Walking mechanical hydrostatic system and control method thereof

Abstract

The invention discloses a walking mechanical hydrostatic system and a control method thereof. The hydrostatic system comprises a variable displacement piston pump, an oil supplementing pump, a first balance valve, a large-flow one-way valve, a pressure reduction main valve element, a first control cover plate, an overflow main valve element, a second control cover plate, a variable displacement piston motor and a second balance valve. The control method comprises the following steps: S1, acquiring a handle value; s2, if the handle value is smaller than 0, or the handle value is larger than or equal to 0, and the engine feeds back the load rate gt; if yes, the target current If of the electric proportional overflow valve in the first control cover plate is set to be 0, and the electric proportional overflow valve is set to be the maximum value; and S3, if the handle value is larger than or equal to 0 and the feedback load rate of the engine is smaller than a, the target current If of the electric proportional overflow valve is obtained through calculation according to the current displacement value of the variable displacement piston pump and the rated braking torque T of the engine, and the electric proportional overflow valve is controlled in real time. The technical problems that system heating is serious, the response speed is low, potential safety hazards exist, and elements are prone to being damaged are solved.

Description

Technical Field

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

[0002] The static hydraulic system has the advantages of good micro-motion, high speed stiffness, high transmission efficiency, easy starting, braking, reversing operation, and easy electro-hydraulic compound control. In recent years, with the development of the hydraulic industry, the reliability of hydraulic components has been continuously improved, and the cost has been continuously reduced. Fully hydraulically driven engineering machinery has appeared at home and abroad.

[0003] With the rapid improvement of China's engineering mechanization level in recent years, customers' requirements for the stability and safety of machinery have also increased accordingly, and the corresponding national standards have also made strict provisions on the safe braking distance of engineering machinery; the existing engineering machinery travel drive part is mainly braked by the static hydraulic system, but some problems have also arisen in the braking process: high-speed braking or long downhill conditions The engine is dragged back and overspeeded, the pressure shock on the braking side is too large to cause component damage, and the forward side cannot replenish oil in time, causing system pressure loss; in response to these problems, relevant technicians have proposed some solutions, and the main improvement technical forms include the following: the series relief valve on the braking side changes the return oil back pressure, the parallel relief valve on the braking side reduces the circuit impact pressure, and the variable pump swash plate angle is controlled to return to zero instantly when the engine is detected to be abnormally overspeed.

[0004] The existing technology currently has the following deficiencies:

[0005] The closed system variable piston pump itself integrates a bidirectional high-pressure relief valve and an oil replenishment 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 flow capacity limitations of the high-pressure relief valve and the oil replenishment check valve, the pressure generated on the braking side usually far exceeds the set value of the relief valve. The resulting ultra-high pressure causes engine overspeed and component damage. At the same time, the loss of pressure on the other side of the system causes problems such as pump and motor knocking. The existing technology mainly proposes corresponding solutions based on this actual situation, such as:

[0006] 1. The patent "Brake control system and control method for closed hydraulic drive system - CN107323446A" connects a relief valve in series on the brake side. When high pressure is generated on the brake side, the return oil back pressure is changed through the relief valve in series on the brake side, and then enters the forward side through the one-way valve. The entire kinetic energy is absorbed by the relief valve in series, and does not act on the hydraulic pump, thereby ensuring that the engine will not be overspeeded due to reverse drag;

[0007] (1) This solution avoids engine overspeeding by eliminating the effect of brake high-pressure oil on the pump. However, the vehicle's kinetic energy is completely absorbed by the overflow valve and the engine's own braking torque is not utilized, resulting in serious overflow heating of the system.

[0008] (2) The system uses the feedback value of the pressure sensor to determine whether the engine is overspeeding, and then adjusts the control pressure in a closed loop to meet its set value. After the relief valve is opened, most of the flow on the braking side will pass through the valve. Due to the influence of the pressure gain of the relief valve, the actual pressure may far exceed the set pressure, and the electric proportional relief valve needs to be readjusted. Because the braking condition time is relatively short, repeated adjustments greatly reduce the response speed of the system;

[0009] 2. The patented "Hydrostatic Braking Solution - CN105546077A" reduces the pressure on the braking side of the motor by connecting a relief valve in parallel with the motor, which can reduce the pressure shock on the braking side to a certain extent and avoid reverse drag due to exceeding the engine braking torque. However, while the pressure is reduced, the reverse braking torque of the hydrostatic system on the motor is also reduced, resulting in an increase in the braking distance of the vehicle and a reduction in the safety of the vehicle.

[0010] 3. The patent "Combined Braking Device of Engine and Hydrostatic System - CN102416942A" adds an electromagnetic reversing valve to the variable mechanism circuit of the variable piston pump. When the engine overspeed is detected, the variable pump swash plate is controlled to return to zero instantly:

[0011] (1) This method can instantly reduce the reverse torque between the pump and the engine, but it will cause a sudden increase in the oil on the brake side, resulting in ultra-high pressure and instantaneous pressure loss on the forward side. In severe cases, it will cause damage to components such as the pump and motor.

[0012] (2) Instantaneously cutting off the pump's displacement can easily cause the vehicle to brake rapidly at a very large deceleration, which is beyond the driver's control and poses a very serious safety hazard. Summary of the invention

[0013] On one hand, the present application provides a static hydraulic system for a walking machine to solve the technical problems of the prior art such as severe heating, slow response speed, potential safety hazards, and easily damaged components.

[0014] The technical solutions adopted in this application are as follows:

[0015] A static hydraulic system for a walking machine includes a variable piston pump, an oil replenishment 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, wherein:

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

[0017] The pressure reducing main valve core comprises a two-way cartridge pressure reducing valve, the first control cover plate comprises an electric proportional relief valve, the A port of the two-way cartridge pressure reducing valve is connected to the B port of the variable piston pump, the B port of the two-way cartridge pressure reducing valve is connected to the B port of the variable piston motor, 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 hole, and the output end of the electric proportional relief valve is connected to the control ports of the variable piston pump and the variable piston motor;

[0018] The input end of the large flow check valve is connected to the B port of the variable piston pump, and at the same time, is connected to the input end of the electric proportional relief valve through the damping hole, and the output end of the large flow check valve is connected to the B port of the variable piston motor;

[0019] The overflow main valve core includes a two-way cartridge overflow valve, the second control cover plate includes a second overflow valve, the A port of the two-way cartridge overflow valve is connected to the B port of the variable plunger motor, the B port of the two-way cartridge overflow valve is connected to the A port of the variable plunger motor, the control port of the two-way cartridge 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 through another damping hole, and the output end of the second overflow valve is connected to the control port of the variable plunger motor.

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

[0021] Preferably, a filtering device is also provided at the output end of the oil replenishment pump.

[0022] Preferably, the large-flow one-way valve adopts a two-way cartridge one-way valve, the A port of the two-way cartridge one-way valve is connected to the B port of the variable piston pump, and at the same time, it is connected to the input end of the electric proportional relief valve through a damping hole, the B port of the two-way cartridge one-way valve is connected to the B port of the variable piston motor, and the control port of the two-way cartridge one-way valve is connected to the B port of the two-way cartridge one-way valve through a damping hole.

[0023] Preferably, the large-flow one-way valve is a tubular one-way valve.

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

[0025] On the other hand, the present application also provides a control method for the static hydraulic system of the walking machine, comprising the steps of:

[0026] S1. Get the 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, then the target current I of the electric proportional relief valve in the first control cover is set.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 comprises the steps of:

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

[0031] S32, according to the rated braking torque T of the engine and the current displacement V of the variable displacement piston pump p Calculate the brake side pump inlet target pressure value P 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 plate 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 prior art, this application has the following beneficial effects:

[0039] The present application provides a static hydraulic system for a walking machinery and a control method thereof. The present application avoids the problems of emergency braking and engine overspeeding during long downhill conditions with a faster response speed by designing a hydraulic circuit and system control logic and adopting a smaller number of commonly used components while ensuring the braking torque. The present application adds a high-pressure, high-flow relief valve at the outlet of a variable piston motor on the braking side, which limits the maximum impact pressure on the braking side and protects the hydraulic components while ensuring the braking distance. The present application compensates for the deficiencies in the flow capacity of the pump's own high-pressure relief valve, the oil replenishment check valve and the oil replenishment capacity of the oil replenishment pump through a high-pressure, high-flow relief valve. The oil overflowing from the braking side can be smoothly replenished to the forward side, avoiding the problem of pump and motor knocking caused by pressure loss on the forward side. The present application adds a high-pressure, high-flow pressure reducing valve at the inlet of the brake side pump to limit the inlet of the brake side pump. of pressure to ensure that the anti-drag torque of the pump on the engine does not exceed the rated braking torque of the engine at all times, thereby avoiding engine overspeed; the hydraulic system of the present 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 relief valve is calculated through the operating parameters of the pump and the rated braking torque of the engine, and the electric proportional relief valve is directly controlled in real time, thereby avoiding repeated adjustments of the system due to factors such as the lag of pressure feedback and pressure gain, and improving the response speed of the system; when the engine is in a critical overspeed condition, the present application controls the anti-drag torque under the action of the high pressure at the inlet of the brake side pump to be basically equal to the rated braking torque of the engine, thereby avoiding engine overspeed and making full use of the engine's own braking energy. The excess capacity is absorbed in the form of heat generated by the hydraulic system, thereby reducing system heat generation and energy loss.

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

[0041] The drawings constituting a part of the present application are used to provide a further understanding of the present application. The illustrative embodiments and descriptions of the present application are used to explain the present application and do not constitute an improper limitation on the present application. In the drawings:

[0042] Figure 1 This is a schematic diagram of the composition principle of the static hydraulic system of a walking machine in a preferred embodiment of the present application;

[0043] Figure 2 It is a flow chart of the control method of the preferred embodiment of the present application;

[0044] Figure 3 It is a schematic diagram of the sub-step flow of step S3 of the preferred embodiment of the present application;

[0045] Figure 4 is a schematic diagram of the sub-step flow of step S3 of another preferred embodiment of the present application;

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

[0048] In the figure: 1. variable piston pump; 2. oil replenishment pump; 3. first balancing valve; 4. first overflow valve; 5. large flow check valve; 6. pressure reducing main valve core; 7. first control cover plate; 8. radiator; 9. overflow main valve core; 10: second control cover plate; 11: variable piston motor; 12. second balancing valve; 13. plug-in type electric proportional overflow valve; 14. plug-in type pressure reducing valve; 15. large flow plug-in type or tubular check valve; 16. plug-in type overflow valve. DETAILED DESCRIPTION

[0049] The embodiments of the present application are described in detail below in conjunction with 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 walking machine static hydraulic system, including a variable piston pump 1, a 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, an overflow 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 replenishment 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 replenishment 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, the first control cover plate includes an electric proportional relief valve, the A port of the two-way cartridge pressure reducing valve is connected to the B port of the variable piston pump 1, the B port of the two-way cartridge pressure reducing valve is connected to the B port 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 hole, 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, and 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 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, and 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.

[0055] 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 piston pump 1 to ensure that the output pressure of the oil replenishment pump 2 is maintained at a set value.

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

[0057] The large flow check valve 5 adopts a two-way cartridge check valve, the A port of the two-way cartridge check valve is connected to the B port of the variable plunger pump 1, and at the same time, is connected to the input end of the electric proportional relief valve through a damping hole, the B port of the two-way cartridge check valve is connected to the B port of the variable plunger motor 11, and the control port of the two-way cartridge check valve is connected to the B port of the two-way cartridge check valve through a damping hole. Preferably, the large flow check valve 5 adopts a tubular check valve.

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

[0059] In the above embodiment, the variable piston pump 1 is the power source of the closed walking system, and the swash plate can be bidirectionally variable; the oil replenishment pump 2 is a gear pump. In this closed system, the main functions of the oil in the oil replenishment pump 2 are: replenishing oil on the low-pressure side of the closed system, flushing the motor housing, and controlling the variable of the swash plate of the variable piston pump 1; the first balancing valve 3 and the second balancing valve 12 are high-pressure relief valves, and their main function is that when the oil pressure on one side of the closed system is higher than the set value of the relief valve, the relief valve opens, and the oil on the high-pressure side enters the low-pressure side, and the pump inlet and outlet circulate their own oil; the first relief valve 4 is an oil replenishment relief valve, which controls the outlet pressure of the oil replenishment pump; the large-flow check valve 5 is a large-flow two-way plug-in check valve, and only when the vehicle is moving backward, the large-flow check valve 5 opens, and the oil passes through the large-flow check valve 5 into the main The oil circuit is used to prevent the high pressure at the B port of the variable piston pump 1 from causing the electric proportional relief valve to tend to the closed state when retreating, and the vehicle cannot move; the pressure-reducing main valve core 6 is a two-way plug-in valve core with a pressure-reducing function, which limits the pressure at the B port of the variable piston pump 1 on the braking side to prevent the pump from generating a large reverse torque and causing the engine to overspeed; the first control cover plate 7 is a pilot control module with an electric proportional relief valve, which adjusts the set value of the electric proportional relief valve in real time according to the pump displacement feedback, so that the pressure at the B port of the pump is always higher than the set value of the electric proportional relief valve by the main valve core spring pressure; when the power is off, the electric proportional relief valve has a maximum set pressure; the radiator 8 dissipates the heat of the small flow hot oil after the internal leakage of the pump and the motor, the flushing flow, and the overflow of the electric proportional relief valve and the second relief valve control oil circuit;

[0060] The overflow main valve core 9 is a two-way plug-in valve core with overflow function, which mainly limits the outlet pressure of the brake side motor. On the one hand, it ensures that there is sufficient braking force at the load end and shortens the braking distance; on the other hand, it reduces the pressure shock between the motor outlet and the components to protect the hydraulic components. At the same time, the larger flow capacity enables the oil to be smoothly replenished to the forward side of the system, avoiding pressure loss on the forward side; the second control cover plate 10 is a pilot control module with a direct-acting overflow valve, and the valve core spring force setting value determines the maximum outlet pressure of the brake side motor; the variable piston motor 11 is a traveling axle drive motor, which provides drive and braking torque for the whole vehicle.

[0061] The present embodiment provides a static hydraulic system for a walking machine, including a variable piston pump 1, an oil replenishment 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. The present embodiment, through the design of the hydraulic circuit and the system control logic, uses a relatively small number of commonly used components to avoid the problems of emergency braking and engine reverse drag overspeed in long downhill conditions with a relatively fast response speed while ensuring the braking torque; the present embodiment adds a high-pressure and large-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, and protects 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 replenishment check valve, and the oil replenishment capacity of the oil replenishment pump through the high-pressure and large-flow overflow valve, and the oil overflow on the braking side can be smoothly replenished to On the forward side, the pump and motor knocking caused by the loss of pressure on the forward side is avoided; in this embodiment, a high-pressure and high-flow pressure reducing valve is added to the brake side pump inlet to limit the pressure of the brake side pump inlet, ensuring that the reverse drag torque of the pump on the engine does not exceed the rated braking torque of the engine at all times, thus avoiding engine overspeed; in this application, when the engine is in critical overspeed condition, the reverse drag torque under the high pressure of the brake side pump inlet is controlled to be substantially equal to the rated braking torque of the engine, thus avoiding engine overspeed and making full use of the engine's own braking energy. The excess capacity is absorbed in the form of heat generation by the hydraulic system, thus reducing system heat generation and energy loss. In this embodiment, by setting a radiator 8, the small-flow high-temperature oil after the overflow of the electric proportional overflow valve in the first control cover plate 7 and the second overflow valve in the second control cover plate 10 is led back to the radiator 8, which can increase the system flushing flow, ensure the system thermal balance, and protect the hydraulic system and components.

[0062] like Figure 2 Another preferred embodiment of the present application further provides a control method for the static hydraulic system of the 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 a reverse working condition, and since the reverse speed is relatively low, there is rarely a problem of high brake pressure or engine back-drag), or 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, normal forward (non-downhill) working condition), 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 brake 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 through the operating parameters of the pump and the rated braking torque of the engine, and the electric proportional relief valve is directly controlled in real time, avoiding repeated adjustments of the system due to the lag of pressure feedback, pressure gain and other factors, thereby improving the response speed of the system.

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

[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 brake side pump inlet target pressure value P p :

[0070]

[0071] S303, according to the target pressure value P p The target current I of the electric proportional relief valve in the first control cover plate 7 is calculated 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 steps of:

[0074] S311, obtaining 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 rated braking torque T of the engine and the current displacement V of the variable displacement piston pump 1 p Calculate the brake side pump inlet target pressure value P p :

[0076]

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

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

[0079] In reality, engine back-drag occurs in two situations:

[0080] The first is a long downhill condition. If only hydrostatic braking is used, the vehicle will accelerate downhill due to its large inertia and the torque generated exceeds the engine's braking torque. In this case, mechanical friction can generally be used to assist braking, but long-term mechanical friction accelerates 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 piston pump's own high-pressure relief valve, the replenishment check valve, and the replenishment capacity of the replenishment pump, the brake side oil cannot overflow to the forward side instantly, resulting in a sharp increase in the brake side pressure. When the torque under the combined action of the high pressure on the brake side and the pump is greater than the rated braking torque of the engine, the engine is reversed and overspeeded.

[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 of the pump and motor and at port B of the pump and motor are basically the same.

[0083] (1) Emergency braking, long downhill working 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 overflow valve integrated on the second control cover 10. The oil opens the direct-acting overflow 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. 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 to the B port through the A port of the overflow main valve core 9 and enters the forward side oil circuit between the pump A port and the motor A port.

[0084] Furthermore, due to the larger flow capacity of the overflow main valve core 9, a lower pressure gain will be generated, and the pressure at port B 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, thereby 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-reducing main valve core 6 is limited, which also protects the electric proportional overflow valve.

[0085] Furthermore, due to the large flow capacity of the overflow main valve core 9, the oil at the B port of the variable piston motor 11 overflows smoothly to the forward side, avoiding the pressure loss between the pump A port and the motor A port on the forward side caused by the insufficient flow capacity of the first balancing valve 3 and the one-way valve of the variable piston pump 1 and the limited oil replenishment capacity of the oil replenishment pump 2, avoiding the knocking problem of the variable piston pump 1 and the variable piston motor 11, and extending the life of the components.

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

[0087] Furthermore, the engine's current load rate feedback is used to determine whether the engine is approaching an overspeed condition. When the load rate is higher than the set value, the engine is not overspeeding, the electric proportional relief valve of the first control cover plate 7 loses power, the pressure-reducing main valve core 6 is set to the maximum value, the main valve core 6 is in a fully open state, and the system pressure is low and cannot trigger the pressure reduction function; when the load rate is lower than the set value, it is determined that the engine has tended to an overspeed condition, and the current displacement of the pump is calculated based on the displacement control current feedback value of the variable piston pump 1; then the limit value of the brake side pump inlet B pressure is calculated based on the engine rated braking torque and the current displacement of the pump, and then the target current value of the electric proportional relief valve of the first control cover plate 7 is calculated. , the controller directly outputs the target current value to the electric proportional relief valve; when the pressure at the outlet A of the pressure reducing main valve core 6 is higher than the set value of the electric proportional relief valve of the first control cover plate 7, the oil opens the relief valve through the control port X of the first control cover plate 7, flows out through the Y port to reach the radiator 8, and the flowing oil generates a pressure difference through the damping hole, so that the pressure reducing main valve core 6 is in a dynamic balance adjustment state, ensuring that the pressure at the outlet A of the pressure reducing main valve core 6 is always higher than the set pressure of the electric proportional relief valve in the first control cover plate 7 by the main valve core spring force, and the excess high-pressure energy of the B port of the pressure reducing main valve core 6 is lost in the form of heat under the action of the valve core, which effectively avoids the brake high pressure from dragging the engine overspeed through the pump;

[0088] Furthermore, the pressure of the electric proportional overflow valve of the first control cover plate 7 is adjusted to control the reverse drag torque of the engine by the high pressure at the inlet port B of the brake side pump to match the braking torque of the engine, thereby avoiding the engine overspeed and making full use of the engine's own braking energy; the excess kinetic energy of the vehicle is then consumed in the form of heat energy of the hydraulic system; this method can minimize the heating of the hydraulic system and reduce energy loss;

[0089] (3) Emergency braking and long downhill conditions: To further reduce the heat generation of the hydraulic system and protect the hydraulic components, the hot oil flowing out through the Y port after the overflow of the first control cover plate 7 and the second control cover plate 10, the housing oil drain 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 rate after the overflow of the control oil circuit is low and has no effect on the system oil replenishment), thereby increasing the flushing flow of the closed system and ensuring the thermal balance of the loop.

[0090] (4) Reverse working condition: When the vehicle is reversing, the high pressure at port B of the variable piston pump 1 acts on port A of the pressure reducing main valve core 6, causing the opening of the pressure reducing main valve core 6 to decrease. In order to prevent the high pressure oil at port B of the variable piston pump 1 from causing the pressure reducing main valve core 6 to tend to be closed and the vehicle cannot be reversed normally, a large flow check valve 5 is connected in parallel with the pressure reducing main valve core 6. The high pressure oil at port B of the variable piston pump 1 enters the main oil circuit between port B of the large flow check valve 5 and port B of the variable piston motor 11 through port A of the large flow check valve 5, and the motor rotates in the reverse direction.

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

[0092] 1. A high-pressure and high-flow relief valve is added to the motor outlet on the brake side to reduce the impact pressure at the motor outlet. While ensuring the braking torque, it protects the hydraulic components and extends the service life of the components.

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

[0094] 3. Add a high-pressure and high-flow pressure reducing valve at the brake pump inlet. When the engine is in critical overspeed condition, the brake pump inlet pressure is adjusted in real time to ensure that the reverse torque generated by the variable piston pump is always less than the rated braking torque of the engine, thus avoiding engine overspeed.

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

[0096] 5. When the engine is in a critical overspeed state, the reverse drag torque under the high pressure of 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 generation in the hydraulic system, reducing system heat generation 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 large-flow two-way plug-in non-return valve is connected in parallel with the electric proportional pressure reducing valve on the braking side. The oil on the braking side is reversely cut off during braking or long downhill conditions. When the vehicle moves backward, the oil opens the non-return valve to drive the motor, preventing high-pressure oil from passing through the electric proportional pressure reducing valve in the reverse direction when the vehicle moves backward, causing the pressure reducing main valve core to show a closing trend and the vehicle cannot 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 plug-in type large flow pressure reducing valve. In this embodiment, the two-way plug-in type large flow pressure reducing valve is replaced with a plug-in type hydraulic valve 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 plug-in check valve is replaced with a large flow plug-in type or tubular check valve 15, and the overflow main valve core 9 and the second control cover plate are combined into a two-way plug-in type large flow relief valve, which 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, and the insert type is an integral valve, the insert type occupies less space and is easier to maintain and disassemble; the degree of integration is higher; 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 parts not described in detail in this application are all known technologies to those skilled in the art. Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of this application and are not limiting. Although this application is described in detail with reference to the preferred embodiments, those skilled in the art should understand that the technical solutions of this application can be modified or replaced by equivalents without departing from the purpose and scope of the technical solutions of this application, which should be included in the scope of the claims of this application.

Claims

1. A static hydraulic system for a walking machine, characterized in that: The invention comprises a variable piston pump (1), an oil replenishment 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 replenishment 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 replenishment 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) comprises a two-way cartridge pressure reducing valve, the first control cover plate comprises an electric proportional relief valve, the A port of the two-way cartridge pressure reducing valve is connected to the B port of the variable piston pump (1), the B port of the two-way cartridge pressure reducing valve is connected to the B port 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 hole, 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); 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 a damping hole, and the output end of the large flow check valve (5) is connected to the B port of the variable piston 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 piston motor (11), the B port of the two-way plug-in overflow valve is connected to the A port of the variable piston 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 piston motor (11) through another damping hole, and the output end of the second overflow valve is connected to the control port of the variable piston motor (11).

2. The static hydraulic system for walking 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 piston pump (1).

3. The static hydraulic system for walking 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 static hydraulic system for walking machinery according to claim 2, characterized in that: The large flow check valve (5) adopts a two-way plug-in check valve, the A port of the two-way plug-in check valve 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 a damping hole, the B port of the two-way plug-in check valve is connected to the B port of the variable piston 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 static hydraulic system for walking machinery according to claim 2, characterized in that: The large flow one-way valve (5) is a tubular one-way valve.

6. The static hydraulic system for walking machinery according to claim 2, characterized in that: It also comprises a radiator (8), the input end of the radiator (8) being connected to the overflow end of the electric proportional overflow valve and the second overflow valve.

7. A control method for a static hydraulic system of a walking machine according to any one of claims 1 to 6, characterized in that: Includes 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 plate (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 comprises the following steps: S31, calculating the current displacement V of the variable displacement piston pump (1) according to its operating parameters p ; S32, 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 brake side pump inlet target pressure value P p : S33, according to the target pressure value P p The target current I of the electric proportional relief valve in the first control cover plate (7) is calculated and obtained. 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 in combination with the displacement-inclination curve of the variable piston pump (1). p =f3(θ p ).

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