperative drive according to claim 1, characterized in that: A piston is provided inside the large-area driving cavity, which divides the large-area driving cavity into a first cavity (11) and a second cavity (12). A first plunger and a second plunger are respectively provided on both sides of the piston, and the first plunger and the second plunger extend into the small-area high-voltage output cavity. The small-area high-voltage output cavity corresponding to the first plunger is the third cavity (13), and the small-area high-voltage output cavity corresponding to the second plunger is the fourth cavity (14). The double-acting booster cylinder (1) also includes a hydraulic rectifier bridge, which includes: a first check valve (15) installed between the oil outlet of the first chamber (11) and the actuator cylinder (10), a second check valve (16) installed between the oil outlet of the third chamber (13) and the actuator cylinder (10), a third check valve (17) installed between the oil outlet of the fourth chamber (14) and the actuator cylinder (10), and a fourth check valve (18) installed between the oil outlet of the second chamber (12) and the actuator cylinder (10). Each check valve cooperates to realize unidirectional high-pressure oil output when the piston of the double-acting booster cylinder (1) reciprocates. The oil outlet of the third chamber (13) is connected to a first high-pressure oil outlet channel and a first oil replenishment channel; a first one-way valve (15) is provided on the path of the first oil replenishment channel, which allows oil to enter the third chamber (13); a second one-way valve (16) is provided on the path of the first high-pressure oil outlet channel, which allows oil to leave the third chamber (13) and enter the actuator cylinder (10); The oil outlet of the fourth chamber (14) is connected to a second high-pressure oil outlet channel and a second oil replenishment channel; the third check valve (17) is provided on the path of the second high-pressure oil outlet channel, and the third check valve (17) allows the oil to leave the fourth chamber and enter the actuator cylinder (10); the fourth check valve (18) is provided on the path of the second oil replenishment channel, and the fourth check valve (18) allows the oil to enter the fourth chamber (14); A first oil driving channel is provided on the side of the first cavity (11) adjacent to the third cavity (13). The first oil driving channel and the first oil replenishment channel are both connected to the first output port of the servo valve (5) and the third output port of the switching valve (6). A second oil driving channel is provided on the side of the second cavity (12) adjacent to the fourth cavity (14). The second oil driving channel and the second oil replenishment channel are both connected to the second output port of the servo valve (5) and the fourth output port of the switching valve (6).

6. A control method for a hydraulic loading system based on heterogeneous valve group cooperative drive, applied to the hydraulic loading system based on heterogeneous valve group cooperative drive as described in any one of claims 1-5, characterized in that, Includes the following steps: Step S1: The controller (4) divides the working mode into precision adjustment mode and high flow rate fast-acting mode according to the flow requirements of the target loading spectrum; Step S2: When in precision adjustment mode, the controller (4) controls the unloading valve (19) to close and the switching valve (6) to close, and only drives the servo valve (5) to work; the servo valve (5) outputs a small flow to drive the double-acting booster cylinder (1) to move slightly, and after being amplified by area gradation, it controls the execution cylinder (10) to achieve high-precision loading; Step S3: When in high flow rate fast-acting mode, the controller (4) controls the unloading valve (19) to close and the switching valve (6) to open; the switching valve (6) quickly fills the double-acting booster cylinder (1) with oil to establish the base load, and at the same time the servo valve (5) works in coordination to dynamically compensate for the pressure fluctuations caused by the intervention of the switching valve (6) and control the actuator cylinder (10) to achieve high flow rate stable loading; Step S4: When entering the unloading return mode, the controller (4) closes the servo valve (5) and the switching valve (6), and controls the unloading valve (19) to open; the oil in the execution cylinder (10) is discharged back to the oil tank (3) through the unloading valve (19) under the action of the return spring, thereby realizing the system reset.

7. The control method for a hydraulic loading system based on heterogeneous valve group cooperative drive according to claim 6, characterized in that: In step S3, the double-acting booster cylinder (1) operates as a hydraulic transformer, using the area ratio of the large-area drive chamber to the small-area output chamber to convert the low-pressure, high-flow input from the switching valve (6) into the high-pressure oil flow required to drive the actuator cylinder (10); the hydraulic rectifier bridge plays a rectifying role in this process to prevent the high-pressure oil from flowing back into the booster cylinder.

8. The control method for a hydraulic loading system based on heterogeneous valve group cooperative drive according to claim 6, characterized in that: In step S4, the opening of the unloading valve (19) provides a bypass pressure relief channel, which overcomes the reverse shut-off characteristic of the one-way valve group inside the double-acting booster cylinder (1), so that the actuator cylinder (10) can overcome the pipeline resistance and complete the return action by relying on its own spring force.

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