A walking machine hydrostatic system and a control method thereof

Abstract

The present application discloses a hydrostatic system for a walking machine and a control method thereof, wherein the hydrostatic system includes a variable piston pump, an oil charge pump, a first balancing valve, a large-flow check valve, a pressure-reducing main valve core, a first control cover plate, an overflow main valve core, a second control cover plate, a variable piston motor, and a second balancing valve; the control method includes the following steps: S1, obtaining a handle value; S2, if the handle value is less than 0, or the handle value is greater than or equal to 0 and the engine feedback load rate is greater than a, setting a target current I of the electric proportional overflow valve in the first control cover plate; f =0, the electric proportional relief valve is set to the maximum value; S3, if the handle value ≥ 0 and when the engine feedback load rate is < a, the target current I of the electric proportional relief valve is calculated based on the current displacement value of the variable piston pump and the rated braking torque T of the engine. f The electric proportional relief valve is controlled in real time. This application solves the technical problems of severe system heating, slow response, potential safety hazards, and easy damage to components.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of hydraulic technology, in particular, to a walking machine hydrostatic system and a control method thereof. BACKGROUND

[0002] Hydrostatic system has the advantages of good micro-motion, large speed stiffness, high transmission efficiency, easy starting, braking, reversing operation, and easy realization of electro-hydraulic composite control. In recent years, with the development of hydraulic industry, the reliability of hydraulic components is continuously improved, and the cost is continuously reduced. Full-hydraulic drive walking engineering machinery has appeared at home and abroad.

[0003] The existing walking drive part of engineering machinery is mainly braked by a hydrostatic system, but some problems occur during braking: engine is overspeed by reverse dragging in high-speed braking or long downhill working condition, element damage caused by excessive pressure impact on the braking side, system pressure loss caused by the inability of the forward side to supply oil in time, etc. In view of these problems, relevant technical personnel has given some solutions, mainly including the following improvement technologies: changing the oil return back pressure by connecting a relief valve in series on the braking side, reducing the circuit impact pressure by connecting a relief valve in parallel on the braking side, and controlling the swash plate angle of the variable pump to be instantaneously zero when the engine is abnormally overspeed.

[0004] The existing technology has the following deficiencies:

[0005] The closed system variable plunger pump itself integrates a bidirectional high-pressure relief valve and an oil supply check valve. When the pressure on one side of the circuit exceeds the set value, it will overflow to the opposite side through the valve. However, due to the limitation of the flow capacity of the high-pressure relief valve and the oil supply check valve, the pressure generated on the braking side usually far exceeds the set value of the relief valve, resulting in engine overspeed, element damage, and system pressure loss on the other side, causing problems such as pump and motor knocking. The existing technology mainly proposes corresponding solutions based on this actual situation, for example:

[0006] 1. Patent "Brake control system applied to closed hydraulic drive system and control method thereof-CN107323446A" connects a relief valve in series on the braking side. When high pressure is generated on the braking side, the oil return back pressure is changed through the relief valve connected in series on the braking side, and then enters the forward side through the check valve. The kinetic energy of the whole machine is completely absorbed by the relief valve connected in series, and does not act on the hydraulic pump, so as to ensure that the engine will not be overspeed by reverse dragging.

[0007] (1) This scheme avoids engine overspeed by reverse dragging by canceling the effect of brake high-pressure oil on the pump, but the kinetic energy of the vehicle is completely overflowed by the relief valve to absorb heat, and the brake torque of the engine itself is not utilized, resulting in serious overflow heat of the system.

[0008] (2) The system judges whether the engine is overspeed through the feedback value of the pressure sensor, and then controls the pressure to meet the set value through closed-loop regulation. After the overflow valve is opened, most of the flow on the brake side will pass through the valve. Due to the influence of the pressure gain of the overflow valve, the actual pressure may be far higher than the set pressure, so the electric proportional overflow valve needs to be adjusted again. Because the braking condition time is relatively short, repeated adjustment greatly reduces the response speed of the system;

[0009] 2. The patent "Hydraulic braking scheme-CN105546077A" reduces the pressure on the brake side of the motor through the parallel overflow valve, which can reduce the pressure impact on the brake side to a certain extent, avoid exceeding the engine braking torque, but at the same time, the hydraulic system reduces the reverse braking torque of the motor, which increases the braking distance of the vehicle and reduces the safety of the vehicle;

[0010] 3. The patent "Engine and hydraulic system combined braking device-CN102416942A" adds an electromagnetic reversing valve to the variable mechanism circuit of the variable piston pump. When the engine is detected to be overspeed, the variable pump swash plate is controlled to return to zero instantaneously:

[0011] (1) This method can instantaneously reduce the reverse torque between the pump and the engine, but this scheme will cause the brake side oil to increase suddenly, resulting in ultra-high pressure and instantaneous pressure loss on the forward side, which may cause damage to the pump, motor and other components;

[0012] (2) Instantaneous cut-off of the pump displacement can easily cause the vehicle to brake rapidly at a very high deceleration, which is beyond the control range of the driver and has a high safety risk. SUMMARY

[0013] The present application provides a walking machine hydraulic system to solve the technical problems of existing technology, such as serious heat generation, slow response, safety risk and easy damage to components.

[0014] The technical scheme adopted by the present application is as follows:

[0015] A walking machine hydraulic system, comprising a variable piston pump, an oil supplement pump, a first balance valve, a large-flow check valve, a pressure reduction main valve core, a first control cover plate, an overflow main valve core, a second control cover plate, a variable piston motor, and a second balance valve, wherein:

[0016] The variable piston pump and the oil supplement pump are coaxially driven. The B port of the variable piston pump is connected to the input end of the first balance valve. The A port of the variable piston pump is connected to the input end of the second balance valve. The output end of the oil supplement pump is connected to the output end of the first balance valve, the output end of the second balance valve, and the control port of the variable piston pump, respectively.

[0017] The pressure reducing main valve core comprises a two-way plug-in pressure reducing valve, the first control cover plate comprises an electric proportional overflow valve, the A port of the two-way plug-in pressure reducing valve is connected with the B port of the variable piston pump, the B port of the two-way plug-in pressure reducing valve is connected with the B port of the variable piston motor, the control port of the two-way plug-in pressure reducing valve is connected with the input end of the electric proportional overflow valve through a damping hole, and the output end of the electric proportional overflow valve is connected with the control port of the variable piston pump and the variable piston motor.

[0018] The input end of the large-flow one-way valve is connected with the B port of the variable piston pump, and simultaneously connected with the input end of the electric proportional overflow valve through a damping hole, and the output end of the large-flow one-way valve is connected with the B port of the variable piston motor.

[0019] The overflow main valve core comprises a two-way plug-in overflow valve, the second control cover plate comprises a second overflow valve, the A port of the two-way plug-in overflow valve is connected with the B port of the variable piston motor, the B port of the two-way plug-in overflow valve is connected with the A port of the variable piston motor, the control port of the two-way plug-in overflow valve is connected with the input end of the second overflow valve through a damping hole, the input end of the second overflow valve is also connected with the B port of the variable piston motor through another damping hole, and the output end of the second overflow valve is connected with the control port of the variable piston motor.

[0020] Preferably, a first overflow valve is further arranged on the pipeline between the output end of the oil supplement pump and the control port of the variable piston pump.

[0021] Preferably, a filtering device is further arranged on the output end of the oil supplement pump.

[0022] Preferably, the large-flow one-way valve adopts a two-way plug-in one-way valve, the A port of the two-way plug-in one-way valve is connected with the B port of the variable piston pump, simultaneously connected with the input end of the electric proportional overflow valve through a damping hole, the B port of the two-way plug-in one-way valve is connected with the B port of the variable piston motor, and the control port of the two-way plug-in one-way valve is connected with the B port of the two-way plug-in one-way valve through a damping hole.

[0023] Preferably, the large-flow one-way valve adopts a pipe one-way valve.

[0024] Preferably, a radiator is further arranged, and the input end of the radiator is connected with the overflow ends of the electric proportional overflow valve and the second overflow valve.

[0025] Another aspect of the present application further provides a control method of the walking machine hydrostatic system, comprising the following steps:

[0026] S1, obtaining a handle value;

[0027] S2, if the handle value is less than 0, or the handle value is greater than or equal to 0 and the engine feedback load rate is greater than a, setting the target current I of the electric proportional overflow valve in the first control cover plate f=0, at this time the electric proportional relief valve is set to the maximum value;

[0028] S3. If the handle value is ≥ 0 and the engine feedback load rate is < a, the target current I of the electric proportional relief valve in the first control cover is calculated based on the current displacement value of the variable piston pump and the rated braking torque T of the engine. f Real-time control of the electric proportional relief valve.

[0029] Preferably, the step S3 specifically includes the steps of:

[0030] S31. Calculate the current displacement V of the variable displacement piston pump according to its operating parameters. p ;

[0031] S32, based on the engine rated braking torque T, the current displacement V of the variable piston pump p Calculate the target pressure value P of the brake pump inlet p :

[0032]

[0033] S33, according to the target pressure value P p Calculate the target current I of the electric proportional relief valve in the first control cover f Real-time control of the electric proportional relief valve:

[0034] I f =f1(P p ).

[0035] Preferably, the current control current feedback value I of the variable piston pump is specifically p , combined with the displacement-current curve of the variable piston pump, the current displacement V of the variable piston pump is calculated p =f2(I p ).

[0036] Preferably, in step S31, the inclination angle θ of the swash plate of the variable displacement piston pump is obtained by an angle sensor. p , combined with the displacement-angle curve of the variable piston pump, calculate the current displacement of the variable piston pump

[0037] V p =f3(θ p ).

[0038] Compared with the existing technology, this application has the following beneficial effects:

[0039] The application provides a walking machine hydrostatic system and a control method thereof. The application adopts a small number of common elements to avoid the problem of engine reverse drag overspeed in emergency braking and long downhill working conditions under the premise of ensuring braking torque and with a faster response speed. The application increases a high-pressure and large-flow overflow valve at the outlet of the brake side variable plunger motor to limit the highest impact pressure of the brake side under the premise of ensuring braking distance and protect the hydraulic elements. The application compensates for the deficiencies of the pump's own high-pressure overflow valve, the flow capacity of the oil supplement check valve and the oil supplement capacity of the oil supplement pump through the high-pressure and large-flow overflow valve. The brake side overflow oil can be smoothly supplemented to the forward side to avoid the pump and motor knocking cylinder problem caused by the pressure loss of the forward side. The application increases a high-pressure and large-flow pressure reducing valve at the inlet of the brake side pump to limit the pressure of the brake side pump inlet and ensure that the reverse drag torque of the pump acting on the engine is always not higher than the rated braking torque of the engine, thereby avoiding engine overspeed. The hydraulic system of the application adopts a more concise control logic. After judging that the engine is in a critical overspeed state, the target current value of the electric proportional overflow valve is calculated through the working parameters of the pump and the rated braking torque of the engine, and the electric proportional overflow valve is directly controlled in real time, thereby avoiding repeated adjustment caused by factors such as pressure feedback lag and pressure gain and improving the response speed of the system. In the critical overspeed working condition of the engine, the reverse drag torque of the brake side pump inlet under the action of high pressure is basically equal to the rated braking torque of the engine, thereby avoiding engine overspeed and fully utilizing the engine's own braking energy. The excess capacity is absorbed in the form of heat through the hydraulic system, thereby reducing system heating and energy loss.

[0040] In addition to the objects, features, and advantages described above, the application has other objects, features, and advantages. The application will be described in further detail below with reference to the drawings. BRIEF DESCRIPTION OF DRAWINGS

[0041] The accompanying drawings, which form a part of this application, are included to provide a further understanding of the application, illustrate the preferred embodiments of the application, and assist in the explanation of the application. In the drawings:

[0042] Figure 1 is a schematic diagram of the composition principle of the walking machine hydrostatic system of the preferred embodiment of the application;

[0043] Figure 2 is a flowchart of the control method of the preferred embodiment of the application;

[0044] Figure 3 is a sub-step flowchart of step S3 of the preferred embodiment of the application;

[0045] Figure 4 is a sub-step flowchart of step S3 of another preferred embodiment of the application;

[0046] Figure 5 This is a schematic diagram of the composition principle of a hydrostatic system for a walking machine according to another preferred embodiment of the present application;

[0047] Figure 6 It is a schematic diagram of the composition principle of a static hydraulic system of a walking machine according to another preferred embodiment of the present application.

[0048] In the figure: 1. Variable piston pump; 2. Oil charge pump; 3. First balancing valve; 4. First relief valve; 5. High-flow check valve; 6. Pressure-reducing main valve core; 7. First control cover; 8. Radiator; 9. Overflow main valve core; 10: Second control cover; 11: Variable piston motor; 12. Second balancing valve; 13: Insert-type electric proportional relief valve; 14: Insert-type pressure-reducing valve; 15: High-flow insert-type or tubular check valve; 16: Insert-type relief valve. DETAILED DESCRIPTION

[0049] The embodiments of the present application are described in detail below with reference to the accompanying drawings, but the present application can be implemented in a variety of different ways defined and covered below.

[0050] like Figure 1 As shown, the preferred embodiment of the present application provides a hydrostatic system for a walking machine, including a variable piston pump 1, an oil charge pump 2, a first balancing valve 3, a large flow check valve 5, a pressure reducing main valve core 6, a first control cover plate 7, a relief main valve core 9, a second control cover plate 10, a variable piston motor 11, a second balancing valve 12, and a radiator 8, wherein:

[0051] The variable piston pump 1 and the oil charge pump 2 are coaxially driven, the B port of the variable piston pump 1 is connected to the input end of the first balancing valve 3, the A port of the variable piston pump 1 is connected to the input end of the second balancing valve 12, and the output end of the oil charge pump 2 is respectively connected to the output end of the first balancing valve 3, the output end of the second balancing valve 12 and the control port of the variable piston pump 1;

[0052] The pressure-reducing main valve core 6 includes a two-way cartridge pressure-reducing valve, and the first control cover plate includes an electric proportional relief valve. Port A of the two-way cartridge pressure-reducing valve is connected to port B of the variable piston pump 1, and port B of the two-way cartridge pressure-reducing valve is connected to port B of the variable piston motor 11. The control port of the two-way cartridge pressure-reducing valve is connected to the input end of the electric proportional relief valve through a damping orifice, and the output end of the electric proportional relief valve is connected to the control ports of the variable piston pump 1 and the variable piston motor 11.

[0053] The input end of the large flow check valve 5 is connected to the B port of the variable piston pump 1, and at the same time, is connected to the input end of the electric proportional relief valve through the damping hole. The output end of the large flow check valve 5 is connected to the B port of the variable piston motor 11;

[0054] The overflow main valve core 9 comprises a two-way plug-in overflow valve, the second control cover plate 10 comprises a second overflow valve, the A port of the two-way plug-in overflow valve is connected with the B port of the variable piston motor 11, the B port of the two-way plug-in overflow valve is connected with the A port of the variable piston motor 11, the control port of the two-way plug-in overflow valve is connected with the input end of the second overflow valve through a damping hole, the input end of the second overflow valve is also connected with the B port of the variable piston motor 11 through another damping hole, and the output end of the second overflow valve is connected with the control port of the variable piston motor 11.

[0055] The pipeline between the output end of the oil supplement pump 2 and the control port of the variable piston pump 1 is also provided with the first overflow valve 4, so as to ensure that the output pressure of the oil supplement pump 2 is kept as a set value.

[0056] The output end of the oil supplement pump 2 is also provided with a filtering device, so as to ensure the cleanliness of the oil in the hydraulic pipeline.

[0057] The large-flow one-way valve 5 adopts a two-way plug-in one-way valve, the A port of the two-way plug-in one-way valve is connected with the B port of the variable piston pump 1, at the same time, is connected with the input end of the electric proportional overflow valve through a damping hole, the B port of the two-way plug-in one-way valve is connected with the B port of the variable piston motor 11, and the control port of the two-way plug-in one-way valve is connected with the B port of the two-way plug-in one-way valve through a damping hole. Preferably, the large-flow one-way valve 5 adopts a pipe one-way valve.

[0058] The input end of the radiator 8 is connected with the overflow ends of the electric proportional overflow valve and the second overflow valve.

[0059] In the above embodiment, the variable plunger pump 1 is a closed walking system power source, and the swash plate can be bidirectional variable; the oil supplement pump 2 is a gear pump, and in the closed system, the oil liquid of the oil supplement pump 2 mainly functions as: low pressure side oil supplement of the closed system, motor shell flushing, and control of the swash plate variable of the variable plunger pump 1; the first balance valve 3 and the second balance valve 12 are high pressure overflow valves, and the main function is that when the oil pressure of one side of the closed system is higher than the set value of the overflow valve, the overflow valve is opened, the high pressure side oil liquid enters the low pressure side, and the pump inlet and outlet oil liquid is circulated; the first overflow valve 4 is an oil supplement overflow valve, which controls the outlet pressure of the oil supplement pump; the large flow one-way valve 5 is a large flow two-way plug-in one-way valve, which is only opened when the vehicle is reversing, and the oil liquid enters the main oil way through the large flow one-way valve 5, avoiding the high pressure of the B port of the variable plunger pump 1 when reversing, causing the electric proportional overflow valve to tend to be closed, and the vehicle cannot walk; the pressure reducing main valve core 6 is a two-way plug-in valve core with pressure reducing function, which limits the B port pressure of the variable plunger pump 1 on the brake side, avoiding the generation of large reverse drag torque of the pump to cause the engine to overspeed; the first control cover plate 7 is a pilot control module with an electric proportional overflow valve, which adjusts the setting value of the electric proportional overflow valve in real time according to the pump displacement feedback, so that the B port pressure of the pump is always higher than the setting value of the electric proportional overflow valve by the spring pressure of the main valve core; the electric proportional overflow valve has the maximum setting pressure when losing power; the radiator 8 radiates the small flow hot oil after the overflow of the pump and motor internal leakage, flushing flow, electric proportional overflow valve and second overflow valve control oil way;

[0060] The overflow main valve core 9 is a two-way plug-in valve core with overflow function, which mainly limits the motor outlet pressure on the brake side, on the one hand, to ensure that the load end has enough braking force and shorten the braking distance; on the other hand, to reduce the pressure impact between the motor outlet and the element, protect the hydraulic element, and at the same time, the large flow capacity makes the oil liquid smoothly supplement to the system forward side, avoiding the pressure loss of the forward side; the second control cover plate 10 is a pilot control module with a direct-acting overflow valve, and the spring force set value of the valve core determines the maximum pressure of the motor outlet on the brake side; the variable plunger motor 11 is a walking axle drive motor, which provides driving and braking torque for the whole vehicle.

[0061] The present embodiment provides a hydrostatic system for a walking machine, comprising a variable piston pump 1, an oil replenishing pump 2, a first balancing valve 3, a high-flow check valve 5, a pressure-reducing main valve core 6, a first control cover plate 7, an overflow main valve core 9, a second control cover plate 10, a variable piston motor 11, and a second balancing valve 12. The present embodiment, through the design of the hydraulic circuit and the system control logic, adopts a relatively small number of commonly used components to avoid the problem of engine overspeeding during emergency braking and long downhill conditions with a relatively fast response speed while ensuring the braking torque; the present embodiment adds a high-pressure, high-flow overflow valve at the outlet of the variable piston motor on the braking side, which limits the maximum impact pressure on the braking side while ensuring the braking distance, thereby protecting the hydraulic components; the present embodiment compensates for the deficiencies in the flow capacity of the pump's own high-pressure overflow valve, the oil replenishing check valve, and the oil replenishing capacity of the oil replenishing pump through the high-pressure, high-flow overflow valve, so that the oil overflowing from the braking side can be smoothly replenished to the On the forward side, this prevents pump and motor knocking caused by pressure loss on the forward side. This embodiment adds a high-pressure, high-flow pressure reducing valve to the brake-side pump inlet, limiting the pressure at the brake-side pump inlet and ensuring that the pump's back-drag torque on the engine never exceeds the engine's rated braking torque, thus preventing engine overspeed. This application controls the back-drag torque under the high pressure at the brake-side pump inlet to be substantially equal to the engine's rated braking torque during critical engine overspeed conditions, preventing engine overspeed while fully utilizing the engine's own braking energy. Excess energy is absorbed through heat generation in the hydraulic system, reducing system heat generation and energy loss. This embodiment provides a radiator 8 to direct the low-flow, high-temperature oil overflowing from the electric proportional relief valve in the first control cover plate 7 and the second relief valve in the second control cover plate 10 back to the radiator 8. This increases the system flushing flow, ensures system thermal balance, and protects the hydraulic system and components.

[0062] like Figure 2 Another preferred embodiment of the present application further provides a method for controlling the static hydraulic system of a walking machine as described above, comprising the steps of:

[0063] S1. Get the handle value;

[0064] S2. If the handle value is less than 0 (indicating that the vehicle is in reverse operation, and since the reverse speed is relatively low, there is rarely a problem of high brake pressure or engine backdraft), or if the handle value is greater than or equal to 0 and the engine feedback load rate is greater than a (indicating that the vehicle is in a stationary standby state or in a normal forward (not downhill) operation), then set the target current I of the electric proportional relief valve in the first control cover plate 7. f =0, at this time the electric proportional relief valve is set to the maximum value;

[0065] S3. If the handle value is ≥ 0 and the engine feedback load rate is < a (indicating that the engine is in a critical overspeed state and the brake pump inlet pressure needs to be limited), the target current I of the electric proportional relief valve in the first control cover plate 7 is calculated based on the current displacement value of the variable piston pump 1 and the rated braking torque T of the engine. f Real-time control of the electric proportional relief valve.

[0066] The hydraulic system of the present application adopts a more concise control logic. After determining that the engine is in a critical overspeed state, the target current value of the electric proportional relief valve is calculated based on the operating parameters of the pump and the rated braking torque of the engine. The electric proportional relief valve is then directly controlled in real time. This avoids repeated adjustments caused by factors such as pressure feedback lag and pressure gain, thereby improving the system's response speed.

[0067] Preferably, if Figure 3 As shown, the step S3 specifically includes the following steps:

[0068] S301, through the current control current feedback value I of the variable piston pump 1 p , combined with the displacement-current curve of the variable piston pump 1, the current displacement V of the variable piston pump 1 is calculated p =f2(I p );

[0069] S302, according to the rated braking torque T of the engine and the current displacement V of the variable displacement piston pump 1 p Calculate the target pressure value P of the brake pump inlet p :

[0070]

[0071] S303, according to the target pressure value P p Calculate the target current I of the electric proportional relief valve in the first control cover 7 f Real-time control of the electric proportional relief valve:

[0072] I f =f1(P p ).

[0073] Preferably, if Figure 4 As shown, the step S3 specifically includes the following steps:

[0074] S311, obtain the inclination angle θ of the swash plate of the variable displacement piston pump 1 through the angle sensor p , combined with the displacement-angle curve of the variable piston pump 1, the current displacement V of the variable piston pump 1 is calculated p =f3(θ p );

[0075] S312, according to the engine rated braking torque T, the current displacement V of the variable displacement piston pump 1 p Calculate the target pressure value P of the brake pump inlet p :

[0076]

[0077] S313, according to the target pressure value P p Calculate the target current I of the electric proportional relief valve in the first control cover 7 f Real-time control of the electric proportional relief valve:

[0078] I f =f1(P p ).

[0079] In reality, engine backdraft occurs in two situations:

[0080] The first is a long downhill condition. If only hydrostatic braking is used, the vehicle's large inertia will cause the torque generated to exceed the engine's braking torque, causing the vehicle to accelerate downward. In this case, mechanical friction is generally the only assisted braking method, but long-term mechanical friction will accelerate the decline in the life of the friction plate.

[0081] The second is the high-speed braking condition. At this time, the motor changes to the pump condition, and the braking side becomes the high-pressure side. Due to the limited flow capacity of the variable displacement piston pump's own high-pressure relief valve, the oil replenishment check valve, and the replenishment capacity of the replenishment pump, the oil on the braking side cannot overflow to the forward side instantly, resulting in a sharp increase in the pressure on the braking side. When the torque under the combined action of the high pressure on the braking side and the pump exceeds the rated braking torque of the engine, the engine is dragged into overspeed.

[0082] like Figure 1 As shown, when the vehicle is moving forward, the pressure at port A of the pump and motor is high, and the pressure at port B is low; when the vehicle is braking or working on a long downhill slope, the pressure at port B of the pump and motor is high, and the pressure at port A is low; when the vehicle is moving backward, the pressure at port B of the pump and motor is high, and the pressure at port A is low, that is, the pressures at port A and port B of the pump and motor are basically the same.

[0083] (1) Emergency braking, long downhill conditions: The oil pressure at the brake-side outlet B of the variable piston motor 11 is greater than the set value of the direct-acting relief valve integrated on the second control cover 10. The oil opens the direct-acting relief valve through the control port and flows out through the Y port to the radiator 8. The flowing oil generates a pressure difference through the damping hole, and the overflow main valve core 9 is opened by the generated pressure difference. The high-pressure oil at the B port of the variable piston motor 11 flows through the A port of the overflow main valve core 9 to the B port and enters the forward oil circuit between the pump A port and the motor A port.

[0084] Further, due to the large flow capacity of the overflow main valve core 9, the pressure gain is low, and the B port pressure of the variable piston motor 11 is basically maintained at the set value of the direct-acting overflow valve spring integrated on the second control cover plate 10, ensuring the braking torque while avoiding damage to the motor caused by excessive high pressure; at the same time, the inlet B pressure of the pressure relief main valve core 6 is limited, which also protects the electric proportional overflow valve.

[0085] Further, due to the large flow capacity of the overflow main valve core 9, the oil in the B port of the variable piston motor 11 smoothly overflows to the forward side, avoiding the loss of pressure from the pump A port to the motor A port on the forward side due to insufficient flow capacity of the first balance valve 3 and the one-way valve of the variable piston pump 1 and the limited oil supplementing capacity of the oil supplementing pump 2, avoiding the problem of variable piston pump 1 and variable piston motor 11 knocking, and prolonging the service life of the components.

[0086] (2) Over-speed working condition: The overflow main valve core 9 and the second control cover plate 10 limit the maximum pressure of the B port of the variable piston motor 11. In order to ensure the braking distance, the pressure is set relatively high. If this pressure directly impacts the B port of the variable piston pump 1, it is easy to cause the engine to reverse and over-speed.

[0087] Further, by feeding back the current load rate of the engine, it is determined whether the engine is approaching an over-speed working condition. When the load rate is higher than the set value and the engine is not over-speed, the electric proportional overflow valve of the first control cover plate 7 loses power, the pressure relief main valve core 6 is set to the maximum value, the main valve core 6 is in the fully open state, and the system pressure is low and cannot trigger the pressure relief function. When the load rate is lower than the set value, it is determined that the engine has tended to an over-speed working condition. According to the feedback value of the displacement control current of the variable piston pump 1, the current displacement of the pump is calculated. Then, according to the rated braking torque of the engine and the current displacement of the pump, the limiting value of the B port pressure of the pump on the braking side is calculated, and then the target current value of the electric proportional overflow valve of the first control cover plate 7 is calculated. The controller directly outputs the target current value to the electric proportional overflow valve. When the A port pressure of the pressure relief main valve core 6 is higher than the set value of the electric proportional overflow valve of the first control cover plate 7, the oil flows through the control port X of the first control cover plate 7 to open the overflow valve, flows out through the Y port to the radiator 8, and the flowing oil generates a pressure difference through the damping hole, so that the pressure relief main valve core 6 is in a dynamic balance adjustment state, ensuring that the A port pressure of the pressure relief main valve core 6 is always higher than the set pressure of the electric proportional overflow valve in the first control cover plate 7 by the spring force of the main valve core. The excess high pressure energy of the B port of the pressure relief main valve core 6 is dissipated in the form of heat under the action of the valve core, effectively avoiding the problem of the engine over-speed caused by the high pressure of the brake dragging the pump to reverse.

[0088] Further, by the pressure regulation of the electric proportional overflow valve of the first control cover plate 7, the reverse drag torque of the high pressure at the brake side pump inlet B port of the engine is matched with the brake torque of the engine, the engine overspeed is avoided, and the brake energy of the engine is fully utilized; the excess kinetic energy of the vehicle is consumed through the hydraulic system heat energy; the heat generation of the hydraulic system is reduced to the greatest extent, and the energy loss is reduced;

[0089] (3) Emergency braking, long downhill working condition: to further reduce the heat generation of the hydraulic system and protect the hydraulic elements, the hot oil overflowing from the first control cover plate 7 and the second control cover plate 10 through the Y port, the shell oil leakage of the variable piston pump 1 and the variable piston motor 11 and the hot oil of the loop flushing are connected back to the radiator 8 (the flow after the overflow of the control oil circuit is low and has no effect on the system oil supplement), the closed system flushing flow is increased, and the loop heat balance is ensured.

[0090] (4) Reverse working condition: when the vehicle is reversing, the high pressure at the B port of the variable piston pump 1 acts on the A port of the pressure relief main valve core 6, so that the opening of the pressure relief main valve core 6 tends to decrease; in order to avoid that the high pressure oil at the B port of the variable piston pump 1 makes the pressure relief main valve core 6 tend to close and the vehicle cannot reverse normally, a large-flow one-way valve 5 is connected in parallel with the pressure relief main valve core 6; the high pressure oil at the B port of the variable piston pump 1 enters the main oil circuit between the B port of the large-flow one-way valve 5 and the B port of the variable piston motor 11 through the A port of the large-flow one-way valve 5 in a forward direction, and the motor rotates reversely.

[0091] In summary, the application has the following technical key points:

[0092] 1. A high-pressure large-flow overflow valve is added to the brake side motor outlet, which reduces the impact pressure of the motor outlet, protects the hydraulic elements and prolongs the service life of the elements while ensuring the brake torque;

[0093] 2. A high-pressure large-flow overflow valve is added to the brake side motor outlet, due to its large flow capacity, the overflowing oil can be supplemented to the forward side in time, avoiding the problem of element knocking caused by pressure loss of the forward side;

[0094] 3. A high-pressure large-flow pressure relief valve is added to the brake side pump inlet, which adjusts the brake side pump inlet pressure in real time when the engine is in a critical overspeed working condition, so that the reverse drag torque generated by the variable piston pump is always less than the rated brake torque of the engine, avoiding engine overspeed;

[0095] 4. Whether the engine is in a critical overspeed state is determined by the engine load rate, the set current value of the electric proportional overflow valve is calculated by the pump displacement control current or the swash plate inclination angle and the rated brake torque of the engine, and is directly assigned, which greatly improves the system response speed;

[0096] 5. When the engine is in a critical overspeed state, the reverse drag torque under the action of the high pressure at the pump inlet is basically equal to the rated braking torque of the engine, which fully utilizes the engine's own braking energy. The excess energy is consumed in the form of heat generated by the hydraulic system, reducing system heat and energy loss.

[0097] 6. The small flow of high-temperature oil after the overflow of the control cover is led back to the radiator to increase the flushing of the closed system, ensure the thermal balance of the system, and protect the hydraulic system and components;

[0098] 7. A high-flow two-way plug-in one-way valve is connected in parallel with the electric proportional pressure reducing valve on the braking side. When the vehicle is braking or in a long downhill slope, the oil on the braking side is reversely cut off. When the vehicle moves backward, the oil opens the one-way valve to drive the motor, preventing high-pressure oil from flowing through the electric proportional pressure reducing valve in the reverse direction when the vehicle moves backward, causing the pressure reducing main valve core to close and the vehicle unable to move backward.

[0099] Preferably, Figure 1 In the embodiment, the pressure reducing main valve core 6 and the first control cover plate 7 are combined into a two-way cartridge-type large flow pressure reducing valve. In this embodiment, the two-way cartridge-type large flow pressure reducing valve is replaced with a hydraulic valve in the form of a plug-in type with a larger flow capacity, including a plug-in type electric proportional relief valve 13 and a plug-in type pressure reducing valve 14; the large flow check valve 5 using a two-way cartridge check valve is replaced with a large flow plug-in type or tubular check valve 15, and the two-way cartridge-type large flow relief valve formed by combining the overflow main valve core 9 and the second control cover plate is replaced with a plug-in type relief valve 16 (see Figure 5 The working principle of this embodiment is similar to Figure 1 Basically the same, by replacing the valve with a plug-in type, the installation space of the hydraulic system can be reduced and the degree of integration can be higher.

[0100] Since the parts of the plug-in type are split, the core type is an integral valve. In comparison, the core type takes up less space and is easier to maintain and disassemble; it has a higher degree of integration; but at the same time, the cost will be correspondingly higher.

[0101] Preferably, Figure 1 middle, Figure 2 The main consideration is the engine overspeed problem caused by the forward working condition; the alternative solution of this embodiment is to add symmetrical valves in the forward and reverse directions of the vehicle to ensure that the engine overspeed problem will not occur in any working condition (see Figure 6 )

[0102] The details which are not described in the present application are the known technology of the skilled in the art. Finally, it is to be explained that the above examples are only used to illustrate the technical solutions of the present application but not to limit the present application. Although the present application has been described in detail with reference to the preferred embodiments, those skilled in the art should understand that the technical solutions of the present application can be modified or equivalently replaced without departing from the purpose and scope of the present application, and they should be covered in the scope of the claims of the present application.

Claims

1. A hydrostatic system for a walking machine, characterized in that: The invention comprises a variable piston pump (1), an oil replenishing pump (2), a first balancing valve (3), a large flow check valve (5), a pressure reducing main valve core (6), a first control cover plate (7), an overflow main valve core (9), a second control cover plate (10), a variable piston motor (11), and a second balancing valve (12), wherein: The variable piston pump (1) and the oil replenishing pump (2) are coaxially driven, the B port of the variable piston pump (1) is connected to the input end of the first balancing valve (3), the A port of the variable piston pump (1) is connected to the input end of the second balancing valve (12), and the output end of the oil replenishing pump (2) is respectively connected to the output end of the first balancing valve (3), the output end of the second balancing valve (12) and the control port of the variable piston pump (1); The pressure reducing main valve core (6) includes a two-way plug-in pressure reducing valve, the first control cover plate includes an electric proportional relief valve, the A port of the two-way plug-in pressure reducing valve is connected to the B port of the variable plunger pump (1), the B port of the two-way plug-in pressure reducing valve is connected to the B port of the variable plunger motor (11), the control port of the two-way plug-in pressure reducing valve is connected to the input end of the electric proportional relief valve through a damping hole, and the output end of the electric proportional relief valve is connected to the control ports of the variable plunger pump (1) and the variable plunger motor (11); The input end of the large flow check valve (5) is connected to the B port of the variable plunger pump (1), and is connected to the input end of the electric proportional relief valve through a damping hole. The output end of the large flow check valve (5) is connected to the B port of the variable plunger motor (11); The overflow main valve core (9) includes a two-way plug-in overflow valve, and the second control cover plate (10) includes a second overflow valve. The A port of the two-way plug-in overflow valve is connected to the B port of the variable plunger motor (11), and the B port of the two-way plug-in overflow valve is connected to the A port of the variable plunger motor (11). The control port of the two-way plug-in overflow valve is connected to the input end of the second overflow valve through a damping hole. The input end of the second overflow valve is also connected to the B port of the variable plunger motor (11) through another damping hole. The output end of the second overflow valve is connected to the control port of the variable plunger motor (11).

2. The hydrostatic system for mobile machinery according to claim 1, characterized in that: A first overflow valve (4) is also provided on the pipeline between the output end of the oil replenishment pump (2) and the control port of the variable displacement plunger pump (1).

3. The hydrostatic system for mobile machinery according to claim 2, characterized in that: The output end of the oil replenishment pump (2) is also provided with a filtering device.

4. The hydrostatic system for mobile machinery according to claim 2, characterized in that: The large flow check valve (5) adopts a two-way plug-in check valve, wherein the A port of the two-way plug-in check valve is connected to the B port of the variable plunger pump (1), and is connected to the input end of the electric proportional relief valve through a damping hole, the B port of the two-way plug-in check valve is connected to the B port of the variable plunger motor (11), and the control port of the two-way plug-in check valve is connected to the B port of the two-way plug-in check valve through a damping hole.

5. The hydrostatic system for mobile machinery according to claim 2, characterized in that: The large flow one-way valve (5) is a tubular one-way valve.

6. The hydrostatic system for mobile machinery according to claim 2, characterized in that: It also includes a radiator (8), the input end of the radiator (8) is connected to the overflow end of the electric proportional overflow valve and the second overflow valve.

7. A method for controlling a static hydraulic system of a walking machine according to any one of claims 1 to 6, characterized in that: Including steps: S1. Get the handle value; S2. If the handle value is less than 0, or the handle value is greater than or equal to 0 and the engine feedback load rate is greater than a, then the target current I of the electric proportional relief valve in the first control cover (7) is set. f =0, at this time the electric proportional relief valve is set to the maximum value; S3. If the handle value is ≥ 0 and the engine feedback load rate is < a, the target current I of the electric proportional relief valve in the first control cover plate (7) is calculated based on the current displacement value of the variable piston pump (1) and the rated braking torque T of the engine. f Real-time control of the electric proportional relief valve.

8. The control method according to claim 7, characterized in that: The step S3 specifically includes the following steps: S31, calculating the current displacement V of the variable piston pump (1) according to its operating parameters p ; S32, based on the rated braking torque T of the engine and the current displacement V of the variable displacement piston pump (1) p Calculate the target pressure value P of the brake pump inlet p : S33, according to the target pressure value P p The target current I of the electric proportional overflow valve in the first control cover plate (7) is calculated f Real-time control of the electric proportional relief valve: I f =f1(P p )。 9. The control method according to claim 8, characterized in that: In the step S31, the current displacement V of the variable piston pump (1) is calculated by combining the current control current feedback value of the variable piston pump (1) with the displacement-current curve of the variable piston pump (1). p =f2(I p ).

10. The control method according to claim 8, characterized in that: In step S31, the inclination angle of the swash plate of the variable piston pump (1) is obtained by an angle sensor, and the current displacement V of the variable piston pump (1) is calculated based on the displacement-inclination curve of the variable piston pump (1). p =f3(θ p ).

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