Control method for oil supply hydraulic servo system of energy accumulator of shipborne helicopter mooring device

Abstract

The invention discloses a control method of an energy accumulator oil supply hydraulic servo system of a shipborne helicopter mooring device, which belongs to the field of mechanical hydraulic control and comprises the following steps: establishing a mathematical model of the energy accumulator oil supply hydraulic servo system; based on the position of a helicopter probe rod, the speed track in the capturing process of a mechanical claw is planned; a smooth speed feedback signal is obtained through a tracking differentiator; generating a feed-forward control quantity by a feed-forward controller; designing an error active disturbance rejection controller comprising a feedback compensator and an extended state observer based on the mathematical model, and generating an error adjustment signal; and the feedforward control quantity and the error adjusting signal are synthesized, the parameters of the error active disturbance rejection controller are adjusted, a final control instruction is formed to drive a hydraulic servo system, and the mechanical claw tracks the planned speed track. The problem of flow-pressure nonlinearity in the oil supply process of the energy accumulator is effectively solved, the capturing speed tracking precision of the mechanical claw is improved, capturing impact is reduced, and the quick mooring device is suitable for helicopters of different tonnages.

Description

Technical Field

[0001] This invention relates to the field of mechanical hydraulic control technology, and in particular to a control method for a hydraulic servo system for the oil supply of an accumulator in a shipborne helicopter mooring device. Background Technology

[0002] When the helicopter lands on the deck, the quick-tether device needs to grab the probe on the helicopter. The grabbing device is driven by a hydraulic system powered by an accumulator. Due to measurement errors, speed control should be used to reduce impact. However, the accumulator pressure will continuously decrease with the continuous supply of oil, exacerbating the flow-pressure nonlinearity. In addition, parameter uncertainties, unknown nonlinear friction, and leakage will all affect the dynamic performance and speed tracking accuracy of the system.

[0003] In electro-hydraulic servo system research, actuator speed control has a crucial impact on the system's dynamic performance. Existing research mainly focuses on position control in constant-pressure oil supply systems, while research on speed control in valve-controlled hydraulic servo systems is relatively weak. Unlike position control, speed control requires maintaining a specific opening of the servo valve in steady state, which places special demands on control system design. Existing solutions mainly employ feedforward control combined with PID-like algorithms, but these have significant limitations in handling system nonlinear characteristics and variable oil supply pressure disturbances. Although Active Disturbance Rejection Control (ADRC) and its improved Error Active Disturbance Rejection Control (EADRC) have demonstrated excellent disturbance suppression capabilities in other fields, research on their application in hydraulic servo control is scarce.

[0004] Therefore, it is necessary to design a control method for the hydraulic servo system of the accumulator oil supply of the shipborne helicopter mooring device to solve the above-mentioned technical problems. Summary of the Invention

[0005] The technical problem to be solved by the present invention is to provide a control method for the hydraulic servo system of the accumulator oil supply of a shipborne helicopter mooring device. Based on motion speed trajectory planning, the method obtains the feedforward control quantity through feedforward gain and designs an extended state controller; it ensures closed-loop control of the entire speed tracking process, enabling the system to cope with different initial oil supply pressures and has stronger robustness.

[0006] To solve the above-mentioned technical problems, the technical solution adopted by the present invention is as follows:

[0007] A control method for a hydraulic servo system for the accumulator oil supply of a shipborne helicopter mooring device includes the following steps:

[0008] S1. Establish a mathematical model of the accumulator oil supply hydraulic servo system;

[0009] S2. Based on the position of the helicopter probe, plan the velocity trajectory during the capture process of the mechanical claw;

[0010] S3. Input the displacement feedback signal collected by the displacement sensor into the tracking differentiator to obtain a smooth velocity feedback signal;

[0011] S4. Compare the speed feedback signal with the speed trajectory planned in S2 to obtain the speed error signal, and generate the feedforward control quantity through the feedforward controller;

[0012] S5. Based on the mathematical model, design an error active disturbance rejection controller including a feedback compensator and an extended state observer. The extended state observer estimates and compensates for the combined disturbances caused by the accumulator pressure decay and system uncertainty in real time, and generates an error adjustment signal.

[0013] S6. Synthesize the feedforward control quantity and the error adjustment signal, adjust the parameters of the error active disturbance rejection controller, and form the final control command to drive the hydraulic servo system so that the mechanical gripper tracks the planned speed trajectory.

[0014] A further improvement to the technical solution of this invention lies in the following: In S1, the mathematical model is used to describe the nonlinear relationship between accumulator pressure change, servo valve flow rate, and hydraulic cylinder dynamics, including three parts: hydraulic cylinder force balance equation, proportional servo valve flow rate equation, and hydraulic cylinder flow rate continuity equation. The mathematical model expression is as follows:

[0015] Define the state variable as The nonlinear state equation of the system is obtained as follows:

[0016]

[0017] In the formula, These refer to the displacement and velocity of the hydraulic cylinder piston, respectively. For load pressure; The equivalent mass of the load and the piston rod of the coarse and fine hydraulic cylinders; These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod-side chamber and rodless chamber of the coarse hydraulic cylinder, respectively. These refer to the pressures inside the rod-side chamber and rodless chamber of the fine hydraulic cylinder, respectively. These refer to the pressures within the rod-side and rodless chambers of the coarse hydraulic cylinder, respectively. It is a nonlinear function that includes nonlinear frictional force, unmodeled dynamics of the system, and unknown external disturbance forces; Oil supply pressure for the accumulator; coefficient They are respectively:

[0018]

[0019] In the formula, The effective bulk modulus of hydraulic oil; ; For servo valve flow gain; ; , These refer to the volumes of the rod-side chamber and the rodless chamber of the fine hydraulic cylinder, respectively. , , and These are the initial volumes of the left and right chambers of the hydraulic cylinder, respectively. This represents the leakage coefficient within the hydraulic cylinder.

[0020] A further improvement to the technical solution of the present invention is that the force balance equation of the hydraulic cylinder is as follows:

[0021]

[0022] In the formula, The equivalent mass of the load and the piston rod of the coarse and fine hydraulic cylinders; These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod-side chamber and rodless chamber of the coarse hydraulic cylinder, respectively. These refer to the pressures within the rod-side and rodless chambers of the coarse hydraulic cylinder, respectively. These refer to the displacement and velocity of the hydraulic cylinder piston, respectively. It is a nonlinear function that includes nonlinear frictional force, unmodeled dynamics of the system, and unknown external disturbance forces; is the viscous damping coefficient.

[0023] A further improvement to the technical solution of the present invention is that the nonlinear flow equation of the servo valve is as follows:

[0024]

[0025] In the formula, For servo valve flow gain; ; Oil supply pressure for the accumulator; flow rate equation , The flow rate into the rod chamber of the fine hydraulic cylinder; The flow rate exiting the rodless chamber of the fine hydraulic cylinder; For lumped control signals, , For error adjustment signal, This is the feedforward control quantity; For load pressure, ,coefficient .

[0026] A further improvement to the technical solution of the present invention is that the hydraulic cylinder load flow equation is as follows:

[0027]

[0028] In the formula, The flow rate into the rod chamber of the fine hydraulic cylinder; The flow rate exiting the rodless chamber of the fine hydraulic cylinder; This refers to the leakage coefficient within the hydraulic cylinder. The effective bulk modulus of hydraulic oil; , These refer to the volumes of the rod-side chamber and the rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively.

[0029] The volume descriptions of the rod-side chamber and rodless chamber of the hydraulic cylinder are as follows:

[0030]

[0031] In the formula, and These are the initial volumes of the rod-side chamber and the rodless chamber of the hydraulic cylinder, respectively.

[0032] A further improvement of the technical solution of the present invention is that, in S2, the speed trajectory planning during the capture process of the mechanical claw based on the position of the helicopter probe specifically includes: during the capture process of the helicopter probe by the mechanical claw, the lateral movement speed of the mechanical claw is planned according to the position of the probe, so as to ensure that the tethering device can complete the capture task within the specified time range while minimizing the speed at which a capture collision occurs.

[0033] The trajectory planning problem is described as follows: For a fixed capture range, a robotic gripper starts from an initial velocity of 0 and flexibly accelerates to a certain velocity within a certain time. It then runs at that constant speed for a period of time; when the robotic gripper approaches the probe, it enters the deceleration phase and decelerates to its speed over the remaining distance. Simultaneously reaching the end of the interval; to ensure the robotic gripper can reach the target speed When capturing the probe, an error buffer zone should be reserved at the end of the capture range, within which the mechanical gripper should always maintain a certain speed. The mechanical gripper moves at a constant speed until the mechanical gripper status detection sensor of the rapid tethering device successfully detects that the mechanical gripper has closed. Then the hydraulic lock is closed, the position of the mechanical gripper is locked, and the speed of the mechanical gripper is reduced to 0.

[0034] The capture area is divided into 3 capture zones. Based on the helicopter probe position issued by the central control system, the capture zone in which the target is located is determined. Combined with the capture time requirements, the capture speed curve of the mechanical claw is dynamically planned.

[0035]

[0036] In the formula, This represents the desired motion velocity trajectory of the robotic gripper. Indicates acceleration time; Indicates the time of uniform motion; Indicates the deceleration time; This represents the maximum expected uniform velocity. This represents the minimum expected uniform motion speed.

[0037] A further improvement to the technical solution of this invention lies in the following: In S3, specifically, it includes: using a tracking differentiator to obtain the motion velocity from the displacement feedback signal; constructing a second-order discrete tracking differentiator as follows:

[0038]

[0039] In the formula, For displacement feedback Tracking; displacement signal The differential; Sampling time; , These are adjustable parameters for the fhan function, which is:

[0040]

[0041] In the formula, Threshold parameters related to acceleration and filtering time; , All are state variables; The predicted value is the discretized velocity value; This is an intermediate quantity representing the nonlinear switching boundary; The predicted value is after nonlinear correction; It is a switching function; This serves as an intermediate output that integrates linear and nonlinear data. This is for secondary switch judgment.

[0042] A further improvement to the technical solution of the present invention lies in: in S4, specifically, it includes: deriving the capture velocity of the tethering device based on the mathematical model as follows:

[0043]

[0044] in, ; ; ;

[0045] To avoid the influence of differential pressure signals on control, the feedforward controller is designed as follows:

[0046]

[0047] In the formula, The desired speed.

[0048] A further improvement to the technical solution of the present invention is that, in S5, the design process of the error active disturbance rejection controller is as follows:

[0049] (1) Design of feedback compensator, the specific process is as follows:

[0050] According to the mathematical model, the state-space expression for the system velocity is:

[0051]

[0052] In the formula, ; ; ; ;

[0053] ; ; ; Lumped control signal , For error adjustment signal, This is a feedforward control signal;

[0054] Redefining The above equation can be rewritten as the chain integrator structure:

[0055]

[0056] in, ;

[0057] ;

[0058] The error in the movement speed and acceleration of the hydraulic cylinder is defined as: , , and These are the desired velocity and desired acceleration, respectively; the hydraulic cylinder velocity error equation is:

[0059]

[0060] In the formula, This is a comprehensive function of the system state. for The estimated value;

[0061] Add derivative feedback terms to both sides of the above equation. The dynamic equations for the expected error of the closed-loop system are formed and simplified to:

[0062]

[0063] In the formula, For total disturbance;

[0064] Therefore, the feedback compensator is designed as follows:

[0065]

[0066] In the formula, the gain of the active disturbance rejection controller is... ;

[0067] (2) The design of the extended state observer is as follows:

[0068] Define state vector The dynamic equation for the expected error of the closed-loop system is rewritten as:

[0069]

[0070] In the formula, ;

[0071] definition To estimate the state vector of an error system, an error-based LESO is designed as follows:

[0072]

[0073] In the formula, the extended state observer gain vector , .

[0074] A further improvement to the technical solution of this invention lies in the following: In S6, the specific process is as follows: the desired closed-loop dynamics are... The pole configuration method is used to make it equal to the desired Hurwitz polynomial. The gain of the active disturbance rejection controller is obtained as follows:

[0075]

[0076] in, For the error active disturbance rejection controller bandwidth;

[0077] The ESO estimation error state matrix is ​​represented as follows: The characteristic polynomial is At that time, the gain of the extended state observer is calculated as follows:

[0078]

[0079] In the formula, the bandwidth of the extended state observer is... ;

[0080] The parameters that need to be tuned are , , The adjustment steps are as follows:

[0081] (1) Calculate the parameters based on the known parameters. The nominal value is then fine-tuned during the experiment.

[0082] (2) Calculate using the nominal values ​​of the parameters according to the mathematical model, and make fine adjustments during the experiment;

[0083] (3) The larger the size, the stronger the anti-interference ability. An increase in noise will amplify the noise, causing the control signal to vibrate and affecting the stability of the system. Adjustments should be made according to the noise intensity and tracking performance requirements.

[0084] (4) Affects dynamic response performance, bandwidth yes 3-5 times.

[0085] The technological advancements achieved by this invention due to the adoption of the above technical solutions are as follows:

[0086] 1. This invention effectively solves the problem of nonlinear changes in supply flow caused by pressure decay during the oil supply process of an accumulator by constructing an error-driven active disturbance rejection controller (error-driven active disturbance rejection speed controller) and integrating feedforward control. The method directly uses the speed tracking error as the input of the error-driven active disturbance rejection controller, which can observe and dynamically compensate for the composite disturbances caused by pressure changes, model uncertainties and external loads, thereby significantly improving the speed trajectory tracking accuracy of the mechanical gripper during rapid approach. Combined with frequency domain stability analysis, it ensures the robust stability of the closed-loop system under all operating conditions, and overcomes the inherent defects of traditional PID algorithms in dealing with such strong nonlinear disturbances from the perspective of control principle.

[0087] 2. The control method provided by this invention enables the rapid mooring device to adapt to mooring tasks of helicopters with different initial pressure conditions and different tonnages, significantly expanding the applicability of the rapid mooring device; through high-precision speed tracking and end-of-line buffer control, the impact on the helicopter probe is greatly reduced, improving the safety and reliability of the mooring process; at the same time, this control method makes full use of the accumulator fuel supply mode, reducing the instantaneous power requirements of the pump station, achieving the effects of saving energy and reducing system size and cost. Attached Figure Description

[0088] To more clearly illustrate the technical solutions in the embodiments of the present invention or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are some embodiments of the present invention. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0089] Figure 1 This is a flowchart of a control method for a hydraulic servo system for oil supply to an accumulator of a shipborne helicopter mooring device, provided in an embodiment of the present invention.

[0090] Figure 2 This is a schematic diagram of the structure of a hydraulic servo system for supplying oil to an accumulator of a shipborne helicopter mooring device, provided in an embodiment of the present invention.

[0091] Figure 3a This is a schematic diagram of the far-end acquisition interval of the helicopter acquisition speed planning curve in an embodiment of the present invention;

[0092] Figure 3b This is a schematic diagram of the near-end acquisition interval of the helicopter acquisition speed planning curve in an embodiment of the present invention;

[0093] Figure 4 This is a block diagram of the transfer function structure of the closed-loop control system in an embodiment of the present invention;

[0094] Figure 5a Here are the frequency response curves for EADRC and FEADRC;

[0095] Figure 5b Frequency domain analysis curves for EADRC and FEADRC;

[0096] Figure 6a For open-loop control, see speed tracking curves under different pressures;

[0097] Figure 6b The graph shows the speed tracking error curves under different pressures for open-loop control.

[0098] Figure 7a Speed ​​tracking curves under different control methods;

[0099] Figure 7b Speed ​​tracking error curves under different control methods;

[0100] Figure 7c Pressure-displacement curves under different control methods;

[0101] Figure 7d Plots showing the control signal and disturbance estimation curves under different control methods;

[0102] Figure 8This is a schematic diagram illustrating the performance indicators of the three different controllers—open-loop control, FPID, and FEADRC—mentioned in Embodiment 2 of the present invention.

[0103] Figure 9a For FEADRC control, speed tracking curves under different pressures are shown.

[0104] Figure 9b The graph shows the speed tracking error curves of FEADRC control under different pressures;

[0105] Figure 9c This is a pressure-displacement curve under FEADRC control.

[0106] Figure 9d The graph shows the control signal and disturbance estimation curves under FEADRC control.

[0107] Figure 10 A schematic diagram showing the performance indicators of the FEADRC controller under different pressures;

[0108] The components include: 1. Oil tank; 2. Electric motor; 3. Hydraulic pump; 4. Check valve; 5. Filter; 6. Relief valve; 7. Servo valve; 8. Hydraulic cylinder; 8.1. Coarse hydraulic cylinder; 8.2. Fine hydraulic cylinder; 9. Pressure sensor; 10.1. First two-position four-way solenoid directional valve; 10.2. Second two-position four-way solenoid directional valve; 10.3. Third two-position four-way solenoid directional valve; 10.4. Fourth two-position four-way solenoid directional valve; 11. Accumulator; 12. Cartridge valve; 13. Stacked double one-way throttle valve; 14. Stacked internal control hydraulic lock; 15. External control high-flow hydraulic lock; 16.1. First safety valve; 16.2. Second safety valve. Detailed Implementation

[0109] It should be noted that the terms "comprising" and "having" and any variations thereof in the specification, claims and accompanying drawings of this invention are intended to cover non-exclusive inclusion. For example, a process, method, system, product or device that includes a series of steps or units is not necessarily limited to those steps or units that are explicitly listed, but may include other steps or units that are not explicitly listed or that are inherent to such processes, methods, products or devices.

[0110] The present invention will be further described in detail below with reference to the accompanying drawings and embodiments:

[0111] like Figure 2As shown, a hydraulic servo system for supplying oil to an accumulator in a shipborne helicopter mooring device includes: an oil tank 1, an electric motor 2, a hydraulic pump 3, a check valve 4, a filter 5, an overflow valve 6, a servo valve 7 (a Y-type electro-hydraulic proportional servo valve), a hydraulic cylinder 8 (a coarse and fine parallel single-rod hydraulic cylinder), a pressure sensor 9, a two-position four-way solenoid directional valve group (including a first two-position four-way solenoid directional valve 10.1, a second two-position four-way solenoid directional valve 10.2, a third two-position four-way solenoid directional valve 10.3, and a fourth two-position four-way solenoid directional valve 10.4), an accumulator 11, a cartridge valve 12, a stacked double one-way throttle valve 13, a stacked internal control hydraulic lock 14, an external control high-flow hydraulic lock 15, a first safety valve 16.1, and a second safety valve 16.2; it also includes a displacement sensor and an inertial load ( Figure 2 (Not shown in the drawing).

[0112] The hydraulic pump 3 is connected to the load oil circuit through the check valve 4 and the filter 5. The relief valve 6 is connected to the load oil circuit to limit the oil circuit pressure from exceeding the maximum pressure setting value in order to protect the safety of the oil circuit.

[0113] In the hydraulic system, the energy storage oil circuit is composed of cartridge valve 12, first and second position four-way solenoid directional valve 10.1, and accumulator 11. Port A of cartridge valve 12 is connected to the load oil circuit, and port B is connected to accumulator 11 and port P of first and second position four-way solenoid directional valve 10.1. Pressure sensor 9 measures the pressure of accumulator 11.

[0114] The rod hydraulic cylinder 8 includes a coarse hydraulic cylinder 8.1 and a fine hydraulic cylinder 8.2;

[0115] The fine hydraulic cylinder 8.2 is connected to the load oil circuit through the servo valve 7, and the A and B ports of the fine hydraulic cylinder 8.2 are respectively connected to the A and B ports of the servo valve 7;

[0116] Ports A and B of the coarse hydraulic cylinder 8.1 are connected to the oil tank 1 via the oil circuits on the A and B sides of the externally controlled high-flow hydraulic lock 15, respectively. Ports A and B of the coarse hydraulic cylinder 8.1 are also connected to the oil tank 1 via the first safety valve 16.1 and the second safety valve 16.2. The control oil circuits on the A and B sides of the externally controlled high-flow hydraulic lock 15 are uniformly connected to port A of the fourth two-position four-way solenoid directional valve 10.4. Ports P and T of the fourth two-position four-way solenoid directional valve 10.4 are connected to the energy storage oil circuit and the oil tank 1, respectively.

[0117] like Figure 1 As shown, a control method for a hydraulic servo system for the accumulator oil supply of a shipborne helicopter mooring device includes the following steps:

[0118] S1. Establish a mathematical model of the accumulator oil supply hydraulic servo system;

[0119] This mathematical model describes the nonlinear relationship between accumulator pressure changes, servo valve flow rate, and hydraulic cylinder dynamics. It comprises three parts: the hydraulic cylinder force balance equation, the servo valve nonlinear flow equation, and the hydraulic cylinder load flow equation. Figure 2 As shown, the hydraulic servo system for the accumulator oil supply of the shipborne helicopter mooring device mainly consists of a hydraulic cylinder 8, a servo valve 7, a displacement sensor, and an inertial load. The control objective of the hydraulic servo system for the accumulator oil supply of the shipborne helicopter mooring device is to ensure that the output displacement of the hydraulic cylinder accurately tracks the expected trajectory. The mathematical model is constructed as follows:

[0120] S11, Establish the hydraulic cylinder force balance equation

[0121] According to Newton's second law, the force balance equation of a hydraulic cylinder can be expressed as:

[0122]

[0123] In equation (1), The equivalent mass of the load and the hydraulic cylinder piston rod; These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod-side chamber and rodless chamber of the coarse hydraulic cylinder, respectively. These refer to the pressures inside the rod-side chamber and rodless chamber of the fine hydraulic cylinder, respectively. These refer to the pressures within the rod-side and rodless chambers of the coarse hydraulic cylinder, respectively. These refer to the displacement and velocity of the hydraulic cylinder piston, respectively. It is a nonlinear function that includes nonlinear frictional force, unmodeled dynamics of the system, and unknown external disturbance forces;

[0124] S12, Establish the nonlinear flow equation for the servo valve

[0125] Define the rightward movement of the mechanical gripper as the positive direction, and the leftward movement of the hydraulic cylinder piston. According to the volumetric effect, hydraulic oil flows into the rod chamber, i.e., the right chamber of the thin hydraulic cylinder. The nonlinear flow equation of the servo valve can be described as:

[0126]

[0127] In equation (2), The flow coefficient at the servo valve orifice is between 0.6 and 1; The valve orifice area gradient; The density of the hydraulic oil; For the displacement of the servo valve spool; Provide oil pressure for the accumulator; This refers to the system return oil pressure. The flow rate into the rod chamber of the fine hydraulic cylinder; The flow rate exiting the rodless chamber of the fine hydraulic cylinder; For servo valve flow gain, For the direction sign function, the formula is... This can be simplified to:

[0128]

[0129] Define the load flow equation , will Substituting the values, the nonlinear flow equation for the servo valve becomes:

[0130]

[0131] In equation (4), For servo valve flow gain; .

[0132] S13. Establish the hydraulic cylinder load-flow equation

[0133] Without considering external leakage of the hydraulic cylinder and pressure loss in the pipeline, the hydraulic cylinder load flow equation is established based on the valve-controlled asymmetric hydraulic cylinder model as follows:

[0134]

[0135] In equation (5), This refers to the leakage coefficient within the hydraulic cylinder. The effective bulk modulus of hydraulic oil; , These refer to the volumes of the rod-side chamber and the rodless chamber of the fine hydraulic cylinder, respectively.

[0136] The volumes of the rod-side chamber and the rodless chamber of a hydraulic cylinder can be described as follows:

[0137]

[0138] In equation (6), and These are the initial volumes of the rod-side chamber and the rodless chamber of the hydraulic cylinder, respectively.

[0139] Define load pressure ,coefficient Joint ,Mode The flow continuity equation for the hydraulic cylinder can be obtained as follows:

[0140]

[0141] Hydraulic cylinder movement process Combined You can get and The expression is:

[0142]

[0143] right and Differentiation yields and :

[0144]

[0145] The formula Substitution From this, we can obtain:

[0146]

[0147] The formula Substitution In this process, the rate of change of load pressure is obtained as follows:

[0148]

[0149] Joint ,Mode The system force balance equations are simplified to:

[0150]

[0151] Define the state variable as The nonlinear state equation of the accumulator-supply hydraulic servo system is obtained as follows:

[0152]

[0153] Mode middle, These refer to the displacement and velocity of the hydraulic cylinder piston, respectively. For load pressure; The equivalent mass of the load and the piston rod of the coarse and fine hydraulic cylinders; These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod-side chamber and rodless chamber of the coarse hydraulic cylinder, respectively. These refer to the pressures inside the rod-side chamber and rodless chamber of the fine hydraulic cylinder, respectively. These refer to the pressures within the rod-side and rodless chambers of the coarse hydraulic cylinder, respectively. It is a nonlinear function that includes nonlinear frictional force, unmodeled dynamics of the system, and unknown external disturbance forces; Provide oil pressure for the accumulator; This is the viscous damping coefficient; coefficient They are respectively:

[0154]

[0155] Mode middle, The effective bulk modulus of hydraulic oil; ; For servo valve flow gain; ; , These refer to the volumes of the rod-side chamber and the rodless chamber of the fine hydraulic cylinder, respectively. , , and These are the initial volumes of the left and right chambers of the hydraulic cylinder, respectively. This represents the leakage coefficient within the hydraulic cylinder.

[0156] S2. Based on the position of the helicopter probe, plan the velocity trajectory during the capture process of the mechanical claw;

[0157] During the process of the mechanical gripper capturing the helicopter probe, the lateral movement speed of the mechanical gripper needs to be planned according to the position of the probe to ensure that the tethering device can complete the capture task within the specified time range while minimizing the speed at which a capture collision occurs.

[0158] The trajectory planning problem can be described as follows: For a certain fixed capture interval, a robotic gripper starts from an initial velocity of 0 and flexibly accelerates to a certain velocity within a certain time. It then runs at that constant speed for a period of time; when the robotic gripper approaches the probe, it enters the deceleration phase and decelerates to its speed over the remaining distance. Simultaneously reaching the end of the interval; to ensure the robotic gripper can reach the target speed When capturing the probe, an error buffer zone should be reserved at the end of the capture range, within which the mechanical gripper should always maintain a certain speed. The device moves at a constant speed until the mechanical gripper status detection sensor of the rapid tethering device successfully detects that the mechanical gripper has closed. Then, the hydraulic lock is closed, the position of the mechanical gripper is locked, and the speed of the mechanical gripper drops to 0.

[0159] Based on the helicopter probe position, plan the velocity trajectory during the robotic gripper capture process: (e.g.) Figure 3a and Figure 3bAs shown, the capture area is divided into 3 capture zones. Based on the helicopter probe position issued by the central control system, the capture zone in which the helicopter is located is determined. Combined with the capture time requirements, the capture speed curve of the mechanical claw is dynamically planned.

[0160]

[0161] In equation (15), This represents the desired motion velocity trajectory of the robotic gripper. Indicates acceleration time; Indicates the time of uniform motion; Indicates the deceleration time; This represents the maximum expected uniform velocity. This represents the minimum expected uniform motion speed.

[0162] S3. Input the displacement feedback signal collected by the displacement sensor into the tracking differentiator to obtain a smooth velocity feedback signal;

[0163] The construction process of the tracking differentiator is as follows:

[0164] The accumulator-supply hydraulic servo system uses speed trajectory tracking as its control objective. Since displacement sensors cannot directly measure the speed of the stationary device, a tracking differentiator is used to obtain the speed from the displacement feedback signal.

[0165] The second-order discrete tracking differentiator is constructed as follows:

[0166]

[0167] Mode middle, displacement signal Tracking; displacement signal The differential; Sampling time; , These are adjustable parameters for the fhan function, which is:

[0168]

[0169] In equation (17), Threshold parameters related to acceleration and filtering time; , For state variables; The predicted value is the discretized velocity value; This is an intermediate quantity representing the nonlinear switching boundary; The predicted value is after nonlinear correction; It is a switching function; This serves as an intermediate output that integrates linear and nonlinear data. This is for secondary switch judgment.

[0170] S4. Compare the speed feedback signal with the speed trajectory planned in S2 to obtain the speed error signal, and generate the feedforward control quantity through the feedforward controller;

[0171] The feedforward controller is designed as follows:

[0172] If we assume that the speed tracking error approaches zero over time, the controller output signal will be zero, leading to the servo valve closing, zero system flow, and ultimately forcing the hydraulic cylinder to stop. This result fundamentally contradicts the control objective of the accumulator-supplyed hydraulic servo system to achieve speed tracking. To solve this technical problem, speed tracking control of the accumulator-supplyed hydraulic servo system requires a feedforward controller that corresponds to the expected speed trajectory to address the speed tracking error.

[0173] According to S1 Chinese style - The capture velocity of the tethered device can be derived as follows:

[0174]

[0175] In equation (18), ; ; ;

[0176] To avoid the influence of differential pressure signals on control, the feedforward controller is designed as follows:

[0177]

[0178] Mode middle, The desired speed.

[0179] S5. An error active disturbance rejection controller, including a feedback compensator and an extended state observer, is designed based on a mathematical model. The extended state observer estimates and compensates for the combined disturbances caused by the accumulator pressure decay and system uncertainty in real time, and generates an error adjustment signal.

[0180] The design process of the error active disturbance rejection controller is as follows:

[0181] (1) Design of feedback compensator, the specific process is as follows:

[0182] According to the formula It can be seen that the velocity state-space expression of the accumulator-supply hydraulic servo system is:

[0183]

[0184] Mode middle, ; ; ; ;

[0185] ; ; ; Lumped control signal , For error adjustment signal, This is the feedforward control signal.

[0186] Redefining , will The rewritten chain integrator structure is as follows:

[0187]

[0188] In equation (21) ;

[0189] .

[0190] The error in the movement speed and acceleration of the hydraulic cylinder is defined as: , , and These are the desired velocity and desired acceleration, respectively. The hydraulic cylinder velocity error equation is:

[0191]

[0192] Mode middle, This is a comprehensive function of the system state. for The estimated value.

[0193] The formula Add derivative feedback terms to both sides simultaneously This constitutes the dynamic equation of the expected error of the closed-loop system, equation... It can be simplified to:

[0194]

[0195] Mode middle This represents the total disturbance.

[0196] Therefore, the feedback compensator is designed as follows:

[0197]

[0198] Mode In the middle, the gain of the automatic disturbance rejection controller. .

[0199] (2) The design of the extended state observer is as follows:

[0200] Define state vector ,Mode It can be rewritten as:

[0201]

[0202] Mode middle,

[0203] definition To estimate the state vector of an error system, an error-based LESO is designed as follows:

[0204]

[0205] Mode In the extended state observer gain vector , .

[0206] (3) Stability analysis of closed-loop system

[0207] like Figure 4 The figure shown is a block diagram of the transfer function of a closed-loop control system. For reference speed signal; For tracking error; For EADRC control output; The feedforward controller controls the output; This is the main control input; External interference; For system output; It is the transfer function of the feedforward controller; It is the transfer function of EADRC.

[0208] Pair Performing a Laplace transform, we obtain the transfer function of EADRC as follows:

[0209]

[0210] Pair Simplifying, we get:

[0211]

[0212] Joint ,Mode have to:

[0213]

[0214] Mode middle, ; ; ; This is the natural frequency of hydraulic pressure. The hydraulic damping ratio; This refers to the total volume of the hydraulic cylinder cavity; Flow-pressure coefficient;

[0215] The transfer function of the feedforward controller can be approximated by a proportional coefficient:

[0216]

[0217] Therefore, the transfer function of the closed-loop control system is:

[0218]

[0219] The closed-loop stability characteristic equation is:

[0220]

[0221] Mode middle, ; ; ; ; ;

[0222] Due to the formula Medium parameters All are greater than zero, therefore According to the Routh stability criterion, equation... The necessary and sufficient condition for stability is:

[0223]

[0224] Mode middle, It is a Hurwitz determinant.

[0225] S6. Synthesize the feedforward control quantity and error adjustment signal, adjust the parameters of the error active disturbance rejection controller, and form the final control command to drive the hydraulic servo system so that the mechanical gripper tracks the planned speed trajectory.

[0226] The specific process is as follows:

[0227] According to the formula ,Mode Therefore, the desired closed-loop dynamics are: The pole configuration method is used to make it equal to the desired Hurwitz polynomial. The gain of the active disturbance rejection controller is obtained as follows:

[0228]

[0229] Mode middle, The bandwidth is the error active disturbance rejection controller.

[0230] According to the formula Japanese style The ESO estimation error state matrix can be represented as follows: The characteristic polynomial is At that time, the gain of the extended state observer can be calculated as:

[0231]

[0232] Mode In the extended state observer bandwidth .

[0233] Bird diagrams of FEADRC and EADRC Figure 5a As shown. By Figure 5a It can be seen that the feedforward controller plays a crucial role in the control loop, reducing the sensitivity of FEADRC to control parameters. For the same set of control parameters, the closed-loop bandwidth of FEADRC is 762 times larger than that of EADRC. The observer's burden is significantly reduced. Assuming the input is given, the transfer function form from the output to the disturbance can be obtained. The corresponding frequency response is as follows. Figure 5b As shown, increasing the bandwidth of both the observer and the controller can enhance the anti-interference performance of FEADRC.

[0234] The parameters that need to be tuned are , , The adjustment steps are as follows:

[0235] (1) Calculate the parameters based on the known parameters. The nominal value is then fine-tuned during the experiment.

[0236] (2) According to the formula The parameters are calculated using their nominal values ​​and fine-tuned during the experiment.

[0237] (3) The larger the size, the stronger the anti-interference ability. An increase in noise will amplify the noise, causing control signal vibration and affecting system stability. Adjustments should be made based on noise intensity and tracking performance requirements.

[0238] (4) The main impact is on dynamic response performance; based on engineering experience, bandwidth... Usually 3-5 times.

[0239] The beneficial effects of the present invention are verified using the following embodiments:

[0240] against Figure 2 The hydraulic servo system for supplying oil to the accumulator of the shipborne helicopter mooring device shown is compared with open-loop control and feedforward PID controller (FPID) using the integrated feedforward controller and error-driven active disturbance rejection controller (FEADRC) constructed in this invention. The experimental verification uses the following parameters: , , , , , , , , .

[0241] Example 1

[0242] This embodiment compares the open-loop control performance using only a feedforward controller under different initial oil supply pressures, and the results are as follows: Figure 6a and 6b As shown. From Figure 7a , 7b As shown in Figures 7c and 7d, the speed tracking curves and tracking error curves indicate that the higher the initial oil supply pressure, the greater the speed overshoot. Conversely, when the initial oil supply pressure is low, the maximum speed is difficult to reach the desired value. In addition, as the oil supply pressure decreases, the speed of the hydraulic cylinder gradually decreases, and the accumulator pressure change curve shows a decreasing trend, resulting in a gradual decrease in the pressure change rate during the movement. The change trend of the hydraulic cylinder displacement curve verifies this. The experimental results show that the open-loop control has poor robustness, and a single feedforward controller cannot meet the requirements of speed tracking accuracy.

[0243] Example 2

[0244] This embodiment compares the control performance of three control methods under the same initial oil supply pressure. Figure 7a , 7b As shown in Figures 7c and 7d, both FEADRC and FPID can improve speed tracking accuracy and mitigate speed drop caused by a decrease in fuel supply pressure. From Figure 7c As shown in the accumulator voltage drop curves, FEADRC exhibits the slowest decreasing trend, indicating that it stores the most energy. Furthermore, closed-loop control offers higher positioning accuracy compared to open-loop control. Figure 7d The control signal curves for three controllers are given. It can be seen that the control signal oscillation is relatively large, which is related to the oscillation of the speed signal. The online estimation of the total disturbance is as follows: Figure 7d As shown. Based on the speed tracking error, three control performance indicators were calculated, such as... Figure 8 As shown, the maximum tracking error, average error, and standard deviation of error, from largest to smallest, are open-loop control, FPID, and FEADRC, further illustrating the superior performance of FEADRC control.

[0245] Example 3

[0246] This embodiment compares the control performance of FEADRC under the same controller but different initial oil supply pressures to verify its robustness. Figure 9a Speed ​​tracking curve and Figure 9b The speed tracking error curve shows that the hydraulic cylinder maintains high tracking accuracy under different initial oil supply pressures, and is almost unaffected by changes in the initial oil supply pressure of the accumulator. The accumulator pressure change curve is shown in the curve. Figure 9c As shown. From Figure 9d The control output curve shows that the amplitude of the control quantity increases as the initial oil supply pressure decreases. Experimental results indicate that FEADRC can adapt to different initial oil supply pressures, exhibiting strong robustness. The oil supply pressure can operate within a wide range, avoiding frequent oil filling and thus improving the lifespan of the hydraulic pump. Furthermore, three performance evaluation indicators calculated from the tracking error, such as... Figure 10 As shown, it can be observed that the absolute value of the maximum error decreases slightly as the initial oil supply pressure gradually increases. Nevertheless, the results indicate that the amplitude of FEADRC does not decrease significantly, while the average tracking error and error standard deviation do not change much, further demonstrating that FEADRC has strong robustness.

[0247] Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention, and not to limit them; although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some or all of the technical features; and these modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the scope of the technical solutions of the embodiments of the present invention.

Claims

1. A control method for a hydraulic servo system for the accumulator oil supply of a shipborne helicopter mooring device, characterized in that, Includes the following steps: S1. Establish a mathematical model of the accumulator oil supply hydraulic servo system; S2. Based on the position of the helicopter probe, plan the velocity trajectory during the capture process of the mechanical claw; S3. Input the displacement feedback signal collected by the displacement sensor into the tracking differentiator to obtain a smooth velocity feedback signal; S4. Compare the speed feedback signal with the speed trajectory planned in S2 to obtain the speed error signal, and generate the feedforward control quantity through the feedforward controller; S5. Based on the mathematical model, design an error active disturbance rejection controller including a feedback compensator and an extended state observer. The extended state observer estimates and compensates for the combined disturbances caused by the accumulator pressure decay and system uncertainty in real time, and generates an error adjustment signal. S6. Synthesize the feedforward control quantity and the error adjustment signal, adjust the parameters of the error active disturbance rejection controller, and form the final control command to drive the hydraulic servo system so that the mechanical gripper tracks the planned speed trajectory.

2. The control method according to claim 1, characterized in that, In S1, the mathematical model is used to describe the nonlinear relationship between accumulator pressure change, servo valve flow rate, and hydraulic cylinder dynamics. It includes three parts: hydraulic cylinder force balance equation, proportional servo valve flow rate equation, and hydraulic cylinder flow rate continuity equation. The mathematical model expression is as follows: Define the state variable as The nonlinear state equation of the system is obtained as follows: ； In the formula, These refer to the displacement and velocity of the hydraulic cylinder piston, respectively. For load pressure; The equivalent mass of the load and the piston rod of the coarse and fine hydraulic cylinders; These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod-side chamber and rodless chamber of the coarse hydraulic cylinder, respectively. These refer to the pressures inside the rod-side chamber and rodless chamber of the fine hydraulic cylinder, respectively. These refer to the pressures within the rod-side and rodless chambers of the coarse hydraulic cylinder, respectively. It is a nonlinear function that includes nonlinear frictional force, unmodeled dynamics of the system, and unknown external disturbance forces; Oil supply pressure for the accumulator; coefficient They are respectively: ； In the formula, The effective bulk modulus of hydraulic oil; ; For servo valve flow gain; ; , These refer to the volumes of the rod-side chamber and the rodless chamber of the fine hydraulic cylinder, respectively. , , and These are the initial volumes of the left and right chambers of the hydraulic cylinder, respectively. This represents the leakage coefficient within the hydraulic cylinder.

3. The control method according to claim 2, characterized in that, The force balance equation for the hydraulic cylinder is as follows: ； In the formula, The equivalent mass of the load and the piston rod of the coarse and fine hydraulic cylinders; These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod-side chamber and rodless chamber of the coarse hydraulic cylinder, respectively. These refer to the pressures within the rod-side and rodless chambers of the coarse hydraulic cylinder, respectively. These refer to the displacement and velocity of the hydraulic cylinder piston, respectively. It is a nonlinear function that includes nonlinear frictional force, unmodeled dynamics of the system, and unknown external disturbance forces; is the viscous damping coefficient.

4. The control method according to claim 2, characterized in that, The nonlinear flow equation for the servo valve is as follows: ； In the formula, For servo valve flow gain; ; Oil supply pressure for the accumulator; flow rate equation , The flow rate into the rod chamber of the fine hydraulic cylinder; The flow rate exiting the rodless chamber of the fine hydraulic cylinder; For lumped control signals, , For error adjustment signal, This is the feedforward control quantity; For load pressure, ,coefficient .

5. The control method according to claim 2, characterized in that, The hydraulic cylinder load flow equation is as follows: ； In the formula, The flow rate into the rod chamber of the fine hydraulic cylinder; The flow rate exiting the rodless chamber of the fine hydraulic cylinder; This refers to the leakage coefficient within the hydraulic cylinder. The effective bulk modulus of hydraulic oil; , These refer to the volumes of the rod-side chamber and the rodless chamber of the fine hydraulic cylinder, respectively. These are the effective working areas of the piston in the rod chamber and rodless chamber of the fine hydraulic cylinder, respectively. The volume descriptions of the rod-side chamber and rodless chamber of the hydraulic cylinder are as follows: ； In the formula, and These are the initial volumes of the rod-side chamber and the rodless chamber of the hydraulic cylinder, respectively.

6. The control method according to claim 1, characterized in that, In S2, based on the position of the helicopter probe, the speed trajectory of the mechanical claw during the capture process is planned. Specifically, during the capture of the helicopter probe by the mechanical claw, the lateral movement speed of the mechanical claw is planned according to the position of the probe to ensure that the tethering device can complete the capture task within the specified time range while minimizing the speed at which a capture collision occurs. The trajectory planning problem is described as follows: For a fixed capture range, a robotic gripper starts from an initial velocity of 0 and flexibly accelerates to a certain velocity within a certain time. It then runs at that constant speed for a period of time; when the robotic gripper approaches the probe, it enters the deceleration phase and decelerates to its speed over the remaining distance. Simultaneously reaching the end of the interval; to ensure the robotic gripper can reach the target speed When capturing the probe, an error buffer zone should be reserved at the end of the capture range, within which the mechanical gripper should always maintain a certain speed. The mechanical gripper moves at a constant speed until the mechanical gripper status detection sensor of the rapid tethering device successfully detects that the mechanical gripper has closed. Then the hydraulic lock is closed, the position of the mechanical gripper is locked, and the speed of the mechanical gripper is reduced to 0. The capture area is divided into three capture zones. Based on the helicopter probe position issued by the central control system, the capture zone in which the target is located is determined. Combined with the capture time requirements, the capture speed curve of the mechanical claw is dynamically planned. ； In the formula, This represents the desired motion velocity trajectory of the robotic gripper. Indicates acceleration time; Indicates the time of uniform motion; Indicates the deceleration time; This represents the maximum expected uniform velocity. This represents the minimum expected uniform motion speed.

7. The control method according to claim 1, characterized in that, In S3, the specific steps include: using a tracking differentiator to obtain the motion velocity from the displacement feedback signal; constructing a second-order discrete tracking differentiator as follows: ； In the formula, For displacement feedback Tracking; displacement signal The differential; Sampling time; , These are adjustable parameters for the fhan function, which is: ； In the formula, Threshold parameters related to acceleration and filtering time; , All are state variables; The predicted value is the discretized velocity value; This is an intermediate quantity representing the nonlinear switching boundary; The predicted value is after nonlinear correction; It is a switching function; This serves as an intermediate output that integrates linear and nonlinear data. This is for secondary switch judgment.

8. The control method according to claim 1, characterized in that, In S4, specifically, based on the mathematical model, the capture velocity of the tethering device is derived as follows: ； in, ; ; ; To avoid the influence of differential pressure signals on control, the feedforward controller is designed as follows: ； In the formula, The desired speed.

9. The control method according to claim 1, characterized in that, In S5, the design process of the error active disturbance rejection controller is as follows: (1) Design of feedback compensator, the specific process is as follows: According to the mathematical model, the state-space expression for the system velocity is: ； In the formula, ; ; ; ; ; ; ; Lumped control signal , For error adjustment signal, This is a feedforward control signal; Redefining The above equation can be rewritten as the chain integrator structure: ； in, ; ； The error in the movement speed and acceleration of the hydraulic cylinder is defined as: , , and These are the desired velocity and desired acceleration, respectively; the hydraulic cylinder velocity error equation is: ； In the formula, This is a comprehensive function of the system state. for The estimated value; Add derivative feedback terms to both sides of the above equation. The dynamic equations for the expected error of the closed-loop system are formed and simplified to: ； In the formula, For total disturbance; Therefore, the feedback compensator is designed as follows: ； In the formula, the gain of the active disturbance rejection controller is... ; (2) The design of the extended state observer is as follows: Define state vector The dynamic equation of the expected error of the closed-loop system is rewritten as: ； In the formula, ; definition To estimate the state vector of an error system, an error-based LESO is designed as follows: ； In the formula, the extended state observer gain vector , .

10. The control method according to claim 1, characterized in that, In S6, the specific process is as follows: the desired closed-loop dynamics are... The pole configuration method is used to make it equal to the desired Hurwitz polynomial. The gain of the active disturbance rejection controller is obtained as follows: ； in, For the error active disturbance rejection controller bandwidth; The ESO estimation error state matrix is ​​represented as follows: The characteristic polynomial is At that time, the gain of the extended state observer is calculated as follows: ； In the formula, the bandwidth of the extended state observer is... ; The parameters that need to be tuned are , , The adjustment steps are as follows: (1) Calculate the parameters based on the known parameters. The nominal value is then fine-tuned during the experiment. (2) Calculate using the nominal values ​​of the parameters according to the mathematical model, and make fine adjustments during the experiment; (3) The larger the size, the stronger the anti-interference ability. An increase in noise will amplify the noise, causing the control signal to vibrate and affecting the stability of the system. Adjustments should be made according to the noise intensity and tracking performance requirements. (4) Affects dynamic response performance, bandwidth yes 3-5 times.

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