Electrically-driven metering pump load-sensitive system, control method and engineering machinery

Abstract

The invention relates to the technical field of engineering machinery, and particularly discloses an electrically-driven metering pump load-sensitive system, a control method and engineering machinery, the electrically-driven metering pump load-sensitive system comprises a control unit, a driving motor assembly, a metering pump, a pilot hydraulic control unit, a pilot handle assembly, a load-sensitive valve assembly, an actuator assembly, a sensor assembly and a pressure difference control module; the driving motor drives the metering pump to supply oil to the load-sensitive valve assembly so as to drive the actuator assembly to operate; the control unit receives the maximum outlet pressure and the maximum load pressure of the pump collected by the pressure sensor, receives a system target pressure difference signal of the pressure difference control module at the same time, and synchronously controls the rotating speed of the motor and the valve target pressure difference output by the proportional pressure reducing valve. The system aims to solve the problems that a traditional upstream compensation load sensitive system lacks a flow saturation resisting function, and a downstream compensation load sensitive system has extra reversing energy consumption and fixed system pressure difference setting.

Description

Technical Field

[0001] The present invention relates to the technical field of engineering machinery, and in particular to a load-sensing system for an electric-driven quantitative pump, a control method, and engineering machinery. Background Art

[0002] In the field of construction machinery, load-sensitive systems are widely used because their pressure compensation characteristics can ensure that the control performance is not affected by the load. However, the electrification and intelligent development of construction machinery have put forward higher energy-saving and controllability requirements for load-sensitive systems.

[0003] Existing load-sensitive systems can be divided into pre-valve compensation (LS) systems and post-valve compensation (LUDV) systems. In the pre-valve compensation system, the actuator flow distribution ratio is unbalanced when the flow is saturated. In the post-valve compensation system, the throttling section and the reversing section of the main valve core are separated, resulting in additional reversing throttling losses, resulting in lower energy efficiency than the pre-valve compensation system. In addition, the pressure difference of the two systems is fixed by the pressure-regulating spring, which not only causes a large throttling loss at the valve port, but also makes it difficult to take into account the flow control requirements of different working conditions such as fine tuning and fast operation.

[0004] Existing improvement solutions or structural changes for flow saturation in pre-valve compensation systems are complex or rely on precise calculations and are susceptible to interference. Variable pressure differential control cannot be applied to pre-valve compensation systems due to structural differences. The traditional power source constant speed-variable displacement method has energy efficiency fluctuations under drastic load changes and the motor and pump responses do not match. Variable speed-variable displacement hybrid control also has similar problems, and the hydraulic pump structure is complex and costly. Therefore, a new system is needed to solve the above problems. Summary of the Invention

[0005] In view of this, the purpose of the present invention is to provide an electric-driven quantitative pump load-sensing system, a control method and engineering machinery thereof, which can effectively solve the problems of the existing load-sensing system's pre-valve compensation being unable to resist flow saturation, the post-valve compensation having additional switching energy consumption and the fixed setting of the system pressure difference.

[0006] The present invention provides an electric-driven quantitative pump load-sensing system, comprising: a control unit, a drive motor assembly, a quantitative pump, a pilot hydraulic control unit, a pilot handle assembly, a load-sensing valve assembly, an actuator assembly, a sensor assembly, and a pressure differential control module; The output end of the control unit is connected to the drive motor assembly and drives the metering pump to operate. The metering pump outputs high-pressure oil to the load-sensing valve assembly. The load-sensing valve assembly is configured to regulate the flow entering the actuator assembly to drive the actuator assembly to operate; the sensor assembly is connected to each valve in the load-sensing valve assembly and measures its pressure signal. The pressure differential control module derives the pressure differential signal between the front and rear ends of each valve based on the pressure signal obtained by the sensor assembly. The input end of the control unit is connected to the sensor assembly and the pressure differential control module. The control unit regulates the valve target pressure differential through the load-sensing valve assembly based on the pressure signal and the pressure differential signal between the front and rear ends of each valve, thereby controlling the target speed of the drive motor assembly and controlling the control pressure of the pilot hydraulic control unit through the pilot handle assembly.

[0007] Preferably, the valves in the load-sensitive valve assembly include a differential relief valve, a first pressure-compensating valve, a second pressure-compensating valve, a first three-position four-way reversing valve, a second three-position four-way reversing valve and a shuttle valve. The differential relief valve is arranged at the oil inlet of the load-sensitive valve assembly, the first pressure-compensating valve is arranged at the oil inlet of the first three-position four-way reversing valve, the second pressure-compensating valve is arranged at the oil inlet of the second three-position four-way reversing valve, and the shuttle valve is arranged at the oil outlet of the three-position four-way reversing valve. The pilot control ends of the first three-position four-way reversing valve and the second three-position four-way reversing valve are connected to the output end of the pilot handle assembly.

[0008] Preferably, one end of the valve core of the first pressure compensating valve and the second pressure compensating valve compensates for the front-end pressure of the three-position four-way reversing valve, and the other end compensates for the load pressure and target control pressure at the rear end of the three-position four-way reversing valve. The target control pressure oil circuits of the first pressure compensating valve and the second pressure compensating valve are connected in parallel and connected to the output end of the pilot hydraulic control unit.

[0009] Preferably, the first pressure compensating valve and the second pressure compensating valve are configured to respectively maintain the front and rear end pressure difference of the first three-position four-way reversing valve and the front and rear end pressure difference of the second three-position four-way reversing valve at the valve target pressure difference of the pressure differential controllable load sensitive valve regulated by the proportional pressure reducing valve in the pilot hydraulic control unit; the first three-position four-way reversing valve and the second three-position four-way reversing valve are configured to control the operating speed of the actuator assembly by adjusting the valve port opening; the fixed differential relief valve is configured to prevent system overpressure; and the shuttle valve is configured to obtain the maximum load pressure of the actuator drive chamber of each branch.

[0010] Preferably, the actuator assembly includes a first actuator and a second actuator, the first actuator is connected to the output port of the first three-position four-way reversing valve, and the second actuator is connected to the output port of the second three-position four-way reversing valve; the first actuator and the second actuator are hydraulic cylinders or hydraulic motors.

[0011] Preferably, the sensor assembly includes a first pressure sensor and a second pressure sensor, the first pressure sensor is arranged on the load-sensing valve assembly and is configured to detect the maximum load pressure of the system, and the second pressure sensor is arranged on the connecting pipeline between the metering pump and the load-sensing valve assembly and is configured to detect the outlet pressure of the metering pump.

[0012] Preferably, the pilot handle assembly includes a first pilot handle and a second pilot handle, the input ends of the first pilot handle and the second pilot handle are connected to the pilot hydraulic control unit, and the output ends of the first pilot handle and the second pilot handle are connected to the main valve core opening control end in the load-sensing valve assembly, and the load-sensing valve assembly is configured to control the target operating speeds of the first actuator and the second actuator according to the output signals of the first pilot handle and the second pilot handle, respectively.

[0013] Preferably, the pressure difference control module is configured to provide a system target pressure difference.

[0014] A control method for the load-sensing system of the electric-driven quantitative pump as described above comprises the following steps: Step S1: obtaining a system target pressure difference signal set by a pressure difference control module; Step S2: obtaining a maximum load pressure signal detected by the first pressure sensor in the sensor assembly and a pump outlet pressure signal detected by the second pressure sensor, and subtracting the maximum load pressure from the pump outlet pressure to obtain an actual system pressure difference; Step S3: setting the valve target pressure difference to the actual system pressure difference, controlling the proportional pressure reducing valve output valve target pressure difference in the pilot hydraulic control unit assembly to the first pressure compensation valve and the second pressure compensation valve in the load sensing valve assembly, thereby controlling the front and rear end pressure differences of the first three-position four-way reversing valve and the second three-position four-way reversing valve to be equal to the current actual system pressure difference; Step S4: With the system target pressure difference as the target value and the system actual pressure difference as the feedback value, closed-loop control is performed through the variable speed control of the drive motor. The closed-loop controller outputs the target speed signal, which is limited and sent to the motor controller to control the motor speed to maintain the system actual pressure difference stable at the system target pressure difference; return to step S1 to perform the next round of control cycle.

[0015] An engineering machine is provided with the above-mentioned electric-driven quantitative pump load sensing system.

[0016] By adopting the above-mentioned technical solution, the present invention achieves the following technical effects: The present invention provides an electric-driven constant-flow pump load-sensing system, control method, and engineering machinery. During normal system operation, the motor speed changes in real time to maintain the system's actual pressure differential at the system's target pressure differential. When the system reaches flow saturation, the motor speed reaches maximum, and the system's actual pressure differential decreases, causing the valve's target pressure differential to decrease simultaneously, resulting in a decrease in the flow demand for each throttle's current flow area. When the total flow demand drops to equal the pump's output flow, the system's actual pressure differential and the valve's target pressure differential stop decreasing and remain constant, reaching a new equilibrium state. During this process, the pump's target pressure differential is the system's target pressure differential, consistently exceeding the system's actual pressure differential, thus maintaining the motor speed at its maximum value. Furthermore, when the flow area of ​​each main valve core decreases, the system's actual pressure differential increases, causing the valve's target pressure differential to increase simultaneously. When the system's actual pressure differential reaches the target pressure differential, the motor speed begins to decrease, maintaining the system's actual pressure differential at the target pressure differential. At this point, the system exits flow saturation. Therefore, by actively regulating the valve's target pressure differential, the system can maintain the flow distribution ratio between the light-load and heavy-load links even when flow is saturated, unaffected by load differences. This achieves the flow saturation resistance of pre-valve compensation. Compared with existing pre-valve compensation load-sensitive systems, this effectively avoids actuator speed imbalance under flow saturation conditions, significantly improving controllability. Furthermore, compared with existing post-valve compensation load-sensitive systems with flow saturation resistance, this system can operate at different system pressure differentials to achieve variable valve port flow gain, avoiding system flow saturation. This system has a wider effective control stroke, significantly improving fine-tuning and flow control characteristics. Furthermore, the system utilizes a pre-valve compensation structure, eliminating additional switching throttling losses and minimizing valve port throttling losses. BRIEF DESCRIPTION OF THE DRAWINGS

[0017] In order to more clearly illustrate the technical solutions of the embodiments of the present invention, the following is a brief introduction to the drawings required for use in the embodiments. It should be understood that the following drawings only illustrate certain embodiments of the present invention and therefore should not be regarded as limiting the scope. For those skilled in the art, other relevant drawings can be obtained based on these drawings without creative work.

[0018] Figure 1 It is a structural diagram of a load-sensing system for an electrically driven metering pump provided by the first embodiment of the present invention.

[0019] Figure 2 1 is a flow chart of a method for controlling a load-sensing system of an electrically driven metering pump provided by the second embodiment of the present invention.

[0020] In the figure: 1-control unit, 2-drive motor assembly, 21-power supply, 22-motor driver, 23-drive motor, 3-dosing pump, 4-pilot hydraulic control unit, 41-pilot pump, 42-pilot relief valve, 43-proportional pressure reducing valve, 5-pilot handle assembly, 51-first pilot handle, 52-second pilot handle, 6-load sensing valve assembly, 61-differential relief valve, 62-first pressure compensating valve, 63-second pressure compensating valve, 64-first three-position four-way reversing valve, 65-second three-position four-way reversing valve, 66-shuttle valve, 7-actuator assembly, 71-first actuator, 72-second actuator, 8-sensor assembly, 81-first pressure sensor, 82-second pressure sensor, 9-differential pressure control module. DETAILED DESCRIPTION

[0021] In order to make the purpose, technical solutions and advantages of the embodiments of the present invention clearer, the technical solutions in the embodiments of the present invention will be clearly and completely described below in conjunction with the drawings in the embodiments of the present invention. Obviously, the described embodiments are part of the embodiments of the present invention, not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by the first technicians in this field without making creative work are within the scope of protection of the present invention. Therefore, the following detailed description of the embodiments of the present invention provided in the drawings is not intended to limit the scope of the invention for which protection is sought, but merely represents selected embodiments of the present invention. Based on the embodiments of the present invention, all other embodiments obtained by the first technicians in this field without making creative work are within the scope of protection of the present invention.

[0022] Example The following are only preferred embodiments of the present invention. The protection scope of the present invention is not limited to the following embodiments. All technical solutions under the concept of the present invention belong to the protection scope of the present invention.

[0023] Reference Manual Figure 1The first embodiment of the present invention provides an electric-driven quantitative pump load-sensing system, comprising: a control unit 1, a drive motor assembly 2, a quantitative pump 3, a pilot hydraulic control unit 4, a pilot handle assembly 5, a load-sensing valve assembly 6, an actuator assembly 7, a sensor assembly 8, and a pressure differential control module 9; wherein the drive motor assembly 2 drives the quantitative pump 3 to operate, and the quantitative pump 3 outputs high-pressure oil to the load-sensing valve assembly 6, and the load-sensing valve assembly 6 is configured to regulate the flow entering the actuator assembly 7 to drive the actuator assembly 7 to operate at a certain speed; the control unit 1 collects signals from the sensor assembly 8 and the pressure differential control module 9, controls the target speed and The control pressure of the pilot hydraulic control unit 4; the drive motor assembly 2 includes a power supply 21, a motor driver 22 and a drive motor 23, the power supply 21 is electrically connected to the input end of the motor driver 22, the output end of the motor driver 22 is electrically connected to the drive motor 23, and the output end of the drive motor 23 is coaxially mechanically connected to the metering pump 3; the motor driver is configured to receive the target speed signal sent by the control unit, and control the drive motor to run at a speed corresponding to the target speed signal; the motor driver 22 is configured to receive the target speed signal sent by the control unit 1, and control the drive motor 23 to run at a speed corresponding to the target speed signal.

[0024] In a possible embodiment of the present invention, the load-sensing valve assembly 6 is a pressure-differential controllable load-sensing valve assembly, specifically including a fixed-differential relief valve 61, a first pressure-compensating valve 62, a second pressure-compensating valve 63, a first three-position four-way reversing valve 64, a second three-position four-way reversing valve 65 and a shuttle valve 66. The fixed-differential relief valve 61 is arranged at the oil inlet of the load-sensing valve assembly 6, the first pressure-compensating valve 62 is arranged at the oil inlet of the first three-position four-way reversing valve 64, the second pressure-compensating valve 63 is arranged at the oil inlet of the second three-position four-way reversing valve 65, and the shuttle valve 66 is arranged at the oil outlets of the first three-position four-way reversing valve 64 and the second three-position four-way reversing valve 65. The input end of the first three-position four-way reversing valve 64 and the input end of the second three-position four-way reversing valve 65 are connected to the output end of the pilot handle assembly 5; one end of the valve core of the first pressure-compensating valve 62 and the second pressure-compensating valve 63 act on the first three-position four-way reversing valve 64, the second three-position four-way reversing valve 65, and the shuttle valve 66 act on the first three-position four-way reversing valve 64 and the second three-position four-way reversing valve 65, respectively. The first and second pressure compensating valves 62 and 63 are connected in parallel to the target control pressure circuits of the first and second pressure compensating valves 62 and 63 and are connected to the target pressure differential output by the pilot hydraulic control unit 4. The first and second pressure compensating valves 62 and 63 are configured to maintain the pressure differential between the front and rear ends of the first and second pressure compensating valves 64 and 65, respectively, at the target pressure differential of the pressure-differential load-sensitive valve controlled by the proportional pressure reducing valve 43 in the pilot hydraulic control unit 4. The first and second pressure compensating valves 64 and 65 are configured to control the operating speed of the actuator assembly 7 by adjusting the valve opening. The fixed differential relief valve 61 is configured to prevent system overpressure. The shuttle valve 66 is configured to obtain the maximum load pressure of the actuator drive chamber of each branch.

[0025] In a possible embodiment of the present invention, the pilot hydraulic control unit 4 includes a pilot pump 41, a pilot relief valve 42 and a proportional pressure reducing valve 43. The pilot relief valve 42 is arranged at the outlet of the pilot pump 41 to adjust the primary pilot pressure, and the primary pilot pressure is supplied to the input end of the proportional pressure reducing valve 43 and the pilot handle assembly 5 respectively; the input end of the proportional pressure reducing valve 43 is connected to the output end of the control unit 1 to realize real-time adjustment of the valve target pressure difference according to the actual pressure difference of the system, and the output end of the proportional pressure reducing valve 43 is connected to the load-sensitive valve assembly 6 to realize providing the load-sensitive valve assembly 6 with a set pressure of the valve target pressure difference.

[0026] In one possible embodiment of the present invention, the actuator assembly 7 includes a first actuator 71 and a second actuator 72. The first actuator 71 is connected to the output port of the first three-position four-way reversing valve 64, and the second actuator 72 is connected to the output port of the second three-position four-way reversing valve 65. The first actuator 71 and the second actuator 72 are hydraulic cylinders or hydraulic motors. In a possible embodiment of the present invention, the sensor assembly 8 includes a first pressure sensor 81 and a second pressure sensor 82. The first pressure sensor 81 is configured on the load-sensing valve assembly 6, and the first pressure sensor 81 is configured to detect the maximum load pressure of the system. The second pressure sensor 82 is configured on the connecting pipeline between the metering pump 3 and the load-sensing valve assembly 6, and the second pressure sensor 82 is configured to detect the outlet pressure of the metering pump 3.

[0027] In a possible embodiment of the present invention, the pilot handle assembly 5 includes a first pilot handle 51 and a second pilot handle 52, the input ends of the first pilot handle 51 and the second pilot handle 52 are connected to the pilot hydraulic control unit 4, and the output ends of the first pilot handle 51 and the second pilot handle 52 are connected to the main valve core opening control end in the load-sensing valve assembly 6, and the load-sensing valve assembly 6 is configured to control the target operating speeds of the first actuator 71 and the second actuator 72 according to the output signals of the first pilot handle 51 and the second pilot handle 52, respectively; the pressure difference control module 9 is configured to provide a system target pressure difference.

[0028] When the system of this embodiment is operating normally, the motor speed changes in real time to maintain the system actual pressure difference at the system target pressure difference. When the system enters flow saturation, the motor speed reaches the maximum, and the system actual pressure difference will decrease, causing the valve target pressure difference to decrease synchronously, resulting in a decrease in the flow demand under the current flow area of ​​each throttle port. When the total flow demand drops to equal the pump output flow, the system actual pressure difference and the valve target pressure difference stop decreasing and remain constant, reaching a new equilibrium state. In this process, the pump target pressure difference is the system target pressure difference, which is always greater than the system actual pressure difference, so the motor speed always remains at the maximum value. Furthermore, when the flow area of ​​each main valve core of the system decreases, the system actual pressure difference will rise, causing the valve target pressure difference to rise synchronously. When the system actual pressure difference rises to equal the system target pressure difference, the motor speed will show a downward trend to maintain the system actual pressure difference at the system target pressure difference. At this point, the system exits the flow saturation state. Therefore, by actively regulating the valve's target pressure differential, the system can maintain the flow distribution ratio between the light-load and heavy-load links even when flow is saturated, unaffected by load differences. This achieves the flow saturation resistance of pre-valve compensation. Compared with existing pre-valve compensation load-sensitive systems, this effectively avoids actuator speed imbalance under flow saturation conditions, significantly improving controllability. Furthermore, compared with existing post-valve compensation load-sensitive systems with flow saturation resistance, this system can operate at different system pressure differentials to achieve variable valve port flow gain, avoiding system flow saturation. This system has a wider effective control stroke, significantly improving fine-tuning and flow control characteristics. Furthermore, the system utilizes a pre-valve compensation structure, eliminating additional switching throttling losses and minimizing valve port throttling losses.

[0029] Reference Manual Figure 2 A second embodiment of the present invention provides a control method for a load-sensing system of an electric-driven quantitative pump, comprising the following steps: Step S1: obtaining a system target pressure difference signal set by a pressure difference control module; Step S2: obtaining a maximum load pressure signal detected by the first pressure sensor in the sensor assembly and a pump outlet pressure signal detected by the second pressure sensor, and subtracting the maximum load pressure from the pump outlet pressure to obtain an actual system pressure difference; Step S3: setting the valve target pressure difference to the actual pressure difference of the system, controlling the target pressure difference of the proportional pressure reducing valve output valve in the pilot hydraulic control unit assembly to the first pressure compensation valve and the second pressure compensation valve in the load sensing valve assembly, thereby controlling the front and rear end pressure differences of the first three-position four-way reversing valve and the second three-position four-way reversing valve to be equal to the current actual pressure difference of the system; Step S4: With the system target pressure difference as the target value and the system actual pressure difference as the feedback value, closed-loop control is performed through the variable speed control of the drive motor. The closed-loop controller outputs the target speed signal, which is subjected to amplitude limiting processing (limiting the maximum speed and the minimum speed) and then sent to the motor controller to control the motor speed to maintain the system actual pressure difference stable at the system target pressure difference; return to step S1 to perform the next round of control cycle.

[0030] The third embodiment of the present invention further provides an engineering machine equipped with the above-mentioned load-sensing system for the electrically driven quantitative pump.

[0031] The engineering machinery may be, for example, a series of large-scale engineering machinery equipment such as excavators and earthmovers.

[0032] The present invention adopts a load-sensitive valve assembly (including a three-position four-way reversing valve, and a pressure compensation valve located at the oil inlet, which belongs to pre-valve compensation), matched with a pilot handle assembly and a pressure differential control module, and indirectly controls the valve group through a hydraulic control signal. The pressure differential control module directly gives the system target pressure differential, and the valve target pressure differential is set to the actual pressure differential detected by the sensor. The actual pressure differential is then stabilized at the target pressure differential through motor variable speed closed-loop control, focusing on solving the problems of anti-flow saturation difference in pre-valve compensation, additional switching energy consumption in post-valve compensation, and narrow flow control range caused by fixed pressure differential.

[0033] The above embodiments are intended only to illustrate the technical concepts and features of the present invention. Their purpose is to enable those skilled in the art to understand the contents of the present invention and implement them accordingly. They are not intended to limit the scope of protection of the present invention. Any equivalent changes or modifications made in accordance with the spirit of the present invention are intended to be covered by the scope of protection of the present invention.

Claims

1. A load-sensing system for an electrically driven quantitative pump, characterized in that: include: Control unit, drive motor assembly, metering pump, pilot hydraulic control unit, pilot handle assembly, load sensing valve assembly, actuator assembly, sensor assembly and differential pressure control module; The output end of the control unit is connected to the drive motor assembly and drives the metering pump to operate. The metering pump outputs high-pressure oil to the load-sensing valve assembly. The load-sensing valve assembly is configured to regulate the flow entering the actuator assembly to drive the actuator assembly to operate; the sensor assembly is connected to each valve in the load-sensing valve assembly and measures its pressure signal. The pressure difference control module derives the pressure difference signal between the front and rear ends of each valve based on the pressure signal obtained by the sensor assembly. The input end of the control unit is connected to the sensor assembly and the pressure difference control module. The control unit regulates the valve target pressure difference through the load-sensing valve assembly based on the pressure signal and the pressure difference signal between the front and rear ends of each valve, thereby controlling the target speed of the drive motor assembly and the control through the pilot handle assembly The control pressure of the pilot hydraulic control unit; the various valves in the load-sensitive valve assembly include a differential relief valve, a first pressure-compensating valve, a second pressure-compensating valve, a first three-position four-way reversing valve, a second three-position four-way reversing valve and a shuttle valve, the differential relief valve is arranged at the oil inlet of the load-sensitive valve assembly, the first pressure-compensating valve is arranged at the oil inlet of the first three-position four-way reversing valve, the second pressure-compensating valve is arranged at the oil inlet of the second three-position four-way reversing valve, and the shuttle valve is arranged at the oil outlet of the three-position four-way reversing valve; one end of the valve core of the first pressure-compensating valve and the second pressure-compensating valve compensates for the front-end pressure of the first three-position four-way reversing valve and the second three-position four-way reversing valve, and the other end compensates for the load pressure and target control pressure at the rear end of the first three-position four-way reversing valve and the second three-position four-way reversing valve.

2. The load sensing system of the electric driven metering pump according to claim 1, characterized in that: The pilot control ends of the first three-position four-way reversing valve and the second three-position four-way reversing valve are connected to the output end of the pilot handle assembly.

3. The load sensing system of the electric driven quantitative pump according to claim 2, characterized in that: The target control pressure oil circuits of the first pressure compensating valve and the second pressure compensating valve are connected in parallel and connected to the output end of the pilot hydraulic control unit.

4. The load sensing system of the electric driven metering pump according to claim 3, characterized in that: The first pressure compensating valve and the second pressure compensating valve are configured to respectively maintain the front and rear end pressure differences of the first three-position four-way reversing valve and the front and rear end pressure differences of the second three-position four-way reversing valve at the valve target pressure difference of the pressure-controllable load-sensitive valve regulated by the proportional pressure reducing valve in the pilot hydraulic control unit; the first three-position four-way reversing valve and the second three-position four-way reversing valve are configured to control the operating speed of the actuator assembly by adjusting the valve port opening; the fixed differential relief valve is configured to prevent system overpressure; and the shuttle valve is configured to obtain the maximum load pressure of the actuator drive chamber of each branch.

5. The load sensing system of the electric driven quantitative pump according to claim 2, characterized in that: The actuator assembly includes a first actuator and a second actuator, the first actuator is connected to the output port of the first three-position four-way reversing valve, and the second actuator is connected to the output port of the second three-position four-way reversing valve; the first actuator and the second actuator are hydraulic cylinders or hydraulic motors.

6. The load sensing system of the electric driven metering pump according to claim 1, characterized in that: The sensor assembly includes a first pressure sensor and a second pressure sensor. The first pressure sensor is arranged on the load-sensing valve assembly and is configured to detect the maximum load pressure of the system. The second pressure sensor is arranged on the connecting pipeline between the metering pump and the load-sensing valve assembly and is configured to detect the outlet pressure of the metering pump.

7. The load sensing system of the electric driven metering pump according to claim 5, characterized in that: The pilot handle assembly includes a first pilot handle and a second pilot handle. The input ends of the first pilot handle and the second pilot handle are connected to the pilot hydraulic control unit. The output ends of the first pilot handle and the second pilot handle are connected to the main valve core opening control end in the load-sensing valve assembly. The load-sensing valve assembly is configured to control the target operating speeds of the first actuator and the second actuator according to the output signals of the first pilot handle and the second pilot handle, respectively.

8. The load sensing system of the electric driven metering pump according to claim 1, characterized in that: The pressure differential control module is configured to provide a system target pressure differential.

9. A control method for a load sensing system of an electric-driven metering pump according to any one of claims 1 to 8, characterized in that: The steps include: Step S1: obtaining a system target pressure difference signal set by a pressure difference control module; Step S2: obtaining a maximum load pressure signal detected by the first pressure sensor in the sensor assembly and a pump outlet pressure signal detected by the second pressure sensor, and subtracting the maximum load pressure from the pump outlet pressure to obtain an actual system pressure difference; Step S3: setting the valve target pressure difference to the actual system pressure difference, controlling the proportional pressure reducing valve output valve target pressure difference in the pilot hydraulic control unit assembly to the first pressure compensation valve and the second pressure compensation valve in the load sensing valve assembly, thereby controlling the front and rear end pressure differences of the first three-position four-way reversing valve and the second three-position four-way reversing valve to be equal to the current actual system pressure difference; Step S4: using the system target pressure difference as the target value and the system actual pressure difference as the feedback value, closed-loop control is performed through the variable speed control of the drive motor. The closed-loop controller outputs a target speed signal which is limited and sent to the motor controller to control the motor speed to maintain the system actual pressure difference stable at the system target pressure difference. Return to step S1 to perform the next control cycle.

10. An engineering machine, characterized in that: It is equipped with an electric-driven quantitative pump load sensing system as claimed in any one of claims 1 to 8.

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