WO2014110901A1 - 基于合流控制方式的液压装置 - Google Patents

基于合流控制方式的液压装置 Download PDF

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Publication number
WO2014110901A1
WO2014110901A1 PCT/CN2013/081502 CN2013081502W WO2014110901A1 WO 2014110901 A1 WO2014110901 A1 WO 2014110901A1 CN 2013081502 W CN2013081502 W CN 2013081502W WO 2014110901 A1 WO2014110901 A1 WO 2014110901A1
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WO
WIPO (PCT)
Prior art keywords
valve
confluence
pilot pressure
load sensing
passage
Prior art date
Application number
PCT/CN2013/081502
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English (en)
French (fr)
Inventor
汪立平
Original Assignee
江苏恒立高压油缸股份有限公司
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Filing date
Publication date
Application filed by 江苏恒立高压油缸股份有限公司 filed Critical 江苏恒立高压油缸股份有限公司
Priority to US14/761,101 priority Critical patent/US9988792B2/en
Priority to JP2015552978A priority patent/JP6257647B2/ja
Priority to EP13871529.7A priority patent/EP2947331B1/en
Publication of WO2014110901A1 publication Critical patent/WO2014110901A1/zh

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Classifications

    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2225—Control of flow rate; Load sensing arrangements using pressure-compensating valves
    • E02F9/2228—Control of flow rate; Load sensing arrangements using pressure-compensating valves including an electronic controller
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2232—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps
    • E02F9/2235—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps including an electronic controller
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B11/00—Servomotor systems without provision for follow-up action; Circuits therefor
    • F15B11/16—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors
    • F15B11/17—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors using two or more pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/0416—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor with means or adapted for load sensing
    • F15B13/0417—Load sensing elements; Internal fluid connections therefor; Anti-saturation or pressure-compensation valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/044—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor operated by electrically-controlled means, e.g. solenoids, torque-motors
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B15/00—Fluid-actuated devices for displacing a member from one position to another; Gearing associated therewith
    • F15B15/08—Characterised by the construction of the motor unit
    • F15B15/14—Characterised by the construction of the motor unit of the straight-cylinder type
    • F15B15/1423—Component parts; Constructional details
    • F15B15/1466—Hollow piston sliding over a stationary rod inside the cylinder
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/205—Systems with pumps
    • F15B2211/20576—Systems with pumps with multiple pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/3056—Assemblies of multiple valves
    • F15B2211/3059—Assemblies of multiple valves having multiple valves for multiple output members
    • F15B2211/30595—Assemblies of multiple valves having multiple valves for multiple output members with additional valves between the groups of valves for multiple output members
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/71—Multiple output members, e.g. multiple hydraulic motors or cylinders
    • F15B2211/7142—Multiple output members, e.g. multiple hydraulic motors or cylinders the output members being arranged in multiple groups

Definitions

  • the invention relates to the technical field of hydraulic control, in particular to a hydraulic device for realizing confluence control of a constant flow throttle speed regulation hydraulic system and a load sensing control hydraulic system.
  • this "constant power" controlled load sensing hydraulic system has a hydraulic motor that drives large mass rotation in the actuator.
  • the actuator needs to overcome the large inertia, the movement is very slow, and the required oil flow is required.
  • the hydraulic motor with a large external load the rotation is relatively slow at the beginning, the load pressure of the hydraulic motor will rise sharply to a very high value, and the variable pump controls the adjustment of the oil pipeline according to the highest load pressure. The pressure is higher than the highest load.
  • the pressure in the oil pipeline directly acts on the constant power control valve, which makes the displacement of the variable displacement piston pump smaller, resulting in slow operation of all actuators, low production efficiency, and power source. The energy loss is large.
  • the technical problem to be solved by the present invention is: to overcome the deficiencies in the prior art, and to provide an efficient, The hydraulic device that realizes the confluence control of the constant flow throttling speed regulating hydraulic system and the load sensing control hydraulic system with low energy consumption.
  • a hydraulic device based on a combined flow control mode comprising a load sensing unit and a throttle speed adjusting unit, wherein the load sensing unit has a first reversing valve and a second reversing valve a valve, the throttle speed regulating unit has a fourth reversing valve, and a parallel valve and a check valve connecting the load sensing unit and the throttle speed regulating unit are arranged on the parallel oil path disposed in parallel with the fourth reversing valve.
  • the confluence valve has a confluence channel for controlling the parallel oil passage to open and close to the load sensing unit to divert the throttle speed regulating unit fluid, and the fourth reversing valve is connected with the fourth actuator for realizing the reversing valve reversal during the operation.
  • the first pilot pressure and the second pilot are controlled when a reversing valve is reversed by its first pilot pressure, the second diverter valve is subjected to its second pilot pressure, and the fourth diverter valve is acted upon by its fourth pilot pressure
  • the pressure is also applied to the confluence valve separately or simultaneously to change the position of the confluence passage to realize the reversal of the confluence valve.
  • the load sensing unit further includes a constant power control valve, a variable displacement mechanism and a variable displacement pump, wherein the first reversing valve is respectively connected with a first compensating valve and a first actuator, and the second reversing valve is respectively connected There is a second compensation valve and a second actuator; the throttle speed adjustment unit further includes a gear pump coaxial with the variable displacement piston pump.
  • the confluent channel includes a disconnecting channel for controlling the on/off of the parallel oil passage, a large liquid resistance passage, and a small liquid resistance passage, and one end of the confluent valve has: a large end face that synchronously receives the first pilot pressure control, and synchronously receives the second pilot.
  • the small end face of the pressure control, the other end of the confluence valve is provided with a return spring, and the fourth reversing valve receives the fourth pilot pressure control and is connected in parallel with the confluence valve.
  • the channel area of the disconnecting channel is zero, the channel area of the large liquid resistance channel and the small liquid resistance channel is not zero, and the channel area of the large liquid resistance channel is larger than the channel area of the small liquid resistance channel.
  • the beneficial effects of the present invention are as follows:
  • the present invention connects the load sensing unit and the throttle speed regulating unit by providing a confluence valve, so that the fluid damping formed by the confluence passage of the confluence valve and the highest displacement of the actuator in the load sensing unit
  • the loads are matched so that it does not affect the actuator operation in the throttle unit.
  • the flow rate of the throttle speed regulating unit can be shunted to the load sensing unit in time, and when the load sensing unit is started to be used alone, the pressure rises sharply due to the need to overcome the inertia of the external load of a large mass, and the load sensing unit
  • the actuators are slow in operation, low in efficiency, and depleted in the energy of the hydraulic motor, thereby achieving high efficiency and low energy consumption of the system.
  • Figure 1 is a schematic view of the structure of the present invention.
  • FIG. 2 is an enlarged schematic view of the confluence valve shown at B in Figure 1.
  • a hydraulic device based on the merge control method shown in Figs. 1 and 2 is used in an embodiment of a hydraulic excavator.
  • the hydraulic device comprises a pressure sensing load sensing unit, a throttle speed regulating unit having a bypass port constant flow rate, and a confluence valve 5 connecting the load sensing unit and the throttle speed regulating unit. And check valve 5.
  • the load sensing unit includes a constant power control valve 8, a variable displacement mechanism 9, a variable displacement piston pump 10 to which the engine 16 is connected, a first reversing valve 1, a second reversing valve 2, and a fifth reversing valve 17
  • the first reversing valve 1, the second reversing valve 2 and the fifth reversing valve 17 are respectively connected with a corresponding first compensating valve 11, a first actuating element 12, a second compensating valve 13, and a second actuating element 14,
  • the fifth compensating valve 18 and the fifth actuating element 19 the first reversing valve 1 receives the first pilot pressure P1 provided by the outside to change direction, and the second reversing valve 2 receives the second pilot pressure P2 provided by the outside to change Toward, the fifth switching valve 17 receives the fifth pilot pressure P5 supplied from the outside to be reversed, and the relief valve 20 is provided on the oil passage at the front end of the constant power control valve 8.
  • the throttle regulating unit includes a fourth reversing valve 3, a sixth reversing valve 21, a gear pump 15 coaxial with the variable displacement piston pump 10, and a fourth reversing valve 3 connected with a corresponding fourth actuator 7
  • the sixth reversing valve 21 is connected to a corresponding sixth actuator 22.
  • the fourth reversing valve 3 receives the fourth pilot pressure P4 provided by the outside to change direction
  • the sixth reversing valve 21 receives the sixth pilot pressure P6 provided by the outside to change direction.
  • the confluence valve 5 is disposed on the parallel oil passage 4 connected in parallel with the fourth reversing valve 3 and communicates with the outlet of the variable displacement piston pump 10.
  • the confluence valve 5 has a control to connect the parallel oil passage 4 to the load sensing unit.
  • Dividing the junction channel 50 of the throttle governing unit fluid, the joining passage 50 includes a breaking passage 51, a large liquid resistance passage 52, and a small liquid resistance passage 53, wherein the passage area of the breaking passage 51 is zero, the large liquid resistance passage 52 and The passage area of the small liquid resistance passage 53 is not zero, and the passage area of the large liquid resistance passage 52 is larger than the passage area of the small liquid resistance passage 53.
  • the merging valve 5 adopts a pilot pressure control mode, and has two pilot control end faces at the end of the merging valve 5: a large end face 54 communicating with the first pilot pressure P1 at one end of the first directional control valve 1 and the second directional control valve 2
  • the small end face 55 of the second pilot pressure P4 of the end is connected, and the other end of the confluence valve 5 is provided with a return spring 56, and the merge valve 5 is connected with the fourth reversing valve 3.
  • the confluence valve 5 When the large end surface 54 of the confluence valve 5 has a hydraulic pressure, the confluence valve 5 can be positioned at the large liquid resistance passage 52, and the small end surface 55 of the confluence valve 5 has a hydraulic pressure, so that the confluence valve 5 can be positioned at the small liquid resistance passage 53.
  • the small end faces 54, 55 are all available after hydraulic pressure. In order to make the confluence valve 5 at the position of the large liquid resistance passage 52, if there is no hydraulic pressure in the large and small end faces 54, 55, the confluence valve 5 can be positioned in the disconnecting passage 51.
  • the combine valve 5 is realized by the common or separate action of the first pilot pressure P1 and the second pilot pressure P2 in the disconnecting passage 51, the large liquid resistance passage 52, and The small liquid resistance channel 53 is transposed to connect the load sensing unit and the throttle speed regulating unit, and the majority of the fluid of the throttle speed regulating unit is shunted and then input to the load sensing unit through the combining valve 5 and the check valve 6.
  • the fluid of the fourth actuator 7 is shunted in time, and the oil in the load sensing unit and the throttle unit is ensured under the premise that the pressure of the fourth actuator 7 is consistent with the external load and the fourth actuator 7 can work normally.
  • the pressure does not rise sharply to a maximum value, avoiding the decrease in the displacement of the constant-power control valve 8 controlled variable displacement piston pump 10 due to the increase in oil pressure, eventually causing all actuators in the system to move slowly, with low production efficiency, power A phenomenon in which the source energy loss is large.
  • the device is completed by the confluence control operation mode of the constant flow throttle control unit and the load sensing unit.
  • the fourth actuator 7 of the throttle speed adjusting unit is operated, the first switching valve 1 of the load sensing unit is subjected to the first pilot pressure P1, and the second switching valve 2 is subjected to the second pilot pressure P2 (any one When the two or the same position is working, the confluence valve 5 is reversed, and the majority of the fluid of the throttle governing unit is diverted and then passed through the confluence valve 5 and the check valve 6 to be input to the load sensing unit, which is embodied in the following three types.
  • the first pilot pressure P1 is input to the first switching valve 1, and the fourth pilot pressure P4 is input to the fourth switching valve 3, so that the first switching valve 1 and the fourth switching valve 3 are reversed.
  • the first pilot pressure P1 also acts on the large end surface 54 of the merging valve 5 at the same time. Since the working area of the large end surface 54 is relatively large, the force on the large end surface 54 of the merging valve 5 is also relatively large. The force can overcome the force of the return spring 56, and the joining passage 50 of the combining valve 5 is reversed from the breaking passage 51 to the large liquid resistance passage 52 having a large passage area, and the fluid of the closing valve 5 located at the end surface of the return spring 56 is freely discharged back to the fuel tank.
  • the fluid of the throttle governing unit is input to the load sensing unit through the large liquid resistance passage 52 of the converging valve 5 and the check valve 6. Same The fluid resistance formed in the large liquid resistance passage 52 matches the external load on the first actuator 12, thereby shunting the fluid on the fourth actuator 7 in time.
  • the second pilot pressure P2 is input on the second directional control valve 2, and the fourth pilot pressure P4 is input to the fourth directional control valve 3, so that the second directional control valve 2 and the fourth directional control valve are reversed.
  • the second pilot pressure P2 also acts on the small end face 55 of the confluence valve 5, the small end face 55 has a relatively small active area, and the force acting on the small end face 55 is relatively small, but the force of the return spring 56 can be overcome to make the confluence valve
  • the merged passage 50 of 5 is reversible by the breaking passage 51 to the small liquid resistance passage 53 having a small passage area, and the fluid of the closing valve 5 located at the end face of the return spring 56 is freely discharged back to the tank, and the fluid of the throttle governing unit passes through the confluence valve
  • the small liquid resistance passage 53 and the check valve 6 of 5 are input to the load sensing unit.
  • the fluid resistance formed in the small liquid resistance passage 53 matches the external load on the second actuator 14, so that the fluid of the fourth actuator 7
  • the force acting on the large and small end faces 54, 55 can overcome the force of the return spring 56, so that the merged passage 50 of the combiner valve 5 is reversed from the open passage 51 to the large liquid resistance passage 52 having a large passage area, and the confluence valve 5
  • the fluid at the end face of the return spring 56 is freely discharged back to the tank, and the fluid of the throttle governor unit is input to the load sensing unit through the large liquid resistance passage 52 of the converging valve 5 and the check valve 6. Since the external load on the first actuator 12 is greater than the external load on the second actuator 14, the pressure in the load sensing unit corresponds to the external load on the first actuator 12, so only the merge valve 5 is required.
  • the fluid resistance formed by the large liquid-blocking passage 52 matches the external load on the first actuator 12, and the fluid on the fourth actuator 7 can be shunted out in time.
  • the throttle speed regulating unit can directly discharge the zero pressure without causing energy loss.
  • the actuator of the load sensing unit can avoid the pressure increase, causing the constant power control valve 8 to make the variable displacement piston pump 10
  • the amount of the power source is reduced, and all the actuators in the system are slow to operate, the production efficiency is low, and the energy of the power source is lost.
  • the pressure of the throttle unit can be increased to a large value, but the power source is only supplied to the gear pump. 15 Providing energy without inefficient production.
  • the present invention provides a fluid damper formed by the merging passage 50 flowing through the merging valve 5 to match the highest external load of the actuator in the load sensing unit by providing the merging valve 5 that communicates the load sensing unit with the throttle speed adjusting unit.
  • the fourth actuator 7 in the throttle speed regulating unit is not affected, and the flow rate of the throttle speed regulating unit can be shunted to the load sensing unit in time, so as to avoid the separate application of the load sensing unit when starting work.
  • the large-scale external load inertia causes a sharp rise in pressure, the actuator in the load sensing unit moves slowly, the efficiency is low, and the energy of the engine 16 is lost, thereby achieving high efficiency and low energy consumption of the system.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Mining & Mineral Resources (AREA)
  • Civil Engineering (AREA)
  • Structural Engineering (AREA)
  • Fluid-Pressure Circuits (AREA)
  • Operation Control Of Excavators (AREA)

Abstract

一种基于合流控制方式的液压装置，包括具有第一、第二换向阀（1,2）的负荷传感单元、具有第四换向阀（3）的节流调速单元，在与第四换向阀（3）并联设置的并联油路上，设有连通负荷传感单元与节流调速单元的合流阀（5）和单向阀（6），合流阀（5）具有控制并联油路通断而向负荷传感单元分流节流调速单元流体的合流通道（50），第一换向阀（1）所受第一先导压力（P1）和第二换向阀（2）所受第二先导压力（P2），单独或同时作用在合流阀（5）上使合流通道（50）位置变化而实现合流阀（5）换向；通过设置合流阀（5）连通负荷传感单元与节流调速单元，可以及时向负荷传感单元分流节流调速单元的流量，避免了***内的执行元件动作慢，效率低及损耗液压电机能量的现象，实现***高效率，低能耗的工作。

Description

基于合流控制方式的液压装置 技术领域
本发明涉及液压控制技术领域， 尤其是一种实现定流量节流调速液压*** 与负荷传感控制液压***合流控制的液压装置。
背景技术
定流量节流调速液压***早期广泛应用于各种机械中， 它具有***组成简 单， 元件响应快等优点， 但它的调速特性受到负载影响， 流体总是优先向低负 载执行供油， 为克服这一缺点， US95195425. 3A发明创造了与负载无关的流量分 配控制 （LUDV ) 方式一一负荷传感液压***， 这种***使得各执行机构无论负 载大小怎么不相等， 流入各执行机构的流体流量都可以按各 "需要" 比例地分 配。 同时， 一般的液压机械工作都只需要 "低压大流量， 高压小流量"， 加之其 动力源一般都是有限的， 因此， 在负荷传感液压***中采用 "恒功率"控制可 以充分利用动力源的功率。
但是， 这种 "恒功率"控制的负荷传感液压***， 在执行元件中有一个带 动大质量转动的液压马达， 开始工作时， 执行元件需要克服大惯量， 动作很缓 慢， 所需油液流量很小， 液压马达由于带着大质量的外负载， 开始时转动的比 较慢， 液压马达的负载压力会急剧升高到很高值， 而变量泵依据最高负载压力 来控制调节油路管道中的压力， 该压力比最高负载高出若干值， 油路管道中的 压力直接作用在恒功率控制阀上， 使得变量柱塞泵排量变小， 造成所有执行元 件动作缓慢， 生产效率低， 动力源的能量损失较大。
发明内容
本发明要解决的技术问题是： 克服现有技术中之不足， 提供一种效率高、 能耗少、 实现对定流量节流调速液压***与负荷传感控制液压***进行合流控 制的液压装置。
本发明解决其技术问题所采用的技术方案是： 一种基于合流控制方式的液 压装置， 包括负荷传感单元、 节流调速单元， 负荷传感单元具有第一换向阀和 第二换向阀， 节流调速单元具有第四换向阀， 在与第四换向阀并联设置的并联 油路上， 设有连通负荷传感单元与节流调速单元的合流阀和单向阀， 所述合流 阀具有控制并联油路通断而向负荷传感单元分流节流调速单元流体的合流通 道， 第四换向阀连接有在动作时实现合流阀换向的第四执行元件， 在第一换向 阀受其第一先导压力、 第二换向阀受其第二先导压力、 第四换向阀受其第四先 导压力作用而换向时， 所述第一先导压力和第二先导压力还单独或同时作用在 合流阀上使合流通道位置变化而实现合流阀换向。
所述的负荷传感单元还包括恒功率控制阀、 可变排量机构以及变量柱塞泵， 第一换向阀分别连接有第一补偿阀和第一执行元件， 第二换向阀分别连接有第 二补偿阀和第二执行元件； 节流调速单元还包括与变量柱塞泵同轴的齿轮泵。
具体说， 所述的合流通道包括控制并联油路通断的断路通道、 大液阻通道 以及小液阻通道， 合流阀一端具有： 同步接受第一先导压力控制的大端面、 同 步接受第二先导压力控制的小端面， 合流阀另一端设有复位弹簧， 第四换向阀 接受第四先导压力控制并与合流阀并联连接。
进一步地， 所述的断路通道的通道面积为零， 大液阻通道和小液阻通道的 通道面积不为零， 且大液阻通道的通道面积大于小液阻通道的通道面积。
本发明的有益效果是： 本发明通过设置合流阀连通负荷传感单元与节流调 速单元， 使得流过合流阀的合流通道所形成的流体阻尼与负荷传感单元内的执 行元件的最高外负载相匹配， 这样既不影响节流调速单元内的执行元件工作， 又可以及时向负荷传感单元分流节流调速单元的流量， 避免单独应用负荷传感 单元开始工作时， 由于需克服大质量的外负载惯性而出现压力急剧升高， 负荷 传感单元内的执行元件动作慢， 效率低及损耗液压电机能量的现象， 从而实现 ***高效率， 低能耗的工作。
附图说明
下面结合附图和实施例对本发明进一步说明。
图 1是本发明的结构原理图。
图 2是图 1中 B处显示的合流阀的放大结构示意图
图中 1.第一换向阀 2.第二换向阀 3.第四换向阀 4.并联油路 5.合流阀 50.合流通道 51.断路通道 52 大液阻通道 53.小液阻通道 54.大端面 55.小端面 56.复位弹簧 6.单向阀 7.第四执行元件 8.恒功 率控制阀 9.可变排量机构 10.变量柱塞泵 11.第一补偿阀 12.第一执行 元件 13.第二补偿阀 14.第二执行元件 15.齿轮泵 16.发动机 17.第五 换向阀 18.第五补偿阀 19.第五执行元件 20.溢流阀 21.第六换向阀 22.第六执行元件 P1.第一先导压力 P2.第二先导压力 P3.第三先导压力 P4.第四先导压力 P5.第五先导压力 P6.第六先导压力
具体实施方式
现在结合附图和优选实施例对本发明作进一步的说明。 这些附图均为简化 的示意图， 仅以示意方式说明本发明的基本结构， 因此其仅显示与本发明有关 的构成。
如图 1、 图 2所示的一种基于合流控制方式的液压装置， 使用在液压挖掘机 上的实施例。 该液压装置包括一个具有压力补偿的负荷传感单元、 一个具有旁 通口定流量的节流调速单元、以及连通负荷传感单元与节流调速单元的合流阀 5 和单向阀 5。
负荷传感单元， 包括恒功率控制阀 8、 可变排量机构 9、 连接有发动机 16 的变量柱塞泵 10、 第一换向阀 1、 第二换向阀 2以及第五换向阀 17， 第一换向 阀 1、 第二换向阀 2和第五换向阀 17各自连接有对应的第一补偿阀 11、 第一执 行元件 12、 第二补偿阀 13、 第二执行元件 14、 第五补偿阀 18、 第五执行元件 19， 第一换向阀 1接收外界提供的第一先导压力 P1作用而换向， 第二换向阀 2 接收外界提供的第二先导压力 P2作用而换向， 第五换向阀 17接收外界提供的 第五先导压力 P5作用而换向， 在恒功率控制阀 8的前端油路上设有溢流阀 20。
节流调速单元， 包括第四换向阀 3、 第六换向阀 21、 与变量柱塞泵 10同轴 的齿轮泵 15， 第四换向阀 3连接有对应的第四执行元件 7， 第六换向阀 21连接 有对应的第六执行元件 22。第四换向阀 3接收外界提供的第四先导压力 P4作用 而换向， 第六换向阀 21接收外界提供的第六先导压力 P6作用而换向，
所述的合流阀 5设置在与第四换向阀 3并联的并联油路 4上并与变量柱塞 泵 10的出口相通， 合流阀 5具有控制并联油路 4通断而向负荷传感单元分流节 流调速单元流体的合流通道 50， 该合流通道 50包括断路通道 51、 大液阻通道 52以及小液阻通道 53， 其中， 断路通道 51的通道面积为零， 大液阻通道 52和 小液阻通道 53的通道面积不为零， 且大液阻通道 52的通道面积大于小液阻通 道 53的通道面积。 合流阀 5采用先导压力控制方式， 在合流阀 5—端具有两个 先导控制端面： 即与第一换向阀 1一端的第一先导压力 P1相通的大端面 54、与 第二换向阀 2—端的第二先导压力 P4相通的小端面 55，合流阀 5另一端设有复 位弹簧 56， 合流阀 5与第四换向阀 3连接。 当合流阀 5的大端面 54存在液压力 后可以使合流阀 5在大液阻通道 52位置， 合流阀 5的小端面 55存在液压力后 可以使合流阀 5在小液阻通道 53位置， 大、 小端面 54、 55都存在液压力后可 以使合流阀 5在大液阻通道 52位置， 若大、 小端面 54、 55都不存在液压力可 以使合流阀 5在断路通道 51位置。 在节流调速单元中的第四执行元件 Ί动作的 前提下， 合流阀 5受第一先导压力 P1和第二先导压力 P2的共同或单独作用实 现在断路通道 51、 大液阻通道 52以及小液阻通道 53之间换位， 从而连通负荷 传感单元与节流调速单元， 将节流调速单元的大部分流体分流后通过合流阀 5、 单向阀 6输入到负荷传感单元， 及时把第四执行元件 7的流体分流， 在保证第 四执行元件 7的压力与外负载一致、 第四执行元件 7可以正常工作前提下， 负 荷传感单元和节流调速单元内的油压不会急剧升高到极大值， 避免了因油压升 高导致恒功率控制阀 8控制变量柱塞泵 10排量变小， 最终造成***内所有执行 元件动作缓慢， 生产效率低， 动力源能量损失大的现象。
该装置在液压控制***的布置方面， 是通过定流量的节流调速单元与负荷 传感单元的合流控制工作方式来完成的。 在节流调速单元的第四执行元件 7 动 作的前提下， 负荷传感单元的第一换向阀 1受到第一先导压力 Pl、 第二换向阀 2受到第二先导压力 P2 (任意一个或两个同时）位置工作时，合流阀 5换向， 把 节流调速单元的流体大部分分流后通过合流阀 5、 单向阀 6， 输入到负荷传感单 元， 具体体现在以下三种形式：
( 1 )、 同时在第一换向阀 1上输入第一先导压力 Pl、 第四换向阀 3上输入 第四先导压力 P4, 使第一换向阀 1、 第四换向阀 3换向， 此时， 第一先导压力 P1还同时作用在合流阀 5的大端面 54上， 由于大端面 54的作用面积比较大， 因此在合流阀 5的大端面 54上的作用力也比较大， 该作用力可以克服复位弹簧 56力， 使合流阀 5的合流通道 50由断路通道 51换向到通道面积较大的大液阻 通道 52， 合流阀 5位于复位弹簧 56那端面的流体自由排回到油箱， 节流调速单 元的流体通过合流阀 5的大液阻通道 52、 单向阀 6， 输入到负荷传感单元。 同 时在大液阻通道 52形成的流体阻力与第一执行元件 12上的外负载相匹配， 从 而及时把第四执行元件 7上的流体分流出去。
( 2 )、 同时在第二换向阀 2上输入第二先导压力 P2、 第四换向阀 3上输入第 四先导压力 P4, 使第二换向阀 2、 第四换向阀换向 3， 第二先导压力 P2还同时 作用在合流阀 5的小端面 55上， 小端面 55作用面积比较小， 作用在小端面 55 上作用力也比较小， 但还是可以克服复位弹簧 56力， 使合流阀 5的合流通道 50 由断路通道 51换向到通道面积较小的小液阻通道 53， 合流阀 5位于复位弹簧 56那端面的流体自由排回到油箱， 节流调速单元的流体通过合流阀 5的小液阻 通道 53、 单向阀 6， 输入到负荷传感单元。 同时在小液阻通道 53形成的流体阻 力与第二执行元件 14上的外负载相匹配， 从而及时把第四执行元件 7的流体分 流出去。
( 3 )、 同时在第一换向阀 1上输入第一先导压力 Pl、 第二换向阀 2上输入第 二先导压力 P2、 第四换向阀 3上输入第四先导压力 P3， 使第一换向阀 1、 第二 换向阀 2、第四换向阀 3换向， 第一先导压力 P1与第二先导压力 P2还同时分别 作用在合流阀 5的大、 小端面 54、 55上， 作用在大、 小端面 54、 55上的作用 力可以克服复位弹簧 56力， 使合流阀 5的合流通道 50由断路通道 51换向到通 道面积较大的大液阻通道 52，合流阀 5位于复位弹簧 56那端面的流体自由排回 到油箱， 节流调速单元的流体通过合流阀 5的大液阻通道 52、 单向阀 6， 输入 到负荷传感单元。 由于第一执行元件 12上的外负载比第二执行元件 14上的外 负载大， 此时负荷传感单元中的压力与第一执行元件 12上的外负载相对应， 因 此只需合流阀 5的大液阻通道 52形成的流体阻力与第一执行元件 12上的外负 载相匹配， 就能及时把第四执行元件 7上的流体分流出去。
当负荷传感单元内相应的执行元件工作而节流调速单元上的所有执行元件 均不工作时， 节流调速单元可以直接零压泄荷， 不造成能量损失， 负荷传感单 元的执行元件还是可以避免因压力升高， 导致恒功率控制阀 8使变量柱塞泵 10 排量变小而***内造成所有执行元件动作缓慢， 生产效率低， 动力源的能量被 损失的现象。
当节流调速单元的执行元件工作而负荷传感单元上的所有执行元件均不工 作时， 节流调速单元的压力虽然能升高到很大值， 但此时动力源只给齿轮泵 15 提供能量， 不会造成生产效率低。
本发明通过设置连通负荷传感单元与节流调速单元的合流阀 5，使得流过合 流阀 5的合流通道 50所形成的流体阻尼与负荷传感单元内的执行元件的最高外 负载相匹配， 这样既不影响节流调速单元内的第四执行元件 7工作， 又可以及 时向负荷传感单元分流节流调速单元的流量， 避免单独应用负荷传感单元开始 工作时， 由于需克服大质量的外负载惯性而出现压力急剧升高， 负荷传感单元 内的执行元件动作慢， 效率低及损耗发动机 16能量的现象， 从而实现***高效 率， 低能耗的工作。
上述实施例只为说明本发明的技术构思及特点， 其目的在于让熟悉此项技 术的人士能够了解本发明的内容并加以实施， 并不能以此限制本发明的保护范 围， 凡根据本发明精神实质所作的等效变化或修饰， 都应涵盖在本发明的保护 范围内。

Claims

权 利 要 求 书
1.一种基于合流控制方式的液压装置， 包括负荷传感单元、 节流调速单元， 负荷传感单元具有第一换向阀 （1) 和第二换向阀 （2)， 节流调速单元具有第四 换向阀 （3)， 其特征是： 在与第四换向阀 （3) 并联设置的并联油路 （4) 上， 设有连通负荷传感单元与节流调速单元的合流阀 （5) 和单向阀 （6)， 所述合流 阀 （5) 具有控制并联油路通断而向负荷传感单元分流节流调速单元流体的合流 通道 （50)， 第四换向阀 （3) 连接有在动作时实现合流阀 （5) 换向的第四执行 元件 （7)， 在第一换向阀 （1) 受其第一先导压力 （Pl)、 第二换向阀 （2) 受其 第二先导压力 （P2)、 第四换向阀 （3) 受其第四先导压力 （P4) 作用而换向时， 所述第一先导压力（P1)和第二先导压力 （P2)还单独或同时作用在合流阀（5) 上使合流通道 （50) 位置变化而实现合流阀 （5) 换向。
2.根据权利要求 1所述的基于合流控制方式的液压装置， 其特征是： 所述 的负荷传感单元还包括恒功率控制阀 （8)、 可变排量机构 （9) 以及变量柱塞 泵（10)， 第一换向阀（1)分别连接有第一补偿阀（11)和第一执行元件（12)， 第二换向阀 （2) 分别连接有第二补偿阀 （13) 和第二执行元件 （14); 节流调 速单元还包括与变量柱塞泵 （10) 同轴的齿轮泵 （15)。
3.根据权利要求 1 所述的基于合流控制方式的液压装置， 其特征是： 所述 的合流通道 （50) 包括控制并联油路 （4) 通断的断路通道 （51)、 大液阻通道
(52) 以及小液阻通道 （53)， 合流阀 （5) —端具有： 同步接受第一先导压力 (P1) 控制的大端面 （54)、 同步接受第二先导压力 （P2) 控制的小端面 （55)， 合流阀（5)另一端设有复位弹簧（56)，第四换向阀（3)接受第四先导压力（P4) 控制并与合流阀 （5) 并联连接。
4.根据权利要求 3所述的基于合流控制方式的液压装置， 其特征是： 所述 的断路通道 （51) 的通道面积为零， 大液阻通道 （52) 和小液阻通道 （53) 的 通道面积不为零， 且大液阻通道 （52) 的通道面积大于小液阻通道 （53) 的通 道面积。
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