Control Method and System for Electric Loader Working Device Based on Composite Pressure Control

Abstract

This invention provides a control method and system for the working device of an electric loader based on composite pressure control, belonging to the field of hydraulic control technology for construction machinery. The method includes: S1, establishing a control model for the working device of the electric loader; S2, controlling the boom raising and bucket lowering actions of the working device of the electric loader through the composite pressure control model of the hydraulic pump; S3, controlling the torque of the servo motor according to the electromagnetic torque control parameters of the servo motor. This invention effectively solves the problems of high energy consumption, low energy recovery efficiency, and poor control accuracy of composite actions in traditional electric loader working devices by adopting differentiated control strategies for individual boom raising or bucket lowering operations, individual boom lowering or bucket tipping operations, and various operations of boom raising and bucket lowering in composite actions, thus significantly improving the endurance and operating efficiency of the electric loader.

Description

Technical Field

[0001] This invention relates to the field of hydraulic control technology for construction machinery, specifically to a control method and system for the working device of an electric loader based on composite pressure control. Background Technology

[0002] As a core product of new energy construction machinery, the energy consumption and energy recovery efficiency of the hydraulic system of the working device of electric loaders directly determine the vehicle's range and operational economy. Traditional electric loaders mostly use valve-controlled hydraulic systems, which use electromagnetic proportional directional valves for throttling speed regulation. During boom lowering and bucket tipping operations, the potential energy of the load's gravity is dissipated as heat through the throttling valve, resulting in serious energy waste. At the same time, during active operations such as boom raising and bucket retraction, the throttling losses caused by speed regulation further reduce the system's energy utilization rate, thus limiting the range performance of electric loaders.

[0003] In existing technologies, some energy recovery solutions use a single energy storage element, such as a lithium battery or supercapacitor, to recover potential energy. However, lithium batteries have low power density and slow charging and discharging response, making them unsuitable for the high-frequency, high-impact potential energy recovery needs of loaders. Supercapacitors have low energy density, and their single energy storage is insufficient to meet the power supply requirements of active working conditions during long-term operation. Furthermore, they lack a hierarchical control strategy designed for the differentiated needs of loaders under various working conditions, such as individual actions and compound actions. This results in uneven flow distribution and system response lag during compound actions, making it difficult to balance operational efficiency and energy recovery efficiency. In addition, traditional load-sensitive systems suffer from slow pump control response and high throttling losses, making them unsuitable for the independent control characteristics of electric loaders' pump sources and hindering the achievement of coordinated control with low losses in active working conditions and high recovery in passive working conditions.

[0004] Therefore, there is an urgent need for a control method and system for electric loader working devices based on composite pressure control, which can solve the technical pain points of high energy consumption, low energy recovery efficiency and poor composite motion control of traditional systems. Summary of the Invention

[0005] To address the shortcomings of existing technologies, this invention aims to provide a control method and system for an electric loader working device based on composite pressure control. For different operating conditions of the electric loader, differentiated control strategies are adopted. For individual boom raising or bucket lowering operations, independent pump-controlled speed regulation is used to reduce throttling losses. For individual boom lowering or bucket tipping operations, a gravity potential energy recovery scheme is employed, with high-pressure oil driving a hydraulic motor to power a generator, which stores the electrical energy in a supercapacitor. In composite actions, composite pressure control enhances the sensitivity of the boom raising and bucket lowering operations through electronically controlled load sensitivity, enabling on-demand flow distribution and improving the coordination of control actions. Cooperative control is used in composite operating conditions, balancing operability and energy saving. This effectively solves the problems of high energy consumption, low energy recovery efficiency, and poor control accuracy in composite actions of traditional electric loader working devices, significantly improving the endurance and operating efficiency of electric loaders.

[0006] Specifically, on the one hand, the present invention provides a control method for the working device of an electric loader based on composite pressure control, which includes the following steps:

[0007] S1: Based on the electromagnetic torque of the servo motor With load torque To establish the dynamic equation of the servo motor of the electric loader based on the balance relationship between them; and based on the actual output flow of the hydraulic pump... and hydraulic pump outlet pressure The dynamic equations of the hydraulic pump in the working device of the electric loader are established; the load pressure control equations in the working device of the electric loader are established; and the control model of the working device of the electric loader is obtained by combining the above.

[0008] S2: Establish a hydraulic pump composite pressure control model that integrates nonlinear eigenmap feedforward compensation, real-time disturbance estimation of the extended state observation model, and sliding mode variable structure control to perform hydraulic pump composite pressure control for the boom lifting and bucket retraction actions of the electric loader's working device; the first-level hydraulic pump composite pressure control model is based on the preset hydraulic pump outlet pressure. and current load traffic Q load To balance flow rate and torque, feedforward compensation control is implemented; the second-layer hydraulic pump composite pressure control model sets the observed estimates of the disturbance extended state variables. As state variables, determine the extended state-space equations of the pressure control model and perform extended state observations; the third-layer hydraulic pump composite pressure control model uses the hydraulic pump outlet pressure as the state variable. Define the sliding surface as the controlled variable. Deriving the sliding mode control law Perform sliding mode variable structure control;

[0009] S3: Observed estimates of the perturbation extended state variables estimated by the extended state observation model. By introducing feedforward compensation into the control law, a composite control is formed, and a new sliding mode control law is output. The total output of the composite pressure control model is obtained as the electromagnetic torque control parameter u of the servo motor, which is used for the torque control of the servo motor.

[0010] Preferably, S1 specifically comprises:

[0011] S11: The speed response of the servo motor shaft is affected by the balance between the electromagnetic torque of the servo motor and the load torque. The dynamic equation of the servo motor in the working device of the electric loader is established.

[0012] S12: Determine the actual output flow rate of the hydraulic pump based on the nonlinear relationship between rotational speed and load pressure; determine the outlet pressure of the hydraulic pump based on the principles of oil continuity and compressibility; determine the dynamic equation of the hydraulic pump in the working device of the electric loader.

[0013] S13: Establish the state-space equations of the pressure control model and obtain the second-order state-space model by combining them; establish the load pressure control equations in the working device of the electric loader.

[0014] Preferably, the load pressure control equation in S13 is described by a state-space equation in affine nonlinear form, specifically as follows:

[0015] ;

[0016] in, The outlet pressure of the hydraulic pump is a state variable. Its first derivative; For the servo motor angular velocity state variable, Its first derivative; This is a nonlinear function in pressure dynamics; This refers to the electromagnetic torque of the servo motor. For the electromagnetic torque of the servo motor The coefficient of influence on the rate of change of pressure; This refers to the lumped disturbance in the pressure loop; It is a nonlinear function in angular velocity dynamics; For the electromagnetic torque of the servo motor The coefficient of influence on the rate of change of angular velocity; This refers to the lumped disturbance in the speed loop; This is a time parameter.

[0017] Preferably, S2 specifically comprises:

[0018] S21: Based on the preset hydraulic pump outlet pressure and current load traffic Q load To achieve flow balance and torque balance under steady-state operating conditions of the system, a feedforward compensation control model is established as the first-level hydraulic pump composite pressure control model.

[0019] S22: Let the observed estimates of the perturbation extended state variables be... For state variables, define the perturbation extended state variable. , To determine the total disturbance of the pressure control model, we define the state-space equation of the extended pressure control model and construct the extended state observation model as the second-layer hydraulic pump composite pressure control model.

[0020] S23: Based on the hydraulic pump outlet pressure Define the sliding surface as the controlled variable. Taking the exponential approach law as the approximation law, the sliding mode control law is derived. A sliding mode variable structure control model was established as a third-layer hydraulic pump composite pressure control model.

[0021] Preferably, the extended state-space equation of the pressure control model in S22 is:

[0022] ;

[0023] in, To extend the state variables for perturbation, Its first derivative; These are the electromagnetic torque control parameters for the servo motor. This is the initial nonlinear nominal function; This is the nominal value of the control gain; Total disturbance of the pressure control model The first derivative; For the electromagnetic torque of the servo motor The influence coefficient; This refers to the displacement of the hydraulic cylinder.

[0024] Preferably, the observation model for the extended state in S22 is:

[0025] ;

[0026] in, The hydraulic pump outlet pressure state variable The observed estimates, Its first derivative; This is the observed estimate of the total disturbance d(t) for the pressure control model. Its first derivative; For the observed estimates of the perturbation extended state variables, Its first derivative; The gain of the first observation model; For the second observation model gain; For the third observation model gain; This is the estimated displacement value of the hydraulic cylinder.

[0027] Preferably, in S23, the sliding surface is set as follows:

[0028] ;

[0029] ;

[0030] in, It is a sliding surface; λ represents the pressure tracking error; λ is the integral gain parameter. For time variables Pressure tracking error during the process; It is a time variable; This refers to the outlet pressure of the hydraulic pump. Preset the hydraulic pump outlet pressure; This is a time parameter.

[0031] Preferably, the servo motor electromagnetic torque control parameters in S3 for:

[0032] ;

[0033] ;

[0034] in, This is the nominal value of the control gain; The electromagnetic torque of the servo motor is output as a feedforward compensation. This is a sliding mode control law; This is the equivalent control portion; For switching control section; These are the electromagnetic torque control parameters for the servo motor.

[0035] On the other hand, the present invention provides a control system for an electric loader working device based on composite pressure control, which includes: a hydraulic power module, an energy recovery module, a composite power supply module and a control module;

[0036] The hydraulic power module provides active operating power for the pressure control model, including an oil tank, servo motor, first coupling, hydraulic pump, first pressure sensor, first relief valve, first pressure compensation valve, second pressure compensation valve, first electromagnetic proportional directional valve, second electromagnetic proportional directional valve, first shuttle valve, second shuttle valve, third shuttle valve, second pressure sensor, boom hydraulic cylinder and bucket hydraulic cylinder.

[0037] The energy recovery module includes a generator, a second coupling, a hydraulic motor, a first check valve, a second check valve, a second relief valve, a third check valve, a fourth check valve, a first switching valve, and a second switching valve.

[0038] The composite power module includes a lithium-ion battery, a battery management system, a supercapacitor, a capacitor management system, and a DC bus. It uses a composite energy storage architecture composed of lithium-ion batteries and supercapacitors to recover gravitational potential energy under the conditions of falling boom and tipping bucket operation.

[0039] The control module includes a boom handle, a bucket handle, a pressure control model, a motor driver, and a generator driver.

[0040] Preferably, the hydraulic motor in the energy recovery module is a fixed displacement motor, which is coaxially connected to the generator through a second coupling. The rotation of the hydraulic motor directly drives the generator to generate electricity, thus converting hydraulic energy into electrical energy.

[0041] The hydraulic motor has three inlet ports: the first port connects to the outlet ports of the first and second switching valves; the second port connects to the second relief valve; and the third port connects to the second check valve. The inlets of the first and second switching valves are connected to the rodless chambers of the boom hydraulic cylinder and the bucket hydraulic cylinder, respectively, to facilitate the flow of high-pressure oil discharged from the rodless chambers of the hydraulic cylinders during boom lowering and bucket tipping. The return port of the second relief valve connects to the oil tank to limit the inlet pressure of the hydraulic motor. The inlet port of the second check valve connects to the oil tank, and the outlet port is connected to the inlet port of the hydraulic motor, automatically replenishing oil when the inlet pressure of the hydraulic motor is lower than the opening pressure of the second check valve.

[0042] The hydraulic motor has three outlet ports: the first port is connected to the first check valve, the second port is connected to the inlet of the third check valve, and the third port is connected to the inlet of the fourth check valve. The outlet of the first check valve is connected to the oil tank to provide stable back pressure for the hydraulic motor. The outlets of the third and fourth check valves are connected to the rod chambers of the boom hydraulic cylinder and the bucket hydraulic cylinder, respectively, so that the oil discharged from the hydraulic motor can flow into the rod chamber of the hydraulic cylinder.

[0043] Compared with the prior art, the beneficial effects of the present invention are as follows:

[0044] (1) This invention is compatible with all tasks and working conditions of the working device of electric loader; it covers all working conditions of standby, individual operation and compound operation, meets the actual operation needs of electric loader, and solves the limitation of existing technical means that only adapt to a single task working condition.

[0045] (2) The present invention completely eliminates energy loss by using independent pump control speed regulation for the separate lifting and bucket retraction of the electric loader working device; the boom drop and bucket tipping adopt gravity potential energy recovery, so that the oil can be reused and the energy utilization rate is greatly improved.

[0046] (3) The present invention can control the compound action of the working device of the electric loader, improve the sensitivity of the load of the lifting arm and the bucket by the electric control, realize the flow distribution on demand, and control the coordination of the action; in the compound working condition, the collaborative control is adopted to take into account both operability and energy saving.

[0047] (4) The present invention optimizes power management by using a combination of lithium-ion battery and supercapacitor composite power supply, and uses supercapacitor to quickly store and recover electrical energy, thereby improving energy recovery efficiency and extending battery life. Attached Figure Description

[0048] Figure 1 This is a control block diagram of the electric loader working device control method based on composite pressure control according to the present invention.

[0049] Figure 2 This is a hydraulic schematic diagram of the electric loader of the present invention;

[0050] Figure 3 This is a control flowchart of the electric loader working device of the present invention;

[0051] Figure 4 This is a block diagram of the composite pressure control system for the electric loader of the present invention;

[0052] Figure 5 This is a step pressure control curve diagram according to an embodiment of the present invention;

[0053] Figure 6 This is a step pressure control error curve diagram according to an embodiment of the present invention;

[0054] Figure 7 This is a sinusoidal pressure control curve diagram of an embodiment of the present invention;

[0055] Figure 8 This is a sinusoidal pressure control error curve diagram according to an embodiment of the present invention.

[0056] Key reference numerals:

[0057] 1. Oil tank; 2. Servo motor; 3. First coupling; 4. Hydraulic pump; 5. First pressure sensor; 6. First relief valve; 7. Generator; 8. Second coupling; 9. Hydraulic motor; 10. First check valve; 11. Second check valve; 12. Second relief valve; 13. First pressure compensation valve; 14. Second pressure compensation valve; 15. First electromagnetic proportional directional valve; 16. Second electromagnetic proportional directional valve; 17. First shuttle valve; 18. Second shuttle valve; 19. 20. Third shuttle valve; 21. Second pressure sensor; 22. Third check valve; 23. Fourth check valve; 24. First switching valve; 25. Second switching valve; 26. Boom hydraulic cylinder; 27. Bucket hydraulic cylinder; 28. Boom handle; 29. ​​Bucket handle; 30. Pressure control model; 31. Motor driver; 32. Generator driver; 33. Battery management system; 34. Lithium-ion battery; 35. Capacitor management system; 36. Supercapacitor; 37. DC bus. Detailed Implementation

[0058] Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

[0059] The first aspect of this invention proposes a control method for the working device of an electric loader based on composite pressure control, such as... Figure 1 As shown, a control model for the working device of an electric loader is established; the boom-lifting and bucket-retracting actions of the electric loader's working device are controlled through a hydraulic pump composite pressure control model; and the torque of the servo motor is controlled according to the electromagnetic torque control parameters of the servo motor. This invention explicitly uses the working device of the electric loader as the controlled object. This term has a specific and generally accepted meaning in the field of construction machinery, specifically referring to the working mechanism composed of the boom, bucket, and their hydraulic drive system, which is the physical carrier and functional implementation platform for all control strategies and working condition divisions in this invention. This embodiment of the invention addresses the composite action conditions of the electric loader's working device, specifically the boom-lifting and bucket-retracting actions of the electric loader's working device, such as... Figure 2 The diagram shown is a hydraulic schematic of the electric loader of the present invention, specifically: the hydraulic pump outlet pressure state variable controlled by the control model. Servo motor angular velocity state variable , This is the coefficient for the fifth operating condition. For boom handle opening, This is the coefficient for the sixth operating condition. To adjust the bucket handle opening; through pressure closed-loop control, the preset hydraulic pump outlet pressure is maintained. , For maximum load pressure, The pressure difference is constant. Due to the multi-source strong nonlinear disturbances faced by the hydraulic power module of the electric loader's working device, such as hydraulic pump flow pulsation, low-speed dead zone, friction nonlinearity, time-varying oil elastic modulus, and uncertain leakage coefficient, high-precision control of the hydraulic pump outlet pressure becomes a key challenge. To address this, this invention establishes a complete mathematical model based on the system's physical mechanism, incorporating servo motor dynamics, hydraulic pump fluid dynamics, and load circuit pressure dynamics. Based on this, a three-layer composite pressure control algorithm is proposed, integrating nonlinear eigenmap feedforward compensation, real-time disturbance estimation using an extended state observation model, and sliding mode variable structure control. The specific steps are as follows:

[0060] S1: Establish a control model for the working device of the electric loader.

[0061] S11: Determine the dynamic equation of the servo motor in the working device of the electric loader; as the driving core of the electric loader's working device, the balance between the electromagnetic torque of the servo motor and the load torque determines the speed response of the motor shaft; based on the basic theory of permanent magnet synchronous motors, the dynamic equation of the servo motor is established as follows:

[0062] ;

[0063] ;

[0064] in, This refers to the angular velocity of the servo motor. This is the total moment of inertia of the servo motor rotor and the hydraulic pump shaft. This refers to the electromagnetic torque of the servo motor. This is the load torque; It is the viscous damping coefficient; It is a nonlinear frictional torque; The pressure reaction torque required to drive the hydraulic pump; This refers to the outlet pressure of the hydraulic pump. This refers to the displacement of the hydraulic pump. Pi (π) is a parameter. For the mechanical efficiency of the hydraulic pump; This is a time parameter.

[0065] In the dynamic equations of the servo motor, this invention makes the following targeted adjustments to the traditional servo motor model: the friction torque T is... f (ω m The model extends the conventional linear viscous friction to a complete nonlinear friction model incorporating the Stribeck effect, describing the positive and negative damping switching characteristics of the motor during low-speed startup and near zero speed; simultaneously, it considers the pressure reaction torque T required to drive the hydraulic pump. pump Explicitly expressed as hydraulic pump outlet pressure p p With the displacement D of the hydraulic pump p The product divided by the hydraulic pump mechanical efficiency η mh Instead of simply attributing it to the load torque T L This approach treats the pump's load characteristics together, decoupling them from the external load and facilitating the independent design of subsequent pressure and speed control circuits. The core improvement lies in the fact that frictional torque is often simplified to B in traditional models. m ω m The single linear term neglects the negative damping effect in the low-speed region. This invention introduces a complete Stribeck friction model, enabling the controller to specifically compensate for low-speed crawling phenomena. This is one of the important improvements that distinguishes this invention from the prior art.

[0066] S12: Determine the dynamic equation of the hydraulic pump in the working device of the electric loader; the output flow of the hydraulic pump is simultaneously affected by the nonlinearity of the rotational speed and load pressure; under ideal conditions, the output flow of the hydraulic pump is linearly related to the rotational speed, but in actual operation, the flow characteristics exhibit strong nonlinearity; the actual output flow of the hydraulic pump is expressed as the difference between the theoretical flow and the internal leakage, specifically:

[0067] ;

[0068] in, This refers to the actual output flow rate of the hydraulic pump. The internal leakage coefficient is related to the hydraulic pump outlet pressure. Related; This refers to the flow pulsation term of the hydraulic pump.

[0069] Based on the principles of hydraulic fluid continuity and compressibility, the dynamic pressure at the hydraulic pump outlet is determined by both the net flow rate entering the control volume and the volumetric compression effect, specifically:

[0070] ;

[0071] in, This refers to the control volume between the hydraulic pump inlet and the load. This is the effective bulk modulus of the hydraulic fluid, a parameter that varies with the outlet pressure of the hydraulic pump. and hydraulic pump outlet oil temperature Time-varying; The outlet oil temperature of the hydraulic pump; This represents the total leakage flow rate of the hydraulic pump. Traffic consumed by the load.

[0072] In the actual output flow equation of the hydraulic pump, this invention makes two key improvements to the traditional pump flow model: First, the internal leakage coefficient C is... ip The constant was changed to the hydraulic pump outlet pressure p. p The function is used to reflect the nonlinear increase in leakage caused by the increase in the gap between the plunger and the cylinder under high pressure. Engineering measurements show that when the outlet pressure increases from 5 MPa to 20 MPa, the internal leakage coefficient increases by 30% to 50%. Using a constant leakage coefficient will introduce model errors. Secondly, the flow pulsation term is independently subtracted from the theoretical flow rate. Instead of treating it as an unmodeled dynamic and passively suppressing it through a feedback controller, the flow pulsation term takes a form related to the number of pistons and their rotational speed, with a fundamental frequency of z·ω. m / (2π), by explicitly incorporating this periodic disturbance at the model level, the present invention enables feedforward compensation to actively counteract the pulsation effect. This is an essential improvement of the present invention compared to the traditional method that treats pulsation as a random disturbance.

[0073] In the dynamic equation of pump inlet pressure, this invention uses the effective bulk elastic modulus β of the oil. e Processed as hydraulic pump outlet pressure p p It is a bivariate function of oil temperature θ, rather than a constant used in traditional models. The physical basis for this is that when the pressure is <5MPa, the air mixed in the oil significantly affects compressibility, β... e As low as 500 MPa; when the pressure rises above 10 MPa, the air gradually dissolves, β eThe pressure rises to 1400–1700 MPa. This pressure-dependent characteristic directly affects the dynamic response speed of pressure. β e The lower the β value, the smaller the rate of pressure change caused by the same change in net flow, and the slower the system response. This invention addresses this by explicitly incorporating β into the model. e The pressure and temperature dependence of the controller enables it to adaptively adjust control parameters according to the current operating conditions, which is a key modeling improvement for high-precision pressure control under all operating conditions.

[0074] S13: Establish the load pressure control equation in the working device of the electric loader; for the independent control of the hydraulic pump outlet pressure, the dynamic expression of the hydraulic pump outlet pressure is as follows:

[0075] ;

[0076] in, Equivalent load area; For hydraulic cylinder displacement, Its first derivative; β is the external leakage coefficient; e This is the effective bulk elastic modulus of the oil.

[0077] The state-space equation of the pressure control model is obtained; the state variable of the hydraulic pump outlet pressure is defined. = servo motor angular velocity state variable = Solving the above equations simultaneously, we obtain the second-order state-space model of the control system as follows:

[0078] ;

[0079] in, The outlet pressure of the hydraulic pump is a state variable. Its first derivative; For the servo motor angular velocity state variable, Its first derivative; Q load The required flow rate for the load.

[0080] Further analysis of the total disturbance of the pressure control model, including parameter uncertainties, unmodeled dynamics, and external disturbances, is combined into the total disturbance of the pressure control model. The pressure control equation is simplified into a more compact affine nonlinear form, specifically:

[0081] ;

[0082] in, This is a nonlinear function in pressure dynamics; For the electromagnetic torque of the servo motor The coefficient of influence on the rate of change of pressure; This refers to the lumped disturbance in the pressure loop; It is a nonlinear function in angular velocity dynamics; For the electromagnetic torque of the servo motor The coefficient of influence on the rate of change of angular velocity; This refers to the lumped disturbance in the speed loop.

[0083] In obtaining the second-order state-space model by simultaneously solving the above equations, this invention decomposes the system into two coupled second-order components: a pressure dynamic subsystem and a speed dynamic subsystem. The structural feature of this state-space model is that it incorporates the electromagnetic torque of the servo motor... As the sole control input, the servo motor angular velocity state variable x2 indirectly affects the hydraulic pump outlet pressure state variable x1, which is a typical cascaded control structure. Based on this model, this invention further merges model uncertainties and external disturbances into lumped disturbance d1(t) in the pressure loop and lumped disturbance d2(t) in the speed loop, providing a clear disturbance estimation object for the subsequent design of the extended state observer.

[0084] The output variable for this step is the first derivative of the hydraulic pump outlet pressure state variable. and the first derivative of the angular velocity state variable of the servo motor This was used in the state-space equations of the pressure control model established in subsequent S22.

[0085] S2: Establish a hydraulic pump composite pressure control model integrating nonlinear eigenmap feedforward compensation, extended state observation model real-time disturbance estimation, and sliding mode variable structure control to perform hydraulic pump composite pressure control for the boom-lifting and bucket-retracting actions of the electric loader's working device. The first layer establishes a nonlinear eigenmap library for flow rate, pressure, and speed through offline identification or online learning, performing open-loop feedforward compensation for the dominant nonlinearity of the control system. The second layer uses an extended state observation model to perform online real-time estimation of model uncertainties and unmodeled dynamics. The third layer uses a sliding mode control model as the core to improve the robustness of the pressure control model to residual disturbances. Figure 4 The diagram shows the control block diagram of a three-layer composite pressure control system. This control system uses the target outlet pressure of the hydraulic pump as input and directly generates the feedforward control quantity through a feedforward compensation control model. Simultaneously, the target pressure at the hydraulic pump outlet and the actual hydraulic pump outlet pressure... deviation The input is fed into the sliding mode variable structure control model, and the output is the feedback control quantity. The sum of the two yields the total control quantity. The pressure is applied to the motor driver, which controls the servo motor to rotate and drive the hydraulic pump to output pressure and flow. The pressure control model outputs actual pressure. Feedback is sent to the input. The extended state observation model simultaneously utilizes the total control input. and actual pressure Real-time estimation of observed estimates of perturbation-extended state variables It is then fed into the sliding mode variable structure control model for feedforward compensation, thereby forming a three-layer composite control loop of feedforward matching, sliding mode feedback, and online disturbance estimation, which improves the suppression performance of nonlinear disturbances such as hydraulic pump flow pulsation and time-varying parameters.

[0086] S21: Establish a feedforward compensation control model, which is the first-level hydraulic pump composite pressure control model. The core objective of feedforward compensation is: based on the flow balance and torque balance relationship under the steady-state operating conditions of the system, and according to the preset hydraulic pump outlet pressure... and current load traffic Q load The required electromagnetic torque feedforward quantity u of the servo motor can be directly solved. ff This allows the pump inlet pressure to approach the set target value in an open-loop sense, reducing the adjustment burden on the feedback controller.

[0087] When the pump inlet pressure reaches the target value and the system is in steady state, the theoretical flow rate output by the pump equals the sum of the load demand flow rate, the internal leakage flow rate, and the flow pulsation compensation. The steady-state flow balance equation is:

[0088] ;

[0089] Among them, Q load D is the required flow rate for the load. p This refers to the displacement of the hydraulic pump. The required reference angular velocity of the motor to achieve the target pressure; The internal leakage coefficient is influenced by pressure. This represents the average value of the flow pulsation.

[0090] Due to the average value of flow pulsation The motor reference angular velocity required to achieve the target pressure The implicit nonlinear function makes it difficult to directly and analytically solve for the motor reference angular velocity required to achieve the target pressure. Numerical methods, such as Newton-Raphson iteration or interpolation in an offline-built inverse mapping table, are required to obtain the rotational speed setpoint that satisfies flow balance.

[0091] ;

[0092] in, The inverse nonlinear mapping operator is identified based on experimental data. This mapping comprehensively analyzes the nonlinear characteristics of the pump, such as displacement, pressure-related leakage, and flow pulsation; θ is the oil temperature, which serves as an environmental compensation parameter.

[0093] The motor reference angular velocity required to achieve the target pressure Under these conditions, the electromagnetic torque of the servo motor must simultaneously overcome the pump's pressure reaction torque, viscous damping torque, and nonlinear friction torque; the steady-state torque balance equation is:

[0094] ;

[0095] Among them, u ff The electromagnetic torque of the servo motor output for feedforward compensation is the feedforward control quantity; η mh T represents the mechanical-hydraulic efficiency of the pump. f It is a nonlinear frictional torque; is the viscous damping coefficient.

[0096] The output variable in this step is the feedforward control variable. This was used in the subsequent S3 process to determine the electromagnetic torque control parameters of the output servo motor.

[0097] In the steady-state flow balance equation of feedforward compensation, each term is handled as follows: This is the pump's theoretical flow rate, which is directly proportional to the rotational speed. The internal leakage flow rate is represented by the internal leakage coefficient C in this embodiment of the invention. ip A cubic polynomial is used for fitting, specifically:

[0098] C ip (p) = c0 + c1p + c2p 2 +c3p 3 ;

[0099] Among them, C ip denoted as the internal leakage coefficient; p is the pressure parameter; c0 is the first coefficient identified through offline leakage testing, c1 is the second coefficient identified through offline leakage testing, c2 is the third coefficient identified through offline leakage testing, and c3 is the fourth coefficient identified through offline leakage testing, in order to accurately characterize the nonlinear growth of leakage as pressure increases; This represents the average value of the flow pulsation over a period of time.

[0100] For a nine-plunger pump, the theoretical pulsation amplitude is 1.5% to 2.5% of the average flow rate, and the average value of the flow pulsation is:

[0101] ;

[0102] Where z is the number of plungers; A pulThis is the pulsation amplitude coefficient; is the initial phase; sin is the sine function.

[0103] In the steady-state solution, the average value over a complete pulsation cycle is taken to smooth the influence of periodic pulsations on the steady-state speed solution. To facilitate understanding of the engineering meaning of the above formula, a typical calculation example is given below. Taking the combined working condition of a certain type of electric loader's boom lifting and bucket retraction as an example, the system parameters are: hydraulic pump displacement D... p =80mL / r, target outlet pressure =18MPa, load demand flow Q load =120L / min, oil temperature θ=40℃, number of plungers z=9. First, solve for the target speed: from the steady-state flow balance equation, substitute C ip (18MPa) = 1.2 × 10 -12 m 3 / (s·Pa), this value was obtained by fitting through offline leakage tests, internal leakage flow rate C ip ·p=1.2×10 -12 ×18×10 6 ≈1.3L / min, average pulse value ≈1.5%×Q load =1.8L / min, then the theoretically required flow rate ≈120 + 1.3 + 1.8 = 123.1 L / min, solving for... ≈1538r / min.

[0104] Then calculate the electromagnetic torque of the feedforward servo motor: substitute η mh (18MPa, 161rad / s) = 0.88, interpolated from the manufacturer's efficiency graph, pressure reaction torque. ≈260.5 N·m; the nonlinear frictional torque is estimated according to the Stribeck model, taking F... c =5 N·m, F s =8 N·m, ω s =5rad / s, B v =0.02 N·m·s / rad, then T f (161) = 8.22 N·m; Viscous damping torque is taken as B. m =0.005 N·m·s / rad, =0.005×161=0.805N·m. The total of the three is u. ff =260.5 + 8.22 + 0.805 ≈ 269.5 N·m. This calculated value is the electromagnetic torque control command of the servo motor output by the feedforward compensation model under the current operating conditions. If the traditional fixed efficiency method is used, take η... mh=0.92 constant, then the pressure reaction torque is 249.2 N·m, and the feedforward torque is 258.2 N·m, with a difference of 11.3 N·m, or 4.2%. For a high-pressure system of 18 MPa, this deviation will directly lead to a steady-state pressure error of 0.75 MPa, demonstrating the necessity of using pressure-related efficiency and a complete friction model for feedforward compensation in this invention.

[0105] S22: Establish an extended state observation model for the second-level hydraulic pump composite pressure control model. The observed estimates of the disturbance extended state variables... Expand to a second state variable, and define the perturbation extended state variable. And assuming it is bounded. Then the extended state-space equation of the pressure control model is:

[0106] ;

[0107] in, To extend the state variables for perturbation, Its first derivative; These are the electromagnetic torque control parameters for the servo motor. ; This is the initial nonlinear nominal function; This is the nominal value of the control gain; Total disturbance of the pressure control model The first derivative; For the electromagnetic torque of the servo motor The influence coefficient; This refers to the displacement of the hydraulic cylinder.

[0108] The extended state observation model is constructed as follows:

[0109] ;

[0110] in, The hydraulic pump outlet pressure state variable The observed estimates, Its first derivative; This is the observed estimate of the total disturbance d(t) for the pressure control model. Its first derivative; Here, represents the observed estimate of the extended state variable due to the disturbance, and represents the rate of change of the disturbance. The observed estimates, Its first derivative; The gain of the first observation model; For the second observation model gain; This is the gain of the third observation model.

[0111] Select an appropriate observation model gain , , The observation error is limited to converge to a bounded region within a finite time, specifically: the observation model gain. , , The characteristic equation is tuned using the bandwidth parameterization method, i.e., let (s+ω) be the characteristic equation. o ) 3 ,Pick =3ω o , =3ω o 2 , =ω o 3 Where bandwidth ω o Typically, ω is chosen to be 3 to 10 times the closed-loop control bandwidth, and this is achieved by gradually increasing ω while satisfying the sampling frequency constraint. o The process continues until the optimal control effect is achieved, at which point the engineering fine-tuning is completed.

[0112] The output variable in this step is the observed estimate of the perturbation-spread state variable. This was used in the subsequent S3 to output the electromagnetic torque control parameters of the servo motor.

[0113] In the extended state observation model, For the observed value of actual export pressure x1, Let be the observed value of the total disturbance d(t) in the pressure control model. Let be the observed value of the rate of change of the disturbance, h(t). The core design idea of ​​the observer is to include all factors that are difficult to model precisely, including parameter perturbations, unmodeled dynamics, and external load changes, into the total disturbance d(t). Then, online estimation and feedforward compensation are performed through extended states, so that the feedback controller only needs to deal with the residual estimation error, rather than directly facing all uncertainties. l1, l2, and l3 are the observation model gains, which are obtained using the bandwidth parameterization method. =3ω o , =3ω o 2 , =ω o 3 Tuning, where ω o This represents the observer bandwidth.

[0114] Taking the working device of the electric loader in this embodiment as an example, the system sampling period T s =1ms, hydraulic natural frequency ω h ≈80 rad / s. The closed-loop control bandwidth ω is taken as... c =0.3ω h=24 rad / s, observer bandwidth ω o =5ω c =120rad / s, then the observation model gain is: l1=3×120=360, l2=3×120 2 =43200, l3=120 3 =1728000. Verify the discretization stability condition: ω o ·T s =120×0.001=0.12<0.3, which meets the requirement, indicating that the observer operates stably at the current sampling frequency with this gain set. Under this gain set, the characteristic equation of the observer is (s+120). 3 =0, all three poles are located at -120 on the real axis, and the convergence time constant of the observation error is 3 / ω. o =25ms, much faster than the response time of the pressure closed loop, which is 220ms in the embodiment, enabling the observer to complete the disturbance estimation before the controller, thus meeting the design requirements of dynamic separation of the system.

[0115] S23: Establish a sliding mode variable structure control model, which is a composite pressure control model for the third-layer hydraulic pump.

[0116] hydraulic pump outlet pressure Let the sliding surface be the controlled variable:

[0117] ;

[0118] in, It is a sliding surface; For pressure tracking error, λ is the integral gain parameter. The introduction of the integral term eliminates the steady-state error, and λ>0. For time variables Pressure tracking error during the process; It is a time variable.

[0119] Taking the exponential law of convergence as the law of convergence, we have:

[0120] ;

[0121] in, The first derivative of the sliding surface; These are the coefficients of the isotropic approaching term; This is a sign function, taking the value 1 when s > 0, -1 when s < 0, and 0 when s = 0; The coefficient of the exponential approach term.

[0122] The sliding mode control law is derived as follows:

[0123] ;

[0124] in, This is a sliding mode control law; This is the equivalent control portion; This is for switching control.

[0125] Equivalent control part pass Solve and switch control section Determined by the exponential reaching law.

[0126] The output variable in this step is the equivalent control part. and switching control section This was used in the subsequent S3 to output the electromagnetic torque control parameters of the servo motor.

[0127] In the sliding surface s, e represents the pressure tracking error. The integral term is introduced to eliminate the steady-state pressure error. The selection of the integral gain parameter λ is based on the following: if λ is too small, the integral action is weak and the steady-state error converges slowly; if λ is too large, it easily causes integral saturation and overshoot. In this invention, λ = 0.5ω is chosen. c =12s -1 , where ω c =24rad / s is the closed-loop bandwidth, so that the corner frequency of the integral element is lower than the closed-loop bandwidth, thus avoiding interference from the integral effect on the main dynamic response of the system.

[0128] In the exponential law In this equation, ε is the constant-rate approaching term coefficient, which means that when the system state is far from the sliding surface, it approaches the sliding surface at a constant rate ε; k is the exponential approaching term coefficient, which means that when the system state is close to the sliding surface, the approaching rate is proportional to the distance to the sliding surface |s|, achieving exponential decay convergence. Their functions are: ε ensures reachability when far from the sliding surface, and k ensures smoothness when approaching the sliding surface. In this embodiment, ε = 0.5 MPa / s and k = 30 s are used. -1 The value of ε must be greater than the upper bound of the rate of change of the total system disturbance, which is estimated to be 0.3 MPa / s through offline identification, so that the sliding mode reaches the condition. <0 is true.

[0129] In sliding mode control law In the equivalent control section By order =0 Solving for this value represents the control quantity required to maintain the system on the sliding surface when the model is completely accurate and undisturbed; switching the control section The exponential reaching law is used to overcome model uncertainties and external disturbances. To eliminate the chattering phenomenon inherent in sliding mode control, this invention uses a saturation function sat(s / Φ) instead of the ideal sign function sgn(s), where the boundary layer thickness is Φ=0.1MPa, which is twice the allowable value of steady-state pressure error, thus achieving a balance between suppressing chattering and improving control accuracy.

[0130] S3: Observed estimates of the perturbation extended state variables estimated by the extended state observation model. By introducing feedforward compensation into the control law to form a composite control, the sliding mode control law is re-output as follows:

[0131] ;

[0132] in, This is the nominal value for the control gain.

[0133] In summary, the total output of the composite pressure control model is obtained as follows: The electromagnetic torque control parameters for the servo motor are:

[0134] ;

[0135] in, These are the electromagnetic torque control parameters for the servo motor. The electromagnetic torque of the servo motor is output for feedforward compensation.

[0136] The output variable for this step is the servo motor electromagnetic torque control parameter. This is used for torque control of servo motors. The embodiments of this invention estimate the total disturbance using the extended state observation model. Introducing control laws to form composite control The physical meaning of this formula is: the feedback control quantity output by the sliding mode controller must not only include the equivalent control u required to maintain the sliding mode motion. eq and switching control to overcome uncertainty sw The total disturbance estimated by the observer should also be subtracted. The compensation amount obtained by dividing by the nominal gain b0 is used to preemptively cancel the effects of disturbances in the control signal, so that the feedback controller only needs to deal with the residual estimation error rather than all disturbances.

[0137] The total output of the composite pressure control model is the electromagnetic torque control parameter of the servo motor. Taking the aforementioned calculation data for the boom-lifting and bucket-retracting working condition as an example: feedforward torque u ff =269.5 N·m, let the total disturbance estimated by the observer at the current time be 269.5 N·m. =5MPa / s, indicating the pressure change rate disturbance caused by sudden load changes or parameter variations, with a nominal gain b0 ≈ 8.91 × 10⁻⁶. 6MPa / (s·N·m), then the disturbance compensation term ≈5.61×10 -7 N·m, this value is extremely small compared to the feedforward torque of several hundred Newton-meters, indicating that under normal operating conditions, feedforward compensation has undertaken the vast majority of the control task, and feedback control only needs fine-tuning. When the system encounters a drastic load change, such as the bucket suddenly cutting into the stockpile, When the pressure increases instantaneously to 50–100 MPa / s, the disturbance compensation term reaches 10. -5 The pressure is on the order of N·m, and feedback control will actively intervene to suppress pressure fluctuations; this clear-cut distribution of control quantities fully embodies the three-layer collaborative design concept of this invention: feedforward coarse adjustment, feedback fine adjustment, and real-time compensation by the observer.

[0138] This invention addresses the actual use of the electric loader's working device by dividing it into three modes: standby mode, individual operation mode, and combined operation mode, specifically:

[0139] (1) In standby mode: the boom handle and bucket handle are not activated. The pressure control model controls the first electromagnetic proportional directional valve first coil YA1, the first electromagnetic proportional directional valve second coil YA2, the second electromagnetic proportional directional valve first coil YA3, the second electromagnetic proportional directional valve second coil YA4, the first switch valve coil DT1, and the second switch valve coil DT2 to be de-energized. The battery management system and capacitor management system are turned off. The servo motor and generator speeds are 0. The pressure control model is in standby mode.

[0140] (2) The individual action conditions are divided into: single action bucket closing, single action arm raising, single action bucket tipping and single action arm lowering.

[0141] In single-action bucket closing and single-action boom raising modes: the control model keeps the second electromagnetic proportional directional valve / first electromagnetic proportional directional valve at its maximum opening, and the first and second switching valves are de-energized and closed; the handle opening is proportional to the servo motor speed, specifically: Single-action bucket closing: servo motor speed , For bucket handle opening, This refers to the hydraulic pump displacement. The first operating condition coefficient; single-action lifting arm: servo motor speed , For boom handle opening, The third operating condition coefficient is set; the battery management system and capacitor management system are turned on, and the composite power supply powers the servo motor, realizing independent pump control speed regulation and eliminating throttling losses.

[0142] In single-action tipping and single-action boom lowering modes: the control model energizes and opens the second / first switching valve, while de-energizing and closing the first and second electromagnetic proportional directional valves; the handle opening is proportional to the generator speed, specifically: Single-action tipping: generator speed... , This is the second operating condition coefficient. Hydraulic motor displacement; Single-acting boom: Generator speed , The fourth working condition coefficient is applied; when the capacitor management system is activated, the gravitational potential energy of the boom / bucket drives the high-pressure oil in the rodless chamber of the hydraulic cylinder through the switching valve to drive the hydraulic motor, which in turn drives the generator to generate electricity, which is then stored in the supercapacitor.

[0143] (3) The compound action conditions are divided into: raising the arm to collect the bucket, raising the arm to flip the bucket, lowering the arm to collect the bucket and lowering the arm to flip the bucket.

[0144] In the boom-and-bucket-retract state: the control model controls the hydraulic pump outlet pressure state variable and the servo motor angular velocity state variable, specifically: the hydraulic pump outlet pressure state variable. servo motor angular velocity state variable Through pressure closed-loop control, the preset hydraulic pump outlet pressure is ensured to meet the following requirements: .

[0145] In the boom-tilting state: the boom uses an independent pump control scheme, and the servo motor speed... Hydraulic pump outlet pressure state variable , This represents the maximum displacement of the first electromagnetic proportional directional valve; the tipping bucket employs an energy recovery scheme, and the generator speed... servo motor angular velocity state variable When the second switch valve is energized and opened, the battery management system and capacitor management system work together to control the lifting arm for power supply and the tipping bucket for recovery.

[0146] In the boom-and-bucket-folding mode: the bucket folding adopts an independent pump control scheme, with servo motor speed... servo motor angular velocity state variable The boom employs an energy recovery system, and the generator speed... Hydraulic pump outlet pressure state variable When the first switch valve is energized and opened, the battery management system and capacitor management system work together to power the bucket and retract the boom.

[0147] In the boom-and-dump configuration: both boom-and-dump configurations utilize energy recovery systems, and the generator speed... Hydraulic pump outlet pressure state variable servo motor angular velocity state variable When the first and second switching valves are energized and opened, the capacitor management system is activated, and the dual-cylinder high-pressure oil drives the hydraulic motor to generate electricity, which is then stored in the supercapacitor.

[0148] To verify the effectiveness of the proposed three-layer composite pressure control method, which integrates nonlinear eigenmap feedforward compensation, real-time disturbance estimation of the extended state observation model, and sliding mode variable structure control, simulation studies were conducted under step pressure command and sinusoidal pressure command for the boom raising and bucket retraction working conditions under combined action conditions.

[0149] The differentiated control proposed in this invention is the core inventive concept at the strategy level. Its essence lies in matching and executing different control logics and energy management schemes according to different operating conditions of the electric loader's working device—including standby, individual actions of boom raising, bucket retraction, boom lowering, and bucket tipping, as well as composite actions of boom raising and bucket retraction, boom raising and bucket tipping, boom lowering and bucket retraction, and boom lowering and bucket tipping—to achieve energy consumption optimization and efficiency improvement across all operating conditions. The three-layer composite pressure control, on the other hand, is the specific implementation method proposed at the algorithm level for the specific operating condition of boom raising and bucket retraction within the composite actions. It consists of a feedforward compensation control layer, an extended state observation compensation layer, and a sliding mode variable structure control layer, used to solve the problem of multi-source strong nonlinear disturbances in the hydraulic pump outlet pressure under this operating condition, thereby achieving high-precision pressure closed-loop control. In short, differentiated control is a macro-level strategic framework that determines when to use which method, while three-layer composite pressure control is a micro-level technical means within this framework for a specific operating condition. The former encompasses the latter but its function is far beyond it; the two have a hierarchical relationship between the strategy level and the method level.

[0150] like Figure 5 The figure shows the step pressure control curve of an embodiment of the present invention. The figure compares the pressure response curves of the PID algorithm and the composite control algorithm of the present invention under a 10MPa step pressure command. As can be seen from the figure, both control algorithms can effectively track the step pressure, but the dynamic response processes differ significantly: the PID control curve exhibits a large overshoot oscillation after the step jump, with a maximum overshoot of 6.33%, and takes 0.36s to stabilize; while the response curve of the composite control algorithm of the present invention rises more smoothly, with a maximum overshoot of only 1.52%, and reaches steady state in 0.22s. The comparison shows that the overshoot of the composite control algorithm is reduced by 4.81%, and the settling time is shortened by 38.89%, indicating that this method has superior dynamic response performance under step pressure conditions. Figure 6This is a step pressure control error curve diagram according to an embodiment of the present invention. The diagram reflects the change of tracking error over time under a step pressure command for both control algorithms. The error curves show that both algorithms generate instantaneous error spikes at the moment the step command is applied, but the peak error of the composite control algorithm is significantly lower than that of the PID algorithm. After entering the dynamic adjustment stage, the error curve of the PID algorithm exhibits multiple alternating positive and negative decaying oscillations, and the time required for the error to converge to near zero is relatively long. In contrast, the error curve of the composite control algorithm shows reduced oscillation amplitude and faster convergence speed, compressing the error to a near-zero steady-state range within approximately 0.22 seconds, with a narrower steady-state error band. This diagram further quantifies the dynamic advantages of the composite control algorithm under step pressure conditions from the perspective of error convergence.

[0151] like Figure 7 The figure shows the sinusoidal pressure control curve of an embodiment of the present invention. It compares the pressure tracking curves of the PID algorithm and the composite control algorithm of the present invention under a sinusoidal pressure command with an amplitude of 10 MPa and a frequency of 1 Hz. As can be seen from the figure, the PID control curve exhibits significant tracking lag and amplitude attenuation at the peak of the sinusoidal curve, making it difficult for the actual response to closely follow the changing trend of the target curve. In contrast, the response curve of the composite control algorithm closely matches the target curve, maintaining good amplitude tracking and phase consistency throughout the entire cycle, with only minor deviations at the peaks and troughs. This figure visually demonstrates the excellent tracking performance of the composite control algorithm for continuously changing signals. Figure 8 This is a graph showing the error curve of the sinusoidal pressure control in an embodiment of the present invention. The graph illustrates the variation of pressure error over time in the sinusoidal tracking process for the two control algorithms. As can be seen from the graph, the error curve of the PID algorithm exhibits obvious periodic fluctuations, with the error amplitude reaching its maximum near the command peak. Its maximum pressure error is 0.79 MPa, and the root mean square error is 0.45 MPa, indicating that the PID algorithm struggles to maintain high-frequency response accuracy in sinusoidal tracking. In contrast, the error curve of the composite control algorithm has a significantly smaller amplitude and smoother fluctuations, with a maximum pressure error of only 0.38 MPa and a root mean square error of 0.18 MPa. Compared to the PID algorithm, the composite control algorithm reduces the maximum pressure error by 52% and the root mean square error by 60%. The comparison of error waveforms fully verifies the high-precision advantage of this method in periodic time-varying pressure tracking.

[0152] The second aspect of this invention proposes a control system for an electric loader working device based on a composite pressure control method, such as... Figure 2 As shown, it specifically includes four major functional modules: a hydraulic power module, an energy recovery module, a composite power supply module, and a control module. Each module works together through hydraulic pipelines, electrical circuits, and control signals to adapt to the full range of working conditions of electric loaders, such as boom lifting and bucket extension and retraction, and to achieve integrated power supply, potential energy recovery, energy storage, and control.

[0153] The hydraulic power module provides active operating power for the pressure control model 29, and mainly includes an oil tank 1, a servo motor 2, a first coupling 3, a hydraulic pump 4, a first pressure sensor 5, a first relief valve 6, a first pressure compensation valve 13, a second pressure compensation valve 14, a first electromagnetic proportional directional valve 15, a second electromagnetic proportional directional valve 16, a first shuttle valve 17, a second shuttle valve 18, a third shuttle valve 19, a second pressure sensor 20, a boom hydraulic cylinder 25, and a bucket hydraulic cylinder 26. The servo motor 2 is rigidly connected to the hydraulic pump 4 via the first coupling 3, eliminating the power loss of the traditional transmission mechanism. The hydraulic pump 4 adopts a fixed displacement piston pump structure, with its suction port directly connected to the oil tank 1 to achieve stable oil suction. The oil outlet is divided into multiple paths. The first path is connected to the first pressure sensor 5 to collect the outlet pressure of the hydraulic pump 4 in real time and feed it back to the control module. The second path is connected to the first relief valve 6, and the return port of the first relief valve 6 is connected to the oil tank 1. When the pressure of the pressure control model 29 exceeds the set threshold, it overflows to relieve pressure and protect the hydraulic pump 4 and subsequent components from overload damage. The third path is connected to the first pressure compensation valve 13. The fourth path is connected to the inlet of the second pressure compensation valve 14. The outlets of the first pressure compensation valve 13 and the second pressure compensation valve 14 are respectively connected to the inlet P of the first electromagnetic proportional directional valve 15 and the second electromagnetic proportional directional valve 16. The return ports T of the two electromagnetic proportional directional valves are both connected to the oil tank 1, forming a complete return oil circuit. The two inlets of the first shuttle valve 17 are connected to the two working ports of the first electromagnetic proportional directional valve 15, and the two inlets of the second shuttle valve 18 are connected to the two working ports of the second electromagnetic proportional directional valve 16. The outlets of the first shuttle valve 17 and the second shuttle valve 18 are connected to the control ports of the first pressure compensation valve 13 and the second pressure compensation valve 14, respectively, to maintain a constant pressure difference between the two electromagnetic proportional directional valves through load pressure feedback, thus achieving electronically controlled load-sensitive control. On the other hand, they are both connected to the inlet of the third shuttle valve 19, and the outlet of the third shuttle valve 19 is connected to the second pressure sensor 20 to collect and feedback the maximum load pressure of the pressure control model 29 in real time, providing pressure closed-loop basis for the control module. The rod-side and rodless-side chambers of the boom hydraulic cylinder 25 and the bucket hydraulic cylinder 26 are respectively connected to the working ports A and B of the first electromagnetic proportional directional valve 15 and the second electromagnetic proportional directional valve 16. Through the switching and opening adjustment of the electromagnetic proportional directional valves, the boom lifting and lowering, as well as the bucket retraction and tipping actions are driven.

[0154] The energy recovery module is used to recover gravitational potential energy during boom and bucket tipping operations. It mainly includes a generator 7, a second coupling 8, a hydraulic motor 9, a first check valve 10, a second check valve 11, a second relief valve 12, a third check valve 21, a fourth check valve 22, a first switching valve 23, and a second switching valve 24. The hydraulic motor 9 is a fixed displacement motor, coaxially connected to the generator 7 via the second coupling 8. The rotation of the hydraulic motor 9 directly drives the generator 7 to generate electricity, converting hydraulic energy into electrical energy. The hydraulic motor 9 has multiple inlet ports. The first port is connected to the outlet ports of the first switching valve 23 and the second switching valve 24. The second port is connected to the second relief valve 12. The third port is connected to the second check valve 11. The inlet ports of the first switching valve 23 and the second switching valve 24 are respectively connected to the rodless chambers of the boom hydraulic cylinder 25 and the bucket hydraulic cylinder 26 to conduct the high-pressure oil discharged from the rodless chambers of the hydraulic cylinders when the boom drops and the bucket tips. The return port of the second relief valve 12 is connected to the oil tank 1 to limit the inlet pressure of the hydraulic motor 9 and prevent high-pressure impact from damaging the motor and the generator 7. The inlet port of the second check valve 11 is connected to the oil tank 1, and the outlet port is connected to the inlet port of the hydraulic motor 9. When the inlet pressure of the hydraulic motor 9 is low, it automatically replenishes oil to prevent cavitation. The hydraulic motor 9 has three outlet ports: the first port is connected to the first check valve 10, the second port is connected to the inlet of the third check valve 21, and the third port is connected to the inlet of the fourth check valve 22. The outlet of the first check valve 10 is connected to the oil tank 1 to provide stable back pressure for the hydraulic motor 9, making the power generation process smooth. The outlets of the third check valve 21 and the fourth check valve 22 are respectively connected to the rod chambers of the boom hydraulic cylinder 25 and the bucket hydraulic cylinder 26, so that the oil discharged from the hydraulic motor 9 is preferentially replenished into the rod chamber of the hydraulic cylinder, the hydraulic oil is recycled, the amount of oil replenished to the oil tank 1 is reduced, and the efficiency of the hydraulic system is improved.

[0155] The composite power module adopts a composite energy storage architecture of lithium-ion battery 33 and supercapacitor 35, balancing high energy density and high power density. It mainly includes lithium-ion battery 33, battery management system 32, supercapacitor 35, capacitor management system 34, and DC bus 36. Lithium-ion battery 33 serves as the main energy storage module, with high energy density to meet the continuous power supply requirements of long-term operation. Supercapacitor 35 serves as the auxiliary energy storage module, with high power density and fast charging and discharging speed, suitable for the high-frequency, high-impact potential energy recovery scenarios of loaders. Lithium-ion battery 33 is connected to DC bus 36 through battery management system 32, and supercapacitor 35 is connected to DC bus 36 through capacitor management system 34. Battery management system 32 and capacitor management system 34 respectively have functions of status monitoring, charge and discharge control, overcharge and over-discharge protection, and equalization management, realizing power on / off and power distribution according to control module commands. During energy-consuming operations such as lifting the boom and retracting the bucket, the battery management system 32 and the capacitor management system 34 are simultaneously turned on, and the lithium-ion battery 33 and the supercapacitor 35 work together to power the servo motor 2. During recovery operations such as lowering the boom and tipping the bucket, the capacitor management system 34 is turned on first, and the recovered electrical energy is quickly stored in the supercapacitor 35 to avoid frequent charging and discharging from damaging the lithium-ion battery 33 and extend the battery life.

[0156] The control module is the core of the pressure control model 29 for decision-making and execution, enabling full-condition control. It mainly includes the boom handle 27, bucket handle 28, pressure control model 29, motor driver 30, and generator driver 31. The boom handle 27 and bucket handle 28 are operation input elements, with their signal outputs directly connected to the pressure control model 29, converting the operator's intentions into electrical signals. The signal outputs of the first pressure sensor 5 and the second pressure sensor 20 are also connected to the pressure control model 29, providing real-time feedback on the outlet pressure of the hydraulic pump 4 and the maximum load pressure of the pressure control model 29. The pressure control model 29 adopts a dedicated engineering machinery control model, which has signal acquisition, logic judgment, and calculation output functions. Its control output terminal is divided into multiple channels. The first channel is connected to the first coil YA1 and the second coil YA2 of the first electromagnetic proportional directional valve 15 and the first coil YA3 and the second coil YA4 of the second electromagnetic proportional directional valve 16, to adjust the opening degree and directional direction of the directional valve. The second channel is connected to the coil DT1 of the first switching valve 23 and the coil DT2 of the second switching valve 24, to control the on and off of the switching valve. The third channel is connected to the motor driver 30. The fourth channel is connected to the generator driver 31. The fifth channel is connected to the battery management system 32. The sixth channel is connected to the capacitor management system 34. One end of the motor driver 30 is connected to the DC bus 36, and the other end of the motor driver 30 is connected to the servo motor 2. The speed of the servo motor 2 is adjusted according to the instructions of the pressure control model 29. One end of the generator driver 31 is connected to the DC bus 36, and the other end of the generator driver 31 is connected to the generator 7. The generator driver 31 realizes power generation rectification, voltage stabilization and energy feedback control, so that the recovered electrical energy is stably stored in the supercapacitor 35.

[0157] like Figure 3 As shown, the operation flow of the pressure control model 29 of the present invention is as follows: When the pressure control model 29 is running, it automatically identifies the working conditions and executes differentiated control strategies based on the action signals of the boom handle 27 and the bucket handle 28 and the feedback signals of the pressure sensor. The specific operation process is as follows:

[0158] In standby mode, there is no operation signal input to the boom handle 27 and bucket handle 28. The pressure control model 29 controls the coils YA1, YA2, YA3, and YA4 of the first electromagnetic proportional directional valve 15 and the second electromagnetic proportional directional valve 16, as well as the coils DT1 and DT2 of the first switching valve 23 and the second switching valve 24 to be de-energized, and all valves are in the closed state. At the same time, the battery management system 32 and the capacitor management system 34 are turned off, the DC bus 36 stops supplying power, and the servo motor 2 and the generator 7 both have a speed of 0. The system is in standby mode with no energy consumption, reducing the standby power consumption of the whole vehicle.

[0159] In the case of raising the boom or lowering the bucket alone, the pressure control model 29 detects the boom or bucket handle signal and controls the first switching valve 23 and the second switching valve 24 to close, cutting off the energy recovery circuit. Simultaneously, it controls the first electromagnetic proportional directional valve 15 or the second electromagnetic proportional directional valve 16 to be energized and at maximum opening, eliminating the throttling channel. The pressure control model 29 calculates and outputs a speed command based on the handle opening ratio, adjusting the speed of the servo motor 2 via the motor driver 30. The servo motor 2 drives the hydraulic pump 4 to output pressurized oil of corresponding flow rate, directly actuating the boom hydraulic cylinder 25 or the bucket hydraulic cylinder 26, achieving independent pump-controlled speed regulation. This completely eliminates the throttling losses of traditional valve control systems and improves the energy utilization rate of active operation. At this time, the battery management system 32 and the capacitor management system 34 are activated, and the composite power supply provides stable power to the servo motor 2.

[0160] In stand-alone boom or stand-alone bucket operation, the pressure control model 29 detects the boom or bucket handle signal and de-energizes and closes the first electromagnetic proportional directional valve 15 and the second electromagnetic proportional directional valve 16, cutting off the active power circuit. Simultaneously, it energizes and opens the first switching valve 23 or the second switching valve 24, connecting the energy recovery circuit. The boom or bucket descends under its own weight, pushing the rodless chamber of the hydraulic cylinder to discharge high-pressure oil. The high-pressure oil enters the hydraulic motor 9 through the switching valve, driving the hydraulic motor 9 to rotate, which in turn drives the generator 7 to generate electricity. The pressure control model 29 adjusts the generator 7 speed according to the handle opening ratio to achieve stable control of the descent speed. The recovered electrical energy is rectified and stabilized by the generator driver 31, and then stored in the supercapacitor 35 through the capacitor management system 34, completing the conversion and storage of gravitational potential energy into electrical energy. The oil discharged from the hydraulic motor 9 is preferentially replenished into the rod chamber of the hydraulic cylinder through a check valve, while excess oil flows back to the oil tank 1 for hydraulic oil recycling.

[0161] The composite action conditions are divided into four scenarios. The pressure control model 29 adopts a collaborative control strategy: In the boom-lifting and bucket-retracting composite action, both electromagnetic proportional directional valves are energized. The pressure control model 29 adjusts the valve opening according to the handle opening. Through pressure closed-loop control, the outlet pressure of the hydraulic pump 4 follows the maximum load pressure. The speed of the servo motor 2 is adjusted by PID closed-loop regulation to achieve on-demand flow distribution and coordinated action without jamming. The composite power supply provides power in tandem. In the boom-lifting and bucket-lowering composite action, the active action is powered by an independent pump control. The recovery action is activated simultaneously to recover potential energy. The battery management system 32 and the capacitor management system 34 work together to provide power and recover energy at the same time. In the boom-lowering and bucket-lowering composite action, both switching valves are energized and opened. The high-pressure oil in the rodless chamber of the boom hydraulic cylinder 25 and the bucket hydraulic cylinder 26 flows into the hydraulic motor 9. The dual potential energy superposition drives the generator 7 to generate electricity efficiently. All energy is stored in the supercapacitor 35.

[0162] The embodiments of this invention, through the integrated design of four major modules and the hierarchical control strategy for all working conditions, completely solve the problems of high energy consumption, poor potential energy recovery, and weak controllability of hydraulic systems in traditional electric loaders. Active operation has no throttling loss, passive operation has efficient potential energy recovery, and composite actions are coordinated. At the same time, the composite power supply optimizes energy management, improves the range performance, operating efficiency and economy of electric loaders, and fully meets the requirements of high efficiency, energy saving and stable operation of new energy construction machinery.

[0163] The beneficial effects of this invention are as follows: This invention proposes a control method and system for an electric loader working device based on composite pressure control, which can adapt to all tasks and working conditions of the electric loader working device, solving the limitation of existing technologies that only adapt to single task conditions; by using independent pump control speed regulation for separate lifting and bucket retraction, energy loss is completely eliminated; gravity potential energy recovery is used for boom lowering and bucket tipping, enabling oil reuse and improving energy utilization; the control of composite actions improves the sensitivity of the electronically controlled load for boom lifting and bucket retraction, achieving on-demand flow distribution and good coordination of control actions; collaborative control is adopted in composite working conditions, balancing operability and energy saving; this invention uses a combination of lithium-ion batteries and supercapacitors to optimize power management, using supercapacitors to quickly store and recover electrical energy, improving energy recovery efficiency and extending battery life.

[0164] The embodiments described above are merely preferred embodiments of the present invention and are not intended to limit the scope of the present invention. Various modifications and improvements made by those skilled in the art to the technical solutions of the present invention without departing from the spirit of the present invention should fall within the protection scope defined by the claims of the present invention.

Claims

1. A control method for the working device of an electric loader based on composite pressure control, characterized in that: It includes: S1: Based on the electromagnetic torque of the servo motor With load torque To establish the dynamic equation of the servo motor of the electric loader based on the balance relationship between them; and based on the actual output flow of the hydraulic pump... and hydraulic pump outlet pressure The dynamic equations of the hydraulic pump in the working device of the electric loader are established; the load pressure control equations of the working device of the electric loader are established; and the control model of the working device of the electric loader is obtained by combining the above. S2: Establish a hydraulic pump composite pressure control model that integrates nonlinear eigenmap feedforward compensation, real-time disturbance estimation of the extended state observation model, and sliding mode variable structure control to perform hydraulic pump composite pressure control for the boom lifting and bucket retraction actions of the electric loader's working device; the first-level hydraulic pump composite pressure control model is based on the preset hydraulic pump outlet pressure. and current load traffic Q load To balance flow rate and torque, feedforward compensation control is implemented; the second-layer hydraulic pump composite pressure control model sets the observed estimates of the disturbance extended state variables. As state variables, determine the extended state-space equations of the pressure control model and perform extended state observations; the third-layer hydraulic pump composite pressure control model uses the hydraulic pump outlet pressure as the state variable. Define the sliding surface as the controlled variable. Deriving the sliding mode control law Perform sliding mode variable structure control; S3: Observed estimates of the perturbation extended state variables estimated by the extended state observation model. By introducing feedforward compensation into the control law, a composite control is formed, and a new sliding mode control law is output. The total output of the composite pressure control model is obtained as the electromagnetic torque control parameter u of the servo motor, which is used for the torque control of the servo motor.

2. The control method for the working device of an electric loader based on composite pressure control according to claim 1, characterized in that: S1 specifically refers to: S11: The speed response of the servo motor shaft is affected by the balance between the electromagnetic torque of the servo motor and the load torque. The dynamic equation of the servo motor in the working device of the electric loader is established. S12: Determine the actual output flow rate of the hydraulic pump based on the nonlinear relationship between rotational speed and load pressure; determine the outlet pressure of the hydraulic pump based on the principles of oil continuity and compressibility; determine the dynamic equation of the hydraulic pump in the working device of the electric loader. S13: Establish the state-space equations of the pressure control model and obtain the second-order state-space model by combining them; establish the load pressure control equations in the working device of the electric loader.

3. The control method for the working device of an electric loader based on composite pressure control according to claim 2, characterized in that: The load pressure control equation in S13 is described by a state-space equation in affine nonlinear form, specifically: ； in, The outlet pressure of the hydraulic pump is a state variable. Its first derivative; For the servo motor angular velocity state variable, Its first derivative; This is a nonlinear function in pressure dynamics; This refers to the electromagnetic torque of the servo motor. For the electromagnetic torque of the servo motor The coefficient of influence on the rate of change of pressure; This refers to the lumped disturbance in the pressure loop; It is a nonlinear function in angular velocity dynamics; For the electromagnetic torque of the servo motor The coefficient of influence on the rate of change of angular velocity; This refers to the lumped disturbance in the speed loop; This is a time parameter.

4. The control method for the working device of an electric loader based on composite pressure control according to claim 1, characterized in that: S2 specifically refers to: S21: Based on the preset hydraulic pump outlet pressure and current load traffic Q load To achieve flow balance and torque balance under steady-state operating conditions of the system, a feedforward compensation control model is established as the first-level hydraulic pump composite pressure control model. S22: Let the observed estimates of the perturbation extended state variables be... For state variables, define the perturbation extended state variable. , To determine the total disturbance of the pressure control model, we define the state-space equation of the extended pressure control model and construct the extended state observation model as the second-layer hydraulic pump composite pressure control model. S23: Based on the hydraulic pump outlet pressure Define the sliding surface as the controlled variable. ; Taking the exponential reaching law as the reaching law, the sliding mode control law is derived. A sliding mode variable structure control model was established as a third-layer hydraulic pump composite pressure control model.

5. The control method for the working device of an electric loader based on composite pressure control according to claim 4, characterized in that: The extended state-space equations of the pressure control model in S22 are: ； in, For perturbation, extend the state variables. Its first derivative; These are the electromagnetic torque control parameters for the servo motor. This is the initial nonlinear nominal function; This is the nominal value of the control gain; Total disturbance of the pressure control model The first derivative; For the electromagnetic torque of the servo motor The influence coefficient; This refers to the displacement of the hydraulic cylinder.

6. The control method for the working device of an electric loader based on composite pressure control according to claim 4, characterized in that: The observation model for the extended state in S22 is: ； in, The hydraulic pump outlet pressure state variable The observed estimates, Its first derivative; This is the observed estimate of the total disturbance d(t) for the pressure control model. Its first derivative; For the observed estimates of the perturbation extended state variables, Its first derivative; The gain of the first observation model; For the second observation model gain; For the third observation model gain; This is the estimated displacement value of the hydraulic cylinder.

7. The control method for the working device of an electric loader based on composite pressure control according to claim 4, characterized in that: In S23, the sliding surface is set as follows: ； ； in, It is a sliding surface; λ represents the pressure tracking error; λ is the integral gain parameter. For time variables Pressure tracking error during the process; It is a time variable; This refers to the outlet pressure of the hydraulic pump. Preset the hydraulic pump outlet pressure; This is a time parameter.

8. The control method for the working device of an electric loader based on composite pressure control according to claim 1, characterized in that: Servo motor electromagnetic torque control parameters in S3 for: ； ； in, This is the nominal value of the control gain; The electromagnetic torque of the servo motor is output as a feedforward compensation. This is a sliding mode control law; This is the equivalent control portion; For switching control section; These are the electromagnetic torque control parameters for the servo motor.

9. A control system for an electric loader working device based on composite pressure control, wherein the electric loader working device control system is used in the control method for an electric loader working device based on composite pressure control as described in any one of claims 1 to 8, and the electric loader working device control system specifically includes: The hydraulic power module, energy recovery module, composite power module and control module are characterized in that: the energy recovery module recovers the gravitational potential energy under the conditions of boom drop and bucket tipping. The hydraulic power module provides active operating power for the pressure control model, including an oil tank, servo motor, first coupling, hydraulic pump, first pressure sensor, first relief valve, first pressure compensation valve, second pressure compensation valve, first electromagnetic proportional directional valve, second electromagnetic proportional directional valve, first shuttle valve, second shuttle valve, third shuttle valve, second pressure sensor, boom hydraulic cylinder and bucket hydraulic cylinder. The energy recovery module includes a generator, a second coupling, a hydraulic motor, a first check valve, a second check valve, a second relief valve, a third check valve, a fourth check valve, a first switching valve, and a second switching valve. The composite power module includes a lithium-ion battery, a battery management system, a supercapacitor, a capacitor management system, and a DC bus. It uses a composite energy storage architecture composed of lithium-ion batteries and supercapacitors to recover gravitational potential energy under the conditions of falling boom and tipping bucket operation. The control module includes a boom handle, a bucket handle, a pressure control model, a motor driver, and a generator driver.

10. The control system for the electric loader working device based on composite pressure control according to claim 9, characterized in that: The hydraulic motor in the energy recovery module is a fixed-displacement motor, which is coaxially connected to the generator through a second coupling. The rotation of the hydraulic motor directly drives the generator to generate electricity, converting hydraulic energy into electrical energy. The hydraulic motor has three inlet ports: the first port connects to the outlet ports of the first and second switching valves; the second port connects to the second relief valve; and the third port connects to the second check valve. The inlets of the first and second switching valves are connected to the rodless chambers of the boom hydraulic cylinder and the bucket hydraulic cylinder, respectively, to facilitate the flow of high-pressure oil discharged from the rodless chambers of the hydraulic cylinders during boom lowering and bucket tipping. The return port of the second relief valve connects to the oil tank to limit the inlet pressure of the hydraulic motor. The inlet port of the second check valve connects to the oil tank, and the outlet port is connected to the inlet port of the hydraulic motor, automatically replenishing oil when the inlet pressure of the hydraulic motor is lower than the opening pressure of the second check valve. The hydraulic motor has three outlet ports: the first port is connected to the first check valve, the second port is connected to the inlet of the third check valve, and the third port is connected to the inlet of the fourth check valve. The outlet of the first check valve is connected to the oil tank to provide stable back pressure for the hydraulic motor. The outlets of the third and fourth check valves are connected to the rod chambers of the boom hydraulic cylinder and the bucket hydraulic cylinder, respectively, so that the oil discharged from the hydraulic motor can flow into the rod chamber of the hydraulic cylinder.

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