US20020007631A1 - Pressure compensating valve, unloading pressure control valve and hydraulically operated device - Google Patents
Pressure compensating valve, unloading pressure control valve and hydraulically operated device Download PDFInfo
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- US20020007631A1 US20020007631A1 US09/964,491 US96449101A US2002007631A1 US 20020007631 A1 US20020007631 A1 US 20020007631A1 US 96449101 A US96449101 A US 96449101A US 2002007631 A1 US2002007631 A1 US 2002007631A1
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- pressure
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- 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
- E02F9/2264—Arrangements or adaptations of elements for hydraulic drives
- E02F9/2267—Valves or distributors
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- 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
- E02F9/2239—Control of flow rate; Load sensing arrangements using two or more pumps with cross-assistance
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- 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
- E02F9/2264—Arrangements or adaptations of elements for hydraulic drives
- E02F9/2271—Actuators and supports therefor and protection therefor
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- 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
- E02F9/2278—Hydraulic circuits
- E02F9/2285—Pilot-operated systems
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- 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
- E02F9/2278—Hydraulic circuits
- E02F9/2292—Systems with two or more pumps
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- 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
- E02F9/2278—Hydraulic circuits
- E02F9/2296—Systems with a variable displacement pump
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- 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
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- 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
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- 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
- F15B2211/2053—Type of pump
- F15B2211/20546—Type of pump variable capacity
- F15B2211/20553—Type of pump variable capacity with pilot circuit, e.g. for controlling a swash plate
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- 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
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- 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
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- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
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- F15B2211/305—Directional control characterised by the type of valves
- F15B2211/30525—Directional control valves, e.g. 4/3-directional control valve
- F15B2211/3053—In combination with a pressure compensating valve
- F15B2211/3054—In combination with a pressure compensating valve the pressure compensating valve is arranged between directional control valve and output member
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- 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
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- F15B2211/305—Directional control characterised by the type of valves
- F15B2211/3056—Assemblies of multiple valves
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- 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
- F15B2211/31—Directional control characterised by the positions of the valve element
- F15B2211/3105—Neutral or centre positions
- F15B2211/3111—Neutral or centre positions the pump port being closed in the centre position, e.g. so-called closed centre
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- 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
- F15B2211/40—Flow control
- F15B2211/405—Flow control characterised by the type of flow control means or valve
- F15B2211/40507—Flow control characterised by the type of flow control means or valve with constant throttles or orifices
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- 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
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- 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
- F15B2211/40—Flow control
- F15B2211/415—Flow control characterised by the connections of the flow control means in the circuit
- F15B2211/41527—Flow control characterised by the connections of the flow control means in the circuit being connected to an output member and a directional control valve
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- 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
- F15B2211/40—Flow control
- F15B2211/42—Flow control characterised by the type of actuation
- F15B2211/428—Flow control characterised by the type of actuation actuated by fluid pressure
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- 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
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- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
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- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
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- F15B2211/473—Flow control in one direction only without restriction in the reverse direction
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- 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
- F15B2211/60—Circuit components or control therefor
- F15B2211/605—Load sensing circuits
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- 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
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- F15B2211/6051—Load sensing circuits having valve means between output member and the load sensing circuit
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- 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
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- 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
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- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
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- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
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- 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
Definitions
- the present invention relates to a pressure compensating valve, an unloading pressure control valve, and a hydraulically operated device.
- FIG. 25 depicts a hydraulically operated device described in Japanese Unexamined-Patent Application 1-247805.
- a variable delivery pump A is connected to a low pressure hydraulic cylinder D via a pressure compensating valve B and a directional control valve (operating valve) C.
- the pump A is also connected to a high pressure cylinder D′ via a pressure compensating valve B′ and a directional control valve C′.
- An actuator E for changing the displacement volume and a flow regulating valve F for controlling the actuator E are attached to the hydraulic pump A.
- the flow regulating valve F controls the actuator E so that the discharge pressure P P of the pump A is always greater than the maximum load pressure P LS
- the cylinders D and D′ are jointly operated by the simultaneous operation of directional control valves C and C′ in the hydraulically operated device.
- the pressure compensating valve B controls the amount of oil supplied to the cylinder D so that the difference between the input pressure and the output pressure of the directional control valve C is constant
- the pressure compensating valve B′ similarly controls the amount of oil supplied to the cylinder D′ so that the difference between the input pressure and the output pressure of the directional control valve C′ is constant.
- the hydraulically operated device equipped with the pressure compensating valves B and B′ can prevent the disadvantage of pressured oil accumulating and being supplied to the cylinder with the lighter load among the operating valve cylinders D and D′.
- the aforementioned hydraulically operated device is provided with a pressure difference sensing device H for sensing the pressure difference P P -P LS between the pressure P P of the pressured oil discharged from the hydraulic pump A and the maximum load pressure P LS , a control force set device I for setting the control force fc based on the pressure difference P P -P LS and the relationship depicted in FIG. 26, and an electromagnetic valve J that is operated by means of the output signals from the control force setting device I.
- control force fc is given by the following equation.
- the electromagnetic valve J allows pressured oil corresponding to the control force fc to act on the pressure receiving components of the pressure compensating valves B and B′ when the pressure difference P P -P LS is at or below the specific pressure difference Pm shown in FIG. 26.
- the pressure compensation characteristics of the pressure compensating valves B and B′ are preferably modified in some cases to improve the operating characteristics, depending on the operating configuration of the aforementioned operating device.
- a technique that is capable of changing the throttle levels for each pressure compensating valve and that is capable of suitably changing the pressure difference before and after the directional control valves C and C′ has been disclosed in the aforementioned patent publication. That is, in this technique, electromagnetic valves J as described above are provided for the pressure compensating valves B and B′, and the control force fc for the pressure compensating valves B and B′ are individually adjusted by these electromagnetic valves J. Accordingly, the throttle levels of the pressure compensating valves B and B′ are individually changed; that is, the pressure differences before and after the direction control valves C and C′ are different from each other.
- a state in which the required flow rate is distributed completely irrespective of load is also referred to in particular as a fully compensated state.
- a first object of the present invention is to provide a pressure compensating valve that allows the pressure compensation characteristics to be arbitrarily modified, that has good response, and that is highly reliable.
- FIG. 27 depicts a hydraulically operated device described in Japanese Unexamined Patent Application 4-250226.
- a flow regulating valve (operating valve) B is operated, by means of the pilot pressure produced by the operating device A, to an extent corresponding to the extent to which the operating device A has been operated, and the discharged pressured oil from a hydraulic pump D is consequently supplied to a hydraulic cylinder (hydraulic actuator) C.
- a pressure compensating valve E for keeping the pressure difference before and after the flow regulating valve B at a constant level is located between the hydraulic pump D and the flow regulating valve (operating valve) B.
- An operating device A′, flow regulating valve (operating valve) B′, hydraulic motor (hydraulic actuator) C′, and pressure compensating valve E′ each correspond to the operating device A, flow regulating valve (operating valve) B, hydraulic cylinder C, and pressure compensating valve E.
- An unloading pressure control valve F is connected in parallel to the hydraulic pump D.
- the higher pressure between the load pressure acting on the hydraulic cylinder C and the load pressure acting on the hydraulic motor C′ is sensed as the maximum load pressure by a shuttle valve G, and this maximum load pressure is allowed to act on the unloading pressure control valve F.
- the unloading pressure control valve F is provided to return the discharged oil from the hydraulic pump D to the tank.
- the amount of the aforementioned discharged oil returned by the unloading pressure control valve F is set by the difference between the maximum load pressure and the discharge pressure of the hydraulic pump D, and by control signals output from a control unit J.
- a computer H connected to the control unit J computes the difference ⁇ P LS between the discharge pressure of the hydraulic pump D and the load pressure of the hydraulic cylinder C or hydraulic motor C′ based on the functional relation shown in FIG. 28 and the output of sensors a 1 , a 2 and a 1 ′, a 2 ′ for sensing the control input of the operating devices A and A′.
- the function shown in FIG. 28 defines a relation in which the pressure difference ⁇ P LS increases proportionally until the control input St of the operating device A reaches a set value, and the pressure difference ⁇ P LS stays at a value ⁇ P LS 1 when the control input St is at or beyond the set value.
- the pressure difference ⁇ P LS is computed by the computer H, so a control signal corresponding to a pressure difference ⁇ P LS2 is output from the control unit J, and the unloading start pressure of the unloading pressure control valve F is set to pressure difference ⁇ P LS2 .
- the amount of pressured oil supplied through the pressure compensating valve E′ and flow regulating valve B′ to the hydraulic motor C′ is the amount defined by the pressure difference ⁇ P LS2 .
- FIG. 29 shows the relation between the amount of oil Q supplied to the hydraulic motor C′ and the pressure difference AP before and after the flow regulating valve B′ when the control input St is 20%.
- the pressure compensating valve E′ supplies pressured oil in a constant oil amount q 2 to the hydraulic motor C′ so that the pressure difference AP of the flow regulating valve B′ is kept at a constant pressure difference ⁇ Pc+ ⁇ P LS ( ⁇ P LS is the pressure loss of the pressure compensating valve E′).
- ⁇ P LS is the pressure loss of the pressure compensating valve E′.
- the pressured oil is supplied to the hydraulic motor C′ in the oil amount q 1 defined by the unloading start pressure ⁇ P LS2 of the unloading pressure control valve F.
- the unloading start pressure of the unloading pressure control valve F is variable.
- the unloading start pressure is set through the computer H and the control unit J. It is accordingly always set after the output from the sensors a 1 , a 2 and a 1 ′, a 2 ′ of the operating devices A and A′, and a resulting problem is the poor response in terms of the hydraulic cylinder C or the hydraulic motor C′. More specifically, when the fluctuations in the load pressure of the hydraulic cylinder C or hydraulic motor C′ are estimated, the unloading start pressure is hopefully pre-modified rapidly irrespective of the control input of the operating devices A and A′. For the reasons described above, however, the unloading start pressure is difficult to modify in advance.
- a second object of the present invention is to provide an unloading pressure control valve allowing the unloading start pressure to be preset so as to improve the response in terms of a hydraulic actuator.
- a pump discharge pressure control means for controlling the displacement volume of a hydraulic pump for controlling the displacement volume of a hydraulic pump (discharge volume per revolution) is provided in a hydraulically operated device in which the pressured oil discharged from a variable delivery pump is supplied to a hydraulic actuator such as a hydraulic cylinder by the operation of an operating valve.
- This pump discharge pressure control means is designed so as to control the displacement volume of a hydraulic pump based on the discharge pressure of a hydraulic pump and the load pressure acting on a hydraulic actuator, so that the aforementioned discharge pressure is greater by a specific pressure than the aforementioned load pressure.
- the displacement volume of the hydraulic pump immediately increases to a magnitude corresponding to the load pressure.
- the actuator is also connected via a pressure compensating valve. A flow rate corresponding to the control input of the operating valve can thus be supplied, irrespective of the magnitude of the load pressure, to the actuator.
- the aforementioned hydraulic actuator is a hydraulic motor or cylinder driving an operating unit in construction machinery (such as the revolving superstructure, boom, arm, or bucket in the case of a hydraulic shovel, for example)
- the rapid start up of the aforementioned hydraulic actuator results in lower operating performance, depending on the operating configuration.
- Hydraulically operated devices such as the following have been proposed in patent publications.
- a bleed valve is connected to the discharge channel of the aforementioned hydraulic pump, and part of the pressured oil discharged by the hydraulic pump is bled through the bleed valve to the tank.
- the bleed valve used in the hydraulically operated device of the aforementioned patent publication bleeds off part of the pressured oil discharged from the hydraulic pump to the tank.
- a large amount of the pressured oil that is supposed to be supplied to the hydraulic actuator ends up being returned to the tank when bled off. This results in significant energy loss.
- a third object of the present invention is to provide a hydraulically operated device that allows energy loss to be minimized to control rapid start up of hydraulic actuators, and that also allows machinery to be made more compact and high-precision control to be achieved.
- Another object of the present invention is to simultaneously achieve the first and second objects.
- Still another object of the present invention is to simultaneously achieve the first and third objects.
- Yet another object of the present invention is to simultaneously achieve the second and third objects.
- Another object of the present invention is to simultaneously achieve the first, second, and third objects.
- the first of the present inventions is a pressure compensating valve through which passes pressured oil that is fed from a hydraulic pump 1 to a hydraulic actuator 5 , characterized by comprising a main valve 20 that is operated in such a way as to increase the area of the opening between an inlet port 24 and an outlet port 25 by means of pressure acting on a first pressure receiving component 21 , that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component 22 and pressure acting on a third pressure receiving component 23 , and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port 24 to act on the first pressure receiving component 21 and the pressure Pb of the load 5 driven by the pressured oil flowing from the outlet port 25 to act on the second pressure receiving component 22 ; and control pressure producing means 7 B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port 24 to act on the third pressure receiving component 23 .
- the first invention allows the desired pressure compensation characteristics to be obtained by changing the control pressure Pe because the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe.
- the second invention is an unloading pressure control valve for introducing discharged pressured oil from a hydraulic pump 1 to a tank according to the pressure difference between the discharge pressure P P of the hydraulic pump 1 and the load pressure P LS of a hydraulic actuator 5 , characterized by comprising a main valve 20 that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure P P of the hydraulic pump 1 acting on a first pressure receiving component 123 , to operate in the cut-off direction upon load pressure P LS to a second pressure receiving component 124 , and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component 125 ; and control pressure producing means 101 for producing the control pressure Pg.
- the second invention allows the unloading start pressure to be set by means of the control pressure Pg acting on the third pressure receiving component 125 .
- the control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump 1 can be increased in advance to improve the response in terms of the hydraulic actuator 5 .
- the third invention is a hydraulically operated device comprising a plurality of hydraulic actuators 5 to which pressured oil discharged from a variable delivery pump 1 is supplied via pressure compensating valves 7 and directional control valves 4 ; means for outputting pressure P LS to a load pressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators; and pump discharge pressure control means for controlling the discharge pressure of the hydraulic pump 1 based on the pressure P LS ; wherein the hydraulically operated device is characterized in that a variable bleed valve 11 is located in the load pressure sensing passage 9 .
- the third invention allows the amount discharged from the hydraulic pump 1 to be controlled by bleeding off the pressured oil in the load pressure sensing passage 9 .
- the amount flowing in the load pressure sensing channel 9 is generally quite low.
- the pump pressure is controlled according to the pressure of the load pressure sensing passage 9 , whereas the pressure of the load pressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact.
- the amount discharged from the hydraulic pump 1 can be controlled with greater precision.
- the fourth of the inventions is a hydraulically operated device comprising a pressure compensating valve through which passes pressured oil that is fed from a hydraulic pump 1 to a hydraulic actuator 5 ; and an unloading pressure control valve for introducing discharged pressured oil from the hydraulic pump 1 to a tank according to the pressure difference between the discharge pressure P P of the hydraulic pump 1 and the load pressure P LS of the hydraulic actuator 5 ; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve 7 itself comprising a pressure compensated main valve 20 that is operated in such a way as to increase the area of the opening between an inlet port 24 and an outlet port 25 by means of pressure acting on a first pressure receiving component 21 for a pressure compensating valve, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component 22 for a pressure compensating valve and pressure acting on a third pressure receiving component 23 for a pressure compensating valve, and that is designed to allow the
- the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.
- control pressure Pe resulting from a reduction in the pressure of the inlet port 24 is allowed to act on the third pressure receiving component 23 for a pressure compensating valve in the main valve 20 for a pressure compensating valve, the control pressure Pe also fluctuates according to the fluctuations in the pressure of the inlet port 24 .
- the pressure compensation characteristics are thus unaffected by the pressure fluctuations in the inlet port 24 of the main valve 20 .
- the unloading start pressure can be set by means of the control pressure Pg acting on the third pressure receiving component 125 for an unloading pressure control valve.
- the control pressure Pg is produced by the control pressure producing means.
- the unloading start pressure can thus be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump 1 can be increased in advance to improve the response in terms of the hydraulic actuator 5 .
- the fifth of the inventions is a hydraulically operated device comprising a plurality of hydraulic actuators 5 to which pressured oil discharged from a variable delivery pump 1 is supplied via pressure compensating valves and directional control valves 4 ; means 8 for outputting pressure P LS to a load pressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators 5 ; and pump discharge pressure control means 12 for controlling the discharge pressure of the hydraulic pump 1 based on the pressure P LS ; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve 7 itself comprising a main valve 20 that is operated in such a way as to increase the area of the opening between an inlet port 24 and an outlet port 25 by means of pressure acting on a first pressure receiving component 21 , that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component 22 and pressure acting on a third pressure receiving component 23 , and that is designed to allow the pressure Pa
- the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.
- control pressure Pe resulting from a reduction in the pressure of the inlet port 24 is allowed to act on the third pressure receiving component 23 of the main valve 20 , the control pressure Pe also fluctuates according to the fluctuations in the pressure of the inlet port 24 .
- the pressure compensation characteristics are thus unaffected by the pressure fluctuations in the inlet port 24 of the main valve 20 .
- the amount discharged from the hydraulic pump 1 can be controlled by bleeding off the pressured oil in the load pressure sensing passage 9 .
- the amount flowing in the load pressure sensing channel 9 is generally quite low.
- the pump pressure is controlled according to the pressure of the load pressure sensing passage 9 , whereas the pressure of the load pressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact.
- the amount discharged from the hydraulic pump 1 can be controlled with greater precision.
- the sixth of the present inventions is a hydraulically operated device comprising a plurality of hydraulic actuators 5 to which pressured oil discharged from a variable delivery pump 1 is supplied via pressure compensating valves 7 and directional control valves 4 ; means 8 for outputting pressure P LS to a load pressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators 5 ; pump discharge pressure control means 12 for controlling the discharge pressure of the hydraulic pump 1 based on the pressure P LS ; and an unloading pressure control valve for introducing discharged pressured oil from the hydraulic pump 1 to a tank according to the pressure difference between the discharge pressure P P of the variable delivery pump 1 and the load pressure P LS of the hydraulic actuators 5 ; wherein the hydraulically operated device is characterized by comprising an unloading pressure control valve 10 itself comprising a main valve 100 that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure P P of the hydraulic pump 1 acting on a first pressure receiving
- the unloading start pressure can be set by means of the control pressure Pg acting on the third pressure receiving component 25 .
- the control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump 1 can be increased in advance to improve the response in terms of the hydraulic actuator 5 .
- the amount discharged from the hydraulic pump 1 can be controlled by bleeding off the pressured oil in the load pressure sensing passage 9 .
- the amount flowing in the load pressure sensing channel 9 is generally quite low.
- the pump pressure is controlled according to the pressure of the load pressure sensing passage 9 , whereas the pressure of the load pressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact.
- the amount discharged from the hydraulic pump 1 can be controlled with greater precision.
- the seventh of the present inventions is a hydraulically operated device comprising a plurality of hydraulic actuators 5 to which pressured oil discharged from a variable delivery pump 1 is supplied via pressure compensating valves and directional control valves 4 ; means 8 for outputting pressure P LS to a load pressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators 5 ; pump discharge pressure control means 12 for controlling the discharge pressure of the variable delivery pump 1 based on the pressure P LS ; and an unloading pressure control valve for introducing discharged pressured oil from the variable delivery pump 1 to a tank according to the pressure difference between the discharge pressure P P of the variable delivery pump 1 and the load pressure P LS of the hydraulic actuators 5 ; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve 7 itself comprising a pressure compensated main valve 20 for a pressure compensating valve, that is operated in such a way as to increase the area of the opening between an inlet port
- the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.
- control pressure Pe resulting from a reduction in the pressure of the inlet port 24 is allowed to act on the third pressure receiving component 23 for a pressure compensating valve in the main valve 20 for a pressure compensating valve, the control pressure Pe also fluctuates according to fluctuations in the pressure of the inlet port 24 .
- the pressure compensation characteristics are thus unaffected by the fluctuations in the inlet port 24 of the main valve 20 for a pressure compensating valve.
- the unloading start pressure can be set by means of the control pressure Pg acting on the third pressure receiving component 125 for an unloading pressure control valve.
- the control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump 1 can be increased in advance to improve the response in terms of the hydraulic actuators 5 .
- the amount discharged from the hydraulic pump 1 can be controlled by bleeding off the pressured oil in the load pressure sensing passage 9 .
- the amount flowing in the load pressure sensing channel 9 is generally quite low.
- the pump pressure is controlled according to the pressure of the load pressure sensing passage 9 , whereas the pressure of the load pressure sensing passage 9 is the pressure corresponding to the load pressure of the actuators and thus reacts exactly to the fluctuations in the load pressure of the actuators. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact.
- the amount discharged from the hydraulic pump 1 can be controlled with greater precision.
- FIG. 1 is a circuit diagram of the oil pressure in a hydraulically operated device relating to the present invention
- FIG. 2 is a circuit diagram of oil pressure, depicting the structure of a pressure compensating valve relating to the present invention
- FIG. 3 is a longitudinal cross section depicting the attachment of a pressure compensating valve and an operating valve relating to the present invention
- FIG. 4 is a longitudinal cross section, depicting the structure of a pressure compensating valve relating to the present invention
- FIG. 5 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 6 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 7 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 8 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 9 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 10 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 11 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 12 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 13 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 14 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 15 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention.
- FIG. 16 is a circuit diagram of oil pressure, depicting the structure of an unloading pressure control valve relating to the present invention.
- FIG. 17 is a cross section depicting a specific structure for an unloading pressure control valve relating to the present invention.
- FIG. 18 is a cross section depicting another embodiment of an unloading pressure control valve relating to the present invention.
- FIG. 19 is a circuit diagram of oil pressure in another hydraulic system involving the application of an unloading pressure control valve relating to the present invention
- FIG. 20 is a circuit diagram of oil pressure, depicting an enlargement of the structure of a variable bleed valve used in the hydraulically operated device of FIG. 1;
- FIG. 21 depicts an embodiment with a variable bleed valve attached to the hydraulically operated device in FIG. 1;
- FIG. 22 is a cross section of line A-A in FIG. 21;
- FIG. 23 is a graph depicting an example of the relation between input and output when a mode set memory means has been set and stored;
- FIG. 24 is a graph depicting another example of the relation between input and output when a mode set memory means has been set and stored;
- FIG. 25 is a circuit diagram of oil pressure, depicting the structure of a conventional hydraulic device equipped with a pressure compensating valve
- FIG. 26 is a graph depicting the relation between pressure difference and control force
- FIG. 27 is a circuit diagram of oil pressure in a conventional hydraulically operated device in which an unloading pressure control valve is used;
- FIG. 28 is a graph depicting the relation between the pressure difference and the control input of an operating device.
- FIG. 29 is a graph depicting the relation between the pressure difference before and after a flow regulating valve and the amount of oil Q supplied to a hydraulic motor.
- FIG. 1 depicts an embodiment of a hydraulically operated device relating to the present invention.
- the hydraulically operated device can be used for a hydraulic shovel, for example.
- the hydraulically operated device comprises a variable delivery pump 1 , auxiliary hydraulic pump 2 , a plurality of closed center operating valves (directional control valves) 4 to which the oil discharged from the hydraulic pump 1 is supplied through an oil passage 3 , and a plurality of hydraulic cylinders 5 corresponding to each operating valve 4 .
- the head oil chambers of the hydraulic cylinders 5 are connected by means of oil passages 6 a and pressure compensating valves 7 to the operating valves 4 , and the bottom oil chambers are connected by means of a pressure compensating valve not shown in the figure in an oil passage 6 b to the operating valves 4 .
- a pressure compensating valve is in fact interposed in the oil passage 6 b , but thus pressure compensating valve has been left out in FIG. 1 to avoid complicating the drawing.
- the load pressure P 1 of the hydraulic cylinders 5 connected thereto act on each of the oil passages 6 a .
- the maximum load pressure among the load pressures P 1 acting on these oil passages 6 a are sensed as the maximum load pressure P LS by a shuttle valve 8 , and the sensed maximum load pressure P LS is allowed by means of an oil passage (load pressure sensing passage) 9 to act on the hydraulic pump 1 , pressure compensating valves 7 , unloading pressure control valve 10 , and variable bleed valve 11 .
- a fixed throttle 13 is interposed between the tank and the oil passage 9 into which the pressured oil with the maximum load pressure P LS is introduced.
- a pump discharge pressure control means 12 is attached to the hydraulic pump 1 .
- the pump discharge pressure control means 12 introduces the discharge pressure P P of the hydraulic pump 1 and the maximum load pressure P LS ′ and controls the displacement volume of the hydraulic pump 1 so that the discharge pressure P P is always slightly higher than the maximum load pressure P LS .
- the structure of the pressure compensating valve 7 relating to the present invention is described below with reference to FIG. 2.
- the pressure compensating valve 7 is composed of a compensator 7 A, a control pressure producing component 7 B, and a pilot pressure supply component 7 C.
- the compensator 7 A has a main valve 20 .
- the main valve 20 comprises a first pressure receiving component 21 , a second pressure receiving component 22 , and a third pressure receiving component 23 .
- the pressure Pa acting on the first pressure receiving component 21 acts in such a way as to increase the area of the opening between the inlet port 24 and outlet port 25 .
- the pressure Pb acting on the second pressure receiving component 22 and the pressure Pc acting on the third pressure receiving component 23 act in such a way as to reduce the area of the opening along with the elastic force of a spring 26 .
- the inlet port 24 is connected to the outlet port of the operating valve 4 depicted in FIG. 1.
- the pressure Pa of the inlet port 24 acts on the first pressure receiving component 21 via an oil passage 27 .
- the outlet port 25 is connected to the oil passage 6 a through a load check valve 28 .
- a shuttle valve 29 senses the load pressure P 1 acting on the oil passage 6 a and the greater pressure Pb among the maximum load pressure P LS , and allows the pressure Pb to act on the second pressure receiving component 22 of the main valve 20 .
- the control pressure producing component 7 B has a variable throttle valve 30 .
- The-variable throttle valve 30 is operated in such a way as to reduce the area of the opening between an inlet port 32 and an outlet port 33 by means of the elastic force of a spring 31 . It is also operated in such a way as to increase the area of the opening by means of the elastic force of a spring 35 and the pilot pressure P 2 acting on a pressure receiving component 34 .
- the spring 35 is used only in the initial fine tuning of the variable throttle valve 30 , and is not indispensable.
- the tank port pressure is allowed to act constantly on the spring 31 to ensure that the variable throttle valve 30 is operated more rapidly.
- the inlet port 32 of the variable throttle valve 30 is connected to the inlet port 24 of the main valve 20 by way of an oil passage 37 equipped with a throttle 36 .
- the pressure Pe of the inlet port 32 of the variable throttle valve 30 acts on the third pressure receiving component 23 of the main valve 20 .
- the outlet port 33 of the variable throttle valve 30 is connected to the oil passage 6 a by way of an oil passage 40 equipped with a check valve 39 .
- the pilot pressure Pd is given as the output pressure of an electromagnetic proportional pressure control valve 50 located in the pilot pressure supply component 7 C.
- the electromagnetic proportional pressure control valve 50 introduces the pressured oil discharged from the pilot hydraulic pump 2 depicted in FIG. 1 to an inlet port 52 .
- the pressure Pa of this pressured oil is lowered to the pilot pressure Pd by means of the electricity applied to a solenoid 53 .
- the pilot pressure Pd displays a magnitude proportional to the amount of electricity to the solenoid 53 .
- pressure compensating valves 7 are interposed not only in oil passages 6 a but also in oil passages 6 b in the hydraulically operated device in FIG. 1.
- FIG. 3 depicts an example of the structure of an operating valve 4 by which pressured oil is selectively supplied to the two pressure compensating valves 7 described above.
- the operating valve 4 has a structure in which a body 60 is provided with a spool 61 , pairs of left and right outlet ports 62 , pairs of left and right pump ports 63 , pairs of left and right actuator ports 64 , and pairs of left and right of tank ports 65 .
- the spool 61 blocks all of the ports 62 through 65 in the center valve state depicted in the figure.
- the outlet ports 62 on one side communicate with the pump ports 63
- the actuator ports 64 on the other side communicate with the tank ports 65 .
- the outlet ports 65 on the other side communicate with the pump ports 63
- the actuator ports 64 on the first side communicate with the tank ports 65 .
- the main valve 20 located in the compensator 7 A of the pressure compensating valve 7 has a valve component 66 A interposed between the actuator port 64 and the outlet port 62 of the operating valve 4 , and a pressing component 67 connected to the valve component 66 A.
- FIG. 4 depicts an enlargement of the pressure compensating valve 7 .
- the valve component 66 comprises a hollow component 68 open at the left end, a hole 69 open in the outer peripheral surface through the hollow component 68 , and a seat surface 71 that presses into contact with a seat 70 formed in the body 60 .
- Pressure-receiving surfaces 66 a and 66 b of the valve component 66 form the first pressure receiving component 21 of the main valve 20 depicted in FIG. 2, and the hole 69 of the valve component 66 forms the outlet port 25 of the main valve 20 .
- the entire valve component 66 functions as the load check valve 39 depicted in FIG. 2.
- the pressing component 67 is positioned on an extension of the central axis of the valve component 66 , and comprises a piston 73 that slides to the left and right in a sleeve 72 fixed to the body 60 , a sliding element 74 that slides to the left and right in the piston 73 , and a spring 26 (see FIG. 2) interposed between the sleeve 72 and the sliding element 74 .
- An annular space 75 into which the pressured oil with the maximum 1 load pressure P LS (see FIG. 2) is introduced is formed between the body 60 and the sleeve 72 .
- the pressured oil with the maximum load pressure P LS introduced into this annular space 75 flows into a stepped hole 80 in the sliding element 74 through a fine hole 76 located in the sleeve 72 , an annular groove 77 , a hole 78 located in the piston 73 , and an inlet port 79 located in the sliding element 74 , and acts on the right side of a bore 81 located in the stepped hole 80 .
- the stepped hole 80 communicates with a pressure chamber 83 through the outlet port 84 and a convex groove 85 located in the outer peripheral surface thereof.
- the stepped hole 80 and bore 81 have the function of sensing the higher oil pressure between the oil pressure P LS and P 1 , and of guiding it into the pressure chamber 83 .
- the shuttle valve 29 depicted in FIG. 2 is composed of the stepped hole 80 and the bore 81 .
- the pressure of the pressured oil introduced into the pressure chamber 83 is pressure Pb depicted in FIG. 2.
- the pressure chamber 83 is a space enclosed by the inner surface of the sleeve 72 , the right end surface of the piston 73 , and the outer peripheral surface of the sliding element 74 , where the right end surface of the piston 73 functions as the second pressure receiving component 22 depicted in FIG. 2.
- control pressure producing component 7 B is described below.
- the control pressure producing component 7 B is located to the side of the compensator 7 A, and is equipped with the variable throttle valve 30 depicted in FIG. 2.
- a spool 88 for changing the flow resistance (throttle level) between the inlet port 32 and outlet port 33 depicted in FIG. 2 is located in the vertical direction in the body 87 of the variable throttle valve 30 .
- the spool 88 is such that downwardly directed force (the direction in which the flow resistance increases) is provided by the spring 31 , and upwardly directed force (the direction in which the flow resistance decreases) is given by the spring 35 in a pressure chamber 90 formed between the spool and an adjusting screw 89 .
- the inner surface of a concave component 91 located in the left surface of the body 87 forms a pressure chamber 92 along with the right end surface of the sliding element 74 and the right end surface of the sleeve 72 of the compensator 7 A.
- the right end surface of the sliding element 74 facing the pressure chamber 92 forms the second pressure receiving component 23 of the main valve 20 depicted in FIG. 2.
- the inlet port 32 of the variable throttle valve 30 communicates through the oil passage 37 equipped with the throttle 36 to the outlet port 62 of the operating valve 4 , that is, to the inlet port 24 of the main valve 20 depicted in FIG. 2, and also communicates-through an oil passage 38 to the pressure chamber 92 .
- the outlet port 33 communicates through the oil passage 40 equipped with the check valve 39 (see FIG. 2) to the actuator port 64 of the operating valve 4 .
- the pilot pressure producing component 7 C is located in the top of the body 87 of the control pressure producing component 7 B.
- the electromagnetic proportional pressure control valve 50 forming the pilot pressure producing component 7 C comprises a spool 94 arranged in the vertical direction in the body 93 , and a solenoid 53 that presses the spool 94 down against the spring 54 .
- the outlet port 55 communicates through an oil passage 95 to the pressure chamber 90 of the variable throttle valve 30 .
- the spool 94 also is positioned on the axis of the spool 88 of the control pressure producing component 7 B.
- the spring 26 also presses the valve component 66 to the left by means of the sliding element 74 .
- Pa ⁇ A 0 Pe ⁇ A 1 +Pb ( A 0 ⁇ A 1 )+ F 0 (1)
- a 0 sum of the area of surfaces 66 a and 66 b of valve component 66
- a 1 area of right end surface of sliding element 74
- a 0 ⁇ A 1 area of right end surface of piston 73
- the pressure Pe in Eq. (1) is the control pressure that changes the pressure compensation characteristics of the pressure compensating valve 7 .
- the control pressure Pe results in pressure Pa when the variable throttle valve 30 of control pressure producing component 7 B depicted in FIG. 2 is closed.
- the relation in Eq. (2) below is obtained by substituting Pa into Pe in Eq. (1).
- Pa ⁇ Pb F 0 /( A 0 ⁇ A 1) (2)
- variable throttle valve 30 of the control pressure producing component 7 B When the variable throttle valve 30 of the control pressure producing component 7 B is not closed, the pressured oil passing through the fixed throttle 36 flows through the variable throttle valve 30 and check valve 39 to the cylinder 5 end.
- the control pressure Pe obtained by dividing the pressure difference between the pressures Pa and P 1 by the throttling ratio between the throttle 36 and variable throttle valve 30 , in other words, the control pressure Pe resulting from the reduction of the pressure Pa, acts on the third pressure receiving component 23 of the main valve 20 in the compensator 7 A.
- the control pressure Pe drops as the amount of electricity to the solenoid 53 of the electromagnetic proportional pressure control valve 50 increases. Accordingly, when an operating unit in construction machinery, for example (such as the boom, arm, or bucket in hydraulic shovels), is driven by cylinders 5 , pressure compensation characteristics suitable for the operating configuration of such an operating unit can be set by controlling the amount of electricity to the solenoid 53 .
- the pressure compensation characteristics of pressure compensating valves 7 for a plurality of cylinders, as shown in FIG. 1, can also be altered, of course.
- the pressure compensation characteristics of the pressure compensating valves 7 for the series of cylinders 5 depicted in FIG. 3 can each be varied so as to alter the operating speeds during extension and retraction of the cylinders 5 .
- the pilot pressure Pd no longer acts on the variable throttle valve 39 of the control pressure producing component 7 B in the event of wire breakage in the solenoid 53 of the electromagnetic proportional pressure control valve 50 or in the event of malfunctions of the pilot pump 2 depicted in FIG. 1, for example.
- the variable throttle 39 is no longer capable of throttling operations.
- FIG. 5 depicts a second example of the structure of a pressure compensating valve 7 .
- This pressure compensating valve 7 differs from the pressure compensating valve 7 in FIG. 4 in that the spring 54 of the electromagnetic proportional pressure control valve 50 is brought into contact with the top end of the spool 88 of the variable throttle valve 30 of the control pressure producing component 7 B.
- FIG. 6 depicts a third example of the structure of the pressure compensating valve 7 .
- This pressure compensating valve 7 is such that the variable throttle valve 30 of the control pressure producing component 7 B and the electromagnetic proportional pressure control valve 50 have a shared body 218 , with the solenoid 53 of the electromagnetic proportional pressure control valve 50 located on the exterior of the body 218 . This allows the structure to be made more compact and the number of parts to be reduced.
- control pressure producing component 7 B in this pressure compensating valve 7 forms a flange 88 a having a tapered peripheral surface on the spool 88 of the variable throttle valve 30 , and the flange 88 a is interposed between the inlet port 32 and outlet port 33 of the variable throttle valve 30 .
- the spool 88 functions as a check valve to prevent the pressured oil with the pressure P 1 from flowing toward the inlet port 32 .
- the body 218 of this pressure compensating valve 7 thus does not require the check valve 39 depicted in FIGS. 4 and 5, making the body 218 easier to fabricate.
- FIG. 7 depicts a fourth example of the structure of the pressure compensating valve 7 .
- This pressure compensating valve 7 has a structure in which a joint 102 is attached to an attachment block 219 secured to the top surface of the body 87 of the control pressure producing component 7 B, and the pressure chamber 90 of the variable throttle valve 30 in the control pressure producing component 7 B communicates through the joint 102 and piping 95 to the outlet port 55 of the electromagnetic proportional pressure control valve 50 .
- this pressure compensating valve 7 the pilot pressure Pd output from the electromagnetic proportional pressure control valve 50 or the pilot pressure output from a manual pilot valve can be allowed to act on the variable throttle valve 30 of the control pressure producing component 7 B by way of the joint 102 .
- This pressure compensating valve 7 is thus suitable for use in cases where the electromagnetic proportional pressure control valve 50 or pilot valve must be located at a distance from the control pressure producing component 7 B because of restricted space or the like.
- variable throttle valve 30 of the control pressure producing component 7 B in this pressure compensating valve 7 has a structure similar to that of the variable throttle valve 30 of the pressure compensating valve 7 depicted in FIG. 4.
- FIG. 8 depicts a fifth example of the structure of the pressure compensating valve 7 .
- This pressure compensating valve 7 has a structure in which a joint 104 is attached to the exterior of the body 103 of the control pressure producing component 7 B, and the pressure chamber 90 located in the variable throttle valve 30 of the control pressure producing component 7 B communicates through the joint 104 and piping 95 to the electromagnetic proportional pressure control valve 50 or a manual pilot valve not shown in the figure.
- the electromagnetic proportional pressure control valve 50 or pilot valve of this pressure compensating valve 7 can be located apart from the control pressure producing component 7 B. Since the joint 104 is located in the body 103 of the control pressure producing component 7 B in this pressure compensating valve 7 , the machine can be made more compact and the number of parts can be reduced.
- variable throttle valve 30 of the control pressure producing component 7 B has a structure similar to that of the variable throttle valve 30 in the pressure compensating valve 7 depicted in FIG. 6.
- the body 103 of the control pressure producing component 7 B in this pressure compensating valve 7 thus requires no check valve in a manner similar to that in the pressure compensating valve 7 depicted in FIG. 6.
- FIG. 9 depicts a sixth example of the structure of the pressure compensating valve 7 .
- This pressure compensating valve 7 is composed of only the compensator 7 A and the control pressure producing component 7 B.
- the compensator 7 A has a structure similar to that of the compensator 7 A depicted in FIG. 4.
- the control pressure producing component 7 B is equipped with a variable throttle valve 30 having a structure allowing the magnitude of the throttling to be manually altered.
- This variable throttle valve 30 has a vertical hole 106 in the body 105 , and a poppet type spool 107 is inserted into this vertical hole 106 .
- the top and bottom of the vertical hole 106 can be rendered communicable and are blocked by the vertical movement of the spool 107 .
- the top of a vertical hole 106 communicates through the oil passage 40 to the actuator port 64 of the operating valve 4 .
- the bottom of the vertical hole 106 communicates through the oil passage 37 equipped with the throttle 36 to the outlet port 62 of the operating valve 4 , and also communicates through the oil passage 38 to the pressure chamber 92 .
- An adjusting screw 108 is threaded into the top of the vertical hole 106 , and a spring 109 with weak elastic force is interposed between the adjusting screw 108 and the spool 107 .
- variable throttle valve 30 constructed in this manner, the pressured oil with the pressure Pa discharged from the outlet port 62 of the operating valve 4 flows through the oil passage 37 into the bottom of the vertical hole 106 .
- the spool 107 is pushed up, and part of the pressured oil with the pressure Pa flows into the oil passage 40 while constricted by the spool 107 .
- the pressure Pe of the pressure chamber 92 is set according to the amount of pressured oil flowing into the oil passage 40 , that is, according to the throttle level of the spool 107 .
- the upward moving stroke of the spool 107 defining the throttle level of the spool 107 can be adjusted by manually rotating the adjusting screw 108 .
- the pressure compensating valve 7 can thus alter the pressure Pe, that is, can alter the pressure compensation characteristics, when the screw 108 is rotated.
- the spool 107 is a poppet valve type, when pressured oil flows from the cylinder 5 into the oil passage 40 , the spool 107 is pushed down, blocking off the top and bottom of the vertical hole 106 from each other. In other words, the spool 107 functions as a check valve.
- FIG. 10 depicts a seventh example of the structure of the pressure compensating valve 7 .
- This pressure compensating valve 7 differs from the pressure compensating valve 7 depicted in FIG. 4 in terms of the structure of the compensator 7 A.
- the main valve 20 of the compensator 7 A depicted in FIG. 10 has a spool S comprising the unification of the valve component 66 and pushing component 67 depicted in FIG. 4.
- the spool S forms a communication hole 113 along the central axis, thereby allowing the outlet port 62 of the operating valve 4 and the pressure chamber 92 to communicate with each other.
- the pressured oil with the pressure Pa flowing from the outlet port 62 of the operating valve 4 flows through the communicating hole 113 into the pressure chamber 92 .
- the communicating hole 113 functions as the oil passage 37 in FIG. 4.
- a fixed throttle 113 a corresponding to the fixed throttle 36 depicted in FIG. 4 is formed at the end on the pressure chamber 92 side of the communication hole 113 .
- variable throttle valve 30 of the control pressure producing component 7 B has a structure similar to that of the variable throttle valve 30 depicted in FIG. 4.
- FIG. 11 depicts an eighth example of the structure of the pressure compensating valve 7 .
- the structure of the compensator 7 A in this pressure compensating valve 7 is similar to that of the pressure compensating valve 7 depicted in FIG. 10, and the structures of the control pressure producing component 7 B and pilot pressure producing component 7 C are similar to those of the pressure compensating valve 7 depicted in FIG. 6.
- FIG. 12 depicts a ninth example of the structure of the pressure compensating valve 7 .
- the structure of the compensator 7 A in this pressure compensating valve 7 is the same as that of the pressure compensating valve 7 depicted in FIG. 10, while the structure of the control pressure producing component 7 B and the location for attaching the joint 102 are the same as that of the pressure compensating valve 7 depicted in FIG. 7.
- FIG. 13 depicts a tenth example of the structure of the pressure compensating valve 7 .
- the structure of the compensator 7 A of this pressure compensating valve 7 is similar to that of the pressure compensating valve 7 depicted in FIG. 10, and the structure of the control pressure producing component 7 B and the position for attaching the joint 140 are the same as in the pressure compensating valve 7 depicted in FIG. 8.
- variable throttle valve 30 of the control pressure producing component 7 B has a structure similar to that of the variable throttle valve 30 in the pressure compensating valve 7 depicted in FIG. 6. The same effects in dispensing with the need to provide the body 103 of the control pressure producing component 7 B with a check valve can be obtained as in the pressure compensating valve 7 depicted in FIG. 6.
- FIG. 14 depicts an eleventh example of the structure of the pressure compensating valve 7 .
- the structure of the compensator 7 A of this pressure compensating valve 7 is similar to that of the pressure compensating valve 7 depicted in FIG. 10, and the structure of the control pressure producing component 7 B is similar to that of the pressure compensating valve 7 depicted in FIG. 9.
- FIG. 15 depicts a twelfth example of the structure of the pressure compensating valve 7 .
- the structure of the compensator 7 A of this pressure compensating valve 7 differs from that of the pressure compensating valve 7 A depicted in FIG. 4.
- the spool S of the main valve 20 of the compensator 7 A depicted in FIG. 15 is equipped with a piston 116 featuring the unification of the valve component 66 and the piston 73 depicted in FIG. 4, and a sliding element 117 located in the piston 116 .
- the piston 116 and the sliding element 117 are located along the central axis through the communication holes 118 and 119 , respectively.
- One end of the communication hole 119 in the sliding element 117 communicates through a check valve 120 to the communication hole 118 of the piston 116 , and the other end communicates through a throttle 119 a corresponding to the throttle 36 depicted in FIG. 2 to the pressure chamber 92 .
- the pressured oil with the pressure Po supplied from the outlet port 62 flows into the pressure chamber 92 through the communication hole 118 , a check valve 120 , a slit 121 formed around the check valve 120 , a port 122 passing through the peripheral wall of the sliding element 117 , the communicating hole 119 , and the throttle 119 a .
- the communication holes 118 and 119 function as the oil passage 37 depicted in FIG. 2.
- the check valve 120 thus has the same function as the check valve 39 depicted in FIG. 2. Accordingly, in this pressure compensating valve 7 , there is no need to provide the body 87 of the control pressure producing component 7 B with the check valve 39 depicted in FIG. 4, which makes the body 87 easier to fabricate.
- the oil passage 40 connected to the outlet port 33 of the variable throttle valve 30 was connected to the actuator port 64 (oil passage 6 a ) of the operating valve 4 depicted in FIG. 3, but this oil passage 40 may also be connected to the tank port 65 .
- FIG. 16 is a circuit diagram of oil pressure, depicting the structure of the unloading pressure control valve 10 .
- the unloading pressure control valve 10 is used to return the oil discharged from a hydraulic pump 1 directly to a tank to keep the hydraulic pump 1 in an unloaded state in a hydraulic system comprising, for example, a variable delivery pump 1 , an auxiliary hydraulic pump (pilot hydraulic pump) 2 , an operating valve 4 to which the oil discharged from the hydraulic pump 1 is supplied through an oil passage 3 , and a hydraulic cylinder (hydraulic actuator) 5 located opposite the operating valve 4 .
- the unloading pressure control valve 10 comprises a main valve 100 and an electromagnetic proportional pressure control valve 101 .
- the main valve 100 has a first pressure receiving component 123 , a second pressure receiving component 124 , a third pressure receiving component 125 , and a fourth pressure receiving component 126 .
- the main valve 100 sets the throttle level (unloading start pressure) between a first inlet port 127 and outlet port 128 by means of the elastic force of a spring 130 and the pressure acting on the first pressure receiving component 123 , second pressure receiving component 124 , third pressure receiving component 125 , and fourth pressure receiving component 126 .
- the first pressure receiving component 123 is connected to the variable delivery pump 1 along with the first inlet port 127 , and receives the discharge pressure P P of the hydraulic pump 1 .
- the second pressure receiving component 124 receives the maximum load pressure P LS by way of a throttle 129 .
- the third pressure receiving component 125 receives the control pressure Pg described below.
- the fourth pressure receiving component 126 is connected to the tank.
- the main valve 100 determines the unloading set pressure by means of the elastic force of the spring 130 and the pressure area of the second pressure receiving component 124 and third pressure receiving component.
- the main valve 100 does not require the spring 130 .
- the unloading start pressure can be set by just the difference between the pressure area of the second pressure receiving component 124 and the third pressure receiving-component.
- the control pressure Pg is given from the electromagnetic proportional pressure control valve 101 . That is, the electromagnetic proportional pressure control valve 101 introduces the pressured oil discharged from an auxiliary hydraulic pump 2 through the inlet port 132 , and the oil pressure resulting from a reduction in the pressure Pc of this pressured oil is output as the control pressure Pg.
- the control pressure Pg changes proportionally to the amount of electricity sent to the solenoid 133 .
- a sliding element 145 is slidably inserted into the left side of the valve body 140 of the main valve 100 , and the left end of a sleeve 148 is fitted to the right side of the valve body 140 .
- the sliding element 145 has a U-shaped cross section, and is brought into contact on the left end surface with an adjusting screw 147 threaded into the left end of the valve body 140 .
- the adjusting screw 147 is locked by a lock nut 148 .
- the interior of the sliding element 145 communicates through a hole 145 a to the tank.
- a spool 150 has a first small diameter component 151 forming a left half, a large diameter component 152 forming a central component, and a second small diameter component 153 forming a right half.
- the left tip of the first small diameter component 151 of the spool 150 is slidably inserted into the sliding element 145 .
- the large diameter component 152 is slidably inserted into a large diameter hole 154 in a sleeve 146 .
- the second small diameter component 153 is slidably inserted into a small diameter hole 155 in the sleeve 146 .
- the right end surface 150 a of the spool 150 forms the first pressure receiving component 123 depicted in FIG. 16.
- the left end surface 150 b of the spool 150 forms the fourth pressure receiving component 126 .
- the spool 150 is designed so that the cross sectional area of the second small diameter component 153 is a size equal to that obtained by subtracting the cross sectional area of the first small diameter component 151 from the cross sectional area of the large diameter component 152 .
- the right end of the sleeve 146 is positioned in the valve body 180 of the operating valve 4 .
- the sleeve 146 forms the first inlet port 127 depicted in FIG. 16 by opening the right end.
- the inlet port 127 communicates with the pump port 181 of the operating valve 4 .
- the sleeve 146 forms the outlet port 128 depicted in FIG. 16 at a position located slightly to the left of the right end opening.
- the outlet port 128 communicates with the tank port 182 of the operating valve 4 .
- the sleeve 146 further comprises a load pressure introduction port 157 and a control pressure introduction port 158 .
- the load pressure introduction port 157 introduces pressured oil with the maximum load pressure P LS
- the control pressure introduction port 158 introduces control pressure Pg through the electromagnetic proportional pressure control valve 101 .
- the load pressure introduction port 157 communicates through an annular space 159 , an oil hole 160 , and a fine hole 161 to a spring chamber 162 .
- the annular space 159 is formed between the inner peripheral surface of the sleeve 146 and the outer peripheral surface of the second small diameter component 153 of the spool 150 .
- the oil hole 60 is formed along the central axis of the spool 150 .
- the fine hole 161 passes diametrically through the spool 150 , forming the throttle 129 depicted in FIG. 16.
- control pressure introduction port 158 communicates with a space 163 formed between the large diameter component 152 of the spool 150 and the sleeve 146 .
- the right end surface 152 a of the spool large diameter component 152 located in the space 163 forms the third pressure receiving component 125 depicted in FIG. 16.
- the spring 130 depicted in FIG. 16 is located in the spring chamber 162 .
- the spring 130 is interposed between a spring receiver 162 a inserted into the first small diameter component 151 of the spool 150 and the right end surface of the sliding element 145 , and pushes the spool 150 to the right.
- the electromagnetic proportional pressure control valve 101 is disposed over the valve body 140 of the main valve 100 .
- a spool 167 for allowing the inlet port 132 and outlet port 135 depicted in FIG. 16 to communicate with each other and to be blocked off from each other is located in the valve body 166 of the electromagnetic proportional pressure control valve 101 .
- the top of the valve body 166 has a solenoid 133 that pushes the spool 167 down against the spring 134 .
- the inlet port 132 is connected to the auxiliary hydraulic pump 2 .
- the outlet port 135 communicates through an oil passage 168 to the control pressure introduction port 158 .
- control pressure Pg supplied from the electromagnetic proportional pressure control valve 101 acts on the right end surface 152 a of the large diameter component of the spool 150 serving as the third pressure receiving component 125 , by way of the oil passage 168 and the control pressure introduction port 158 , so that the spool 150 is pushed to the left.
- the spring 130 located in the spring chamber 162 pushes the spool 150 to the right.
- the load pressure P LS is introduced through the load pressure introduction port 157 , annular space 159 , oil hole 160 , and fine hole 161 (throttle 129 ) into the spring chamber 162 .
- the load pressure P LS thus acts on the left end surface 152 b of the large diameter component of the spool 150 which is the second pressure receiving component 124 , and the spool 150 is pushed to the right.
- a 1 area of right end surface 150 a of spool 150
- a 2 area of large diameter component 152 of spool 150
- a 3 area of left end surface 150 b of spool 150
- the pressure difference P P ⁇ P LS determines the unloading start pressure.
- the unloading pressure control valve 10 thus allows the unloading start pressure to be arbitrarily set by controlling the amount of electricity to the solenoid 133 of the electromagnetic proportional pressure control valve 101 to change the control pressure Pg.
- the main valve 100 of the unloading pressure control valve 10 is interposed between the hydraulic pump 1 and the tank.
- the pressure difference P P ⁇ P LS reaches the unloading start pressure, the oil discharged from the hydraulic pump 1 is returned to the tank during continuous operation.
- the electromagnetic proportional pressure control valve 101 of the unloading pressure control valve 10 produces pilot control pressure Pg resulting from the reduction of the discharge oil pressure Pc of the auxiliary hydraulic pump 2 . Meanwhile, in the main valve 100 , the operating start pressure (unloading start pressure) changes according to the control pressure Pg given by the electromagnetic proportional pressure control valve 101 .
- control signals to the solenoid 133 of the electromagnetic proportional pressure control valve 101 can be changed to set the unloading start pressure to the desired magnitude.
- FIG. 18 depicts another embodiment of the unloading pressure control valve relating to the present invention.
- This unloading pressure control valve 10 comprises an attachment block 185 , a piping joint 187 , and an oil pressure pilot valve 188 .
- the attachment block 185 is fixed to the upper surface of the valve body 140 .
- the piping joint 187 is screwed into a threaded hole 186 located in the attachment block 185 , and is thus secured.
- the oil pressure pilot valve 188 is manually operated.
- the threaded hole 186 passes through the control pressure introduction port 158 .
- the inlet port 188 b of the oil pressure pilot valve 188 is connected to the auxiliary hydraulic pump 2 .
- the outlet port 188 a is connected to the piping joint 187 .
- This unloading pressure control valve 10 allows the unloading start pressure to be arbitrarily set according to the control pressure Pg.
- the oil pressure pilot valve 188 which is the means for producing the control pressure Pg can also be disposed apart from the main valve 100 . It can thus be freely disposed, enabling manual remote control of the unloading start pressure, and the like.
- control pressure Pg acted as the force moving the spool 150 to the left (the direction passing through the first inlet port 127 and outlet port 128 of the main valve 100 ).
- control pressure Pg acts as the force moving the spool 150 to the right.
- the pressing force of the spring 130 acts in the direction opposite that described above (the direction in which the spool is pushed to the left).
- FIG. 19 depicts a hydraulic system featuring the use of two hydraulic pumps 1 A and 1 B.
- the hydraulic pumps 1 A and 1 B are connected to corresponding operating valves 4 A and 4 B by means of a switching valve 191 in a converged flow component 190 .
- a switching valve 192 switches between the communication and blockage of pressured oil, with a maximum load pressure P LS-A sensed by one shuttle valve 8 A, and pressured oil with a maximum load pressure P LS-B sensed by another shuttle valve 8 B.
- the switching valves 191 and 192 of the converged flow component 190 in this case allow the oil discharged by the hydraulic pumps 1 A and 1 B to converge, and also allow the pressured oil with the load pressures P LS-A and P LS-B to converge.
- the maximum load pressure P LS-A sensed by the shuttle valve 8 A is the highest among the plurality of hydraulic cylinders 5 driven by the hydraulic pump 1 A.
- the maximum load pressure P LS-B sensed by the shuttle valve 8 B is the highest among the plurality of hydraulic cylinders 5 driven by the hydraulic pump 1 B.
- the load pressure P LS-A is supplied to the unloading pressure control valve 10 and the volume control component (pump discharge pressure control means) 12 of the hydraulic pump 1 A.
- the load pressure P LS-A is also supplied through a check valve 193 A to the load pressure bleed valve 11 .
- the load pressure P LS-B is supplied to the unloading pressure control valve 10 and the volume control component 12 of the hydraulic pump 1 B.
- the load pressure P LS-B is also supplied through a check valve 193 B to the load pressure bleed valve 11 .
- the pressured oil with the maximum load pressure P LS-A is allowed to communicate with the tank during the operation of the one unloading pressure control valve 10 A. That is, the pressured oil with the maximum load pressure P LS-A is introduced from in front of the throttle 129 through the branched piping into the unloading pressure control valve 10 A, and this pressured oil is also output through a throttle 169 from the unloading pressure control valve 10 A so as to be returned to the tank.
- This allows the maximum load pressure P LS confined in the piping leading from the shuttle valve 8 A to the main valve 20 to escape to the tank, and prevents the discharge pressure P P of the hydraulic pump 1 A from increasing.
- the discharge pressure P P of the hydraulic pump 1 B can similarly be prevented from increasing during the operation of the other unloading pressure control valve 10 B.
- variable bleed valve 11 The structure of the variable bleed valve 11 relating to the present invention is described below with reference to FIG. 20.
- variable bleed valve 11 comprises a variable throttle valve 110 and an electromagnetic proportional pressure control valve 111 , as shown in the enlargement in FIG. 20.
- variable throttle valve 110 is operated so as to increase the area of the opening-between an inlet port 196 and an outlet port 197 by means of the elastic force of a spring 95 and the pilot pressure Pg acting on a pressure receiving component 194 , and is operated so as to reduce the area of the opening by means of the elastic force of a spring 198 .
- the electromagnetic proportional pressure control valve 111 introduces pressured oil with a standard pressure Pc discharged from the auxiliary hydraulic pump 2 into the inlet port 199 , and the pressure Pc of the pressured oil is reduced to the pilot pressure Pg.
- the pressured oil with the pilot pressure Pg is allowed to act on the pressure receiving component 194 of the variable throttle valve 110 by way of the outlet port 200 .
- the pilot pressure Pg changes proportionally to the amount of electricity to the solenoid 201 .
- variable bleed valve 11 is connected to a controller 300 .
- the controller 300 gives a corresponding control signal to the solenoid 201 of the electromagnetic proportional pressure control valve 111 based on operation commands such as a command to open the operating valve 4 by the operation of an operating lever (not shown in figure).
- FIG. 21 depicts the variable bleed valve 11 while mounted. It may also be seen from FIG. 21 that the variable bleed valve 11 is provided as a valve block along with a plurality of operating valves 4 and 4 . That is, the variable bleed valve 11 is attached by means of a support block 202 to the operating valve 4 located on the outermost side of the plurality of operating valves 4 joined in parallel. The symbol 4 a indicates the spool of the operating valve 4 .
- FIG. 22 is a cross section of line A-A in FIG. 21. It may be seen from FIG. 22 that the variable throttle valve 110 is such that the spool 206 is inserted into the spool hole 205 of the valve body 204 . The spool hole 205 is formed in the vertical direction.
- the spool 206 is interposed between the inlet port 196 and outlet port 197 of the variable throttle valve 110 .
- the spool 206 is such that downward force (the direction in which the area of the opening between the ports 196 and 197 is reduced) is urged by the spring 215 . Meanwhile, the upward force (the direction in which the area of the opening between the ports 196 and 197 is increased) is urged by the spring 195 in the pressure chamber 209 formed between the spool and an adjustment screw 62 .
- the bottom end surface of the spool 205 facing the pressure chamber 209 forms the pressure receiving component 194 depicted in FIG. 20.
- the elastic force of the spring 195 can be fine tuned by the operation of the adjustment screw 217 .
- the inlet port 196 communicates with the load pressure introduction hole 203 through a load pressure introduction oil passage 210 leading from the valve body 204 to the support block 202 .
- the outlet port 197 communicates with the tank through a tank oil hole 211 that opens into the attachment surface 204 of the valve body 204 .
- the electromagnetic proportional pressure control valve 111 is disposed on the upper surface of the aforementioned valve body 204 .
- the electromagnetic proportional pressure control valve 111 comprises a spool 214 and the solenoid 201 .
- the spool 214 is vertically disposed in the valve body 213 .
- the spool 214 is disposed coaxially relative to the spool 206 of the variable throttle valve 110 .
- the solenoid 201 pushes the spool 214 down against the spring 215 according to the amount of electricity.
- the outlet port 200 communicates with the pressure chamber 209 by way of an oil passage 216 located in the valve body 213 and an oil passage 212 located in the valve body 204 of the variable throttle valve 110 .
- the spring 215 is in contact with the upper tip of the spool 206 of the variable throttle valve 110 .
- the spring chamber 207 of the valve body 204 communicates with the tank by way of a tank oil hole 208 that opens into the attachment surface 204 a of the valve body 204 .
- variable bleed valve 11 The operation of the variable bleed valve 11 is described below.
- variable bleed valve 11 allows the amount of pressured oil with the maximum load pressure P LS that is bled off to be arbitrarily adjusted by controlling the amount of electricity to the electromagnetic proportional pressure control valve 111 .
- the rate of increase in the maximum load pressure P LS in the oil passage 9 can be arbitrarily adjusted by controlling the aforementioned amount of electricity.
- variable bleed valve 11 allows the start up response of the hydraulic cylinder 5 to be adjusted by controlling the amount of electricity to the solenoid 201 of the electromagnetic proportional pressure control valve 111 .
- the amount discharged by the hydraulic pump 1 is controlled to bleed off the pressured oil in the oil passage 9 for sensing the maximum load pressure P LS serving as the pilot pressure.
- the amount flowing in the load pressure sensing channel 9 is generally quite low.
- the pump pressure is controlled according to the pressure of the load pressure sensing passage 9 , whereas the pressure of the load pressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact.
- the amount discharged from the hydraulic pump 1 can be controlled with greater precision.
- the amount of pressured oil that is bled off can be arbitrarily adjusted by means of control signals output by a controller 300 described below, making such control easier to manage. Since pressured oil should be supplied by the application of electricity to the electromagnetic proportional pressure control valve 111 only when bleed off is needed, not only can pressured oil energy loss be further minimized, but electrical energy can also be economized.
- the aforementioned hydraulically operated device is equipped with a controller 300 connected to the variable bleed valve 11 as described above, and this controller 300 comprises, as shown in FIG. 20, a mode setting memory component 310 , a mode select setting component 320 , and a control signal output component 330 .
- the mode setting memory component 310 sets and stores a plurality of input-output relations according to the operating configuration of the hydraulic cylinder 5 .
- three different modes comprising an ordinary mode which is the ordinary operating state, a heavy operating mode requiring considerable force, and a more precise operating mode requiring highly precise manipulations are set and stored in terms of the input-output relations between the open command to the operating valve 4 and the control signals to the solenoid 201 of the electromagnetic proportional pressure control valve 111 , that is, the area of the opening of the variable throttle valve 110 .
- the area of the opening of the variable throttle valve 110 in terms of open commands to the same operating valve 4 is preset and stored so as to increase in ascending order from heavy operating mode, to ordinary mode, to precision operating mode.
- the mode select setting component 320 selects and sets one of the three input-output relations set and stored in the mode setting memory component 310 .
- This mode select setting component 320 selects and sets a corresponding input-output relation according to the operation of a mode select switch not shown in the figure and located in the driver seat of a hydraulic shove, for example.
- the control signal output component 330 converts the open command for the operating valve 4 based on the input-output relation selected by the mode select setting component 320 , and the converted control signal is given to the solenoid 201 of the electromagnetic proportional pressure control valve 111 .
- a control signal output from the controller 300 in response to an open command for the operating valve 4 can be modified according to the operating configuration of the hydraulic cylinder 5 .
- the area of the opening of the variable throttle valve 110 for open commands to the operating valve 4 can be further reduced.
- more pressured oil can be supplied to the hydraulic cylinder 5 and the hydraulic cylinder can be rapidly operated, even though the control input of the operating lever (not shown in figure) is the same.
- the area of the opening in the variable throttle valve 110 can be further increased for open commands to the operating valve 4 .
- less pressured oil can be supplied to the hydraulic cylinder 5 and the hydraulic cylinder can be moderately operated, even though the control input of the operating lever (not shown in figure) is the same.
- the hydraulic cylinder 5 is provided to drive the operating unit of the hydraulic shovel (such as a boom, arm, or bucket).
- a hydraulically operated device equipped with the variable bleed valve 11 can thus provide operating speeds and operating sensitivity for an operating unit that are suitable for the operating configuration of the aforementioned hydraulic shovel.
- the plurality of input-output relations set and stored in the mode setting memory component 310 are not limited to those depicted in FIG. 5.
- FIG. 24 is a graph depicting another example of input-output relations set and stored by the mode setting memory component 310 .
- the input-output relations depicted in FIG. 24 are designed so that the rate of change increases in the order from the heavy operating mode, to ordinary mode, to precision operating mode.
- the use of this mode setting selection means results in a different proportion of change in the speed by which the pressured oil in the load pressure sensing passage 9 is bled off into the tank, allowing the operating speeds and operating sensitivity of the hydraulic cylinder 5 to be set with even greater precision according to the operating configuration.
- a combination of the input-output relations depicted in FIGS. 23 and 24 can provide input-output relations such as that indicated by the broken lines in FIG. 23 for the heavy operating mode and precision operating mode in relation to ordinary mode. In this case, the input-output relations can be set even more precisely than those depicted in FIG. 24.
- variable throttle valve 110 is constructed in such a way as to increase the area of the opening between the inlet port 196 and outlet port 197 by means of the action of the pilot pressure Pg, but conversely it can also be constructed in such a way as to reduce the area of the aforementioned opening by means of the action of the pilot pressure Pg.
- variable throttle valve 110 is also constructed in such a way that the spool 206 is pressed in the cut-off direction (downward in FIG. 22) by the spring 198 and the spool 206 is pressed in the communicating direction (upward in FIG. 22) by the pressure in the pressure chamber 209 , but it can also be constructed in such a way that the elastic force of the spring 198 and the pressure of the pressure chamber 209 act in directions opposite those described above.
- variable bleed valve 11 is such that the spool 206 of the variable throttle valve 110 and the spool 214 of the electromagnetic proportional pressure control valve 111 are located coaxially, making it possible to achieve more compact shapes with a shorter lateral length. That is, when the lay out of the variable bleed valve 11 , for example, is like that depicted in FIG. 21, a more compact embodiment can be devised because the electromagnetic proportional pressure control valve 111 can be mounted further inside than the spring case 4 b of the operating valve 4 , that is, inside the surface defined by the spring case 4 b when a valve block is used in a generally right-angled parallelepiped form.
- the variable bleed valve 11 is such that the spring 215 of the electromagnetic proportional pressure control valve 111 is in contact with the upper end of the spool 206 of the variable throttle valve 110 .
- the operating force of the spool 206 is mechanically fed back to the spool 214 of the electromagnetic proportional pressure control valve 111 through the spring 215 when the spool 206 of the variable throttle valve 110 is operated.
- the operating characteristics (response) of the spool 206 of the variable throttle valve 110 can thus be improved, allowing more precise bleed off operations to be managed.
- variable bleed valve 11 is also designed to allow the elastic force of the spring 195 of the variable throttle valve 110 to be fine tuned by rotating the adjustment screw 217 .
- the machining precision of the various parts and the elastic force of the spring 198 used in the individual variable bleed valves 11 are not uniform.
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Abstract
A pressure compensating valve (7) comprises a main valve (20) that is operated in such a way as to increase the area of the opening between an inlet port (24) and an outlet port (25) by means of pressure acting on a first pressure receiving component (21), that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component (22) and pressure acting on a third pressure receiving component (23), and that is designed to allow the pressure (Pa) of the pressured oil flowing to the inlet port (24) to act on the first pressure receiving component (21) and the pressure (Pb) of the load driven by the pressured oil flowing from the outlet port (25) to act on the second pressure receiving component (22); and control pressure producer (7B) for allowing control pressure (Pe) resulting from a reduction in the pressure (Pa) of the inlet port (24) to act on the third pressure receiving component (23). An unloading pressure control valve comprises a main valve (100) that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure (PP) of the hydraulic pump (1) acting on a first pressure receiving component (123), to operate in the cut-off direction upon load pressure (PLS) to a second pressure receiving component (124), and to change the balance of the operating force in each direction by means of control pressure (Pg) acting on a third pressure receiving component (125); and control pressure producer (101) for producing the control pressure (Pg). A hydraulically operated device has a variable bleed valve (11) located in a load pressure sensing passage (9). This allows pressure compensated characteristics to be changed as desired. The unloading start pressure is also preset so as to improve response in terms of the hydraulic actuators. Energy loss can also be minimized, allowing the rapid start-up of the hydraulic actuators to be controlled, machines can be made smaller, and high precision control can be achieved.
Description
- The present invention relates to a pressure compensating valve, an unloading pressure control valve, and a hydraulically operated device.
- FIG. 25 depicts a hydraulically operated device described in Japanese Unexamined-Patent Application 1-247805.
- In this hydraulically operated device, a variable delivery pump A is connected to a low pressure hydraulic cylinder D via a pressure compensating valve B and a directional control valve (operating valve) C. The pump A is also connected to a high pressure cylinder D′ via a pressure compensating valve B′ and a directional control valve C′.
- An actuator E for changing the displacement volume and a flow regulating valve F for controlling the actuator E are attached to the hydraulic pump A.
- The higher load pressure among the load pressures that are produced during the operation of the cylinders D and DT is sensed by a shuttle valve G as the maximum load pressure PLS, and this maximum load pressure PLS is output as the pilot pressure to the flow regulating valve F.
- The flow regulating valve F controls the actuator E so that the discharge pressure PP of the pump A is always greater than the maximum load pressure PLS The cylinders D and D′ are jointly operated by the simultaneous operation of directional control valves C and C′ in the hydraulically operated device. At this time, the pressure compensating valve B controls the amount of oil supplied to the cylinder D so that the difference between the input pressure and the output pressure of the directional control valve C is constant, and the pressure compensating valve B′ similarly controls the amount of oil supplied to the cylinder D′ so that the difference between the input pressure and the output pressure of the directional control valve C′ is constant.
- The hydraulically operated device equipped with the pressure compensating valves B and B′ can prevent the disadvantage of pressured oil accumulating and being supplied to the cylinder with the lighter load among the operating valve cylinders D and D′.
- According to the aforementioned Japanese Unexamined Patent Application 1-247805, however, the discharge pressure of pump A decreases upon the supply of large amounts of pressured oil to the hydraulic cylinder D with the lower pressure during periods of considerable control input to the directional control valves C and C′. In such cases, the pressure difference before and after the pressure compensating valve B fails to reach the compensated pressure difference, and the pressure compensating valve B thus fails to achieve pressure compensation. That is, the pressure compensating valve B remains open.
- While the pressure compensating valve B fails to achieve pressure compensation, the amount of pressured oil supplied to the hydraulic cylinder D with the lower pressure is uncontrolled, so no pressured oil is supplied to the hydraulic cylinder D′ with the higher pressure, and the hydraulic cylinder D′ with the higher pressure is thus not operated. The operator must then operate the directional control valve C in the slightly open direction to control the flow rate to the hydraulic cylinder D with the lower pressure.
- To prevent such a situation from developing, the aforementioned hydraulically operated device is provided with a pressure difference sensing device H for sensing the pressure difference PP-PLS between the pressure PP of the pressured oil discharged from the hydraulic pump A and the maximum load pressure PLS, a control force set device I for setting the control force fc based on the pressure difference PP-PLS and the relationship depicted in FIG. 26, and an electromagnetic valve J that is operated by means of the output signals from the control force setting device I.
- The control force fc is given by the following equation.
- fc=f−a(P P −P LS )
- Where
- f: the pressing force of springs b and b− in pressure compensating valves B and B′
- α: constant
- The electromagnetic valve J allows pressured oil corresponding to the control force fc to act on the pressure receiving components of the pressure compensating valves B and B′ when the pressure difference PP-PLS is at or below the specific pressure difference Pm shown in FIG. 26.
- This allows the control force fc against the pressing force f of the aforementioned springs b and b− to be exerted on the springs in the pressure compensating valves B and B′. The force fc increases the discharge pressure of the pump A by increasing the flow resistance of the pressure compensating valves B and B′, allowing pressured oil to be supplied to the hydraulic cylinder D′ with the higher pressure too.
- When the cylinders D and D′ are cylinders that operate an operating device in construction machinery (such as a hydraulic shovel boom, arm, or bucket), the pressure compensation characteristics of the pressure compensating valves B and B′ are preferably modified in some cases to improve the operating characteristics, depending on the operating configuration of the aforementioned operating device.
- A technique that is capable of changing the throttle levels for each pressure compensating valve and that is capable of suitably changing the pressure difference before and after the directional control valves C and C′ has been disclosed in the aforementioned patent publication. That is, in this technique, electromagnetic valves J as described above are provided for the pressure compensating valves B and B′, and the control force fc for the pressure compensating valves B and B′ are individually adjusted by these electromagnetic valves J. Accordingly, the throttle levels of the pressure compensating valves B and B′ are individually changed; that is, the pressure differences before and after the direction control valves C and C′ are different from each other.
- A state in which the required flow rate is distributed completely irrespective of load is also referred to in particular as a fully compensated state.
- The conventional devices described above suffer from the following drawbacks, however.
- In some cases, pressure compensation is not possible when the mechanism for producing control force fc to modify the pressure compensation characteristics malfunctions. Furthermore, the electromagnetic valves J are operated by computations after the pressure difference has been sensed by a pressure difference detector21H, resulting in poor response.
- In view of the foregoing, a first object of the present invention is to provide a pressure compensating valve that allows the pressure compensation characteristics to be arbitrarily modified, that has good response, and that is highly reliable.
- FIG. 27 depicts a hydraulically operated device described in Japanese Unexamined Patent Application 4-250226. When the operating device A in this hydraulically operated device is operated, a flow regulating valve (operating valve) B is operated, by means of the pilot pressure produced by the operating device A, to an extent corresponding to the extent to which the operating device A has been operated, and the discharged pressured oil from a hydraulic pump D is consequently supplied to a hydraulic cylinder (hydraulic actuator) C.
- A pressure compensating valve E for keeping the pressure difference before and after the flow regulating valve B at a constant level is located between the hydraulic pump D and the flow regulating valve (operating valve) B.
- An operating device A′, flow regulating valve (operating valve) B′, hydraulic motor (hydraulic actuator) C′, and pressure compensating valve E′ each correspond to the operating device A, flow regulating valve (operating valve) B, hydraulic cylinder C, and pressure compensating valve E.
- An unloading pressure control valve F is connected in parallel to the hydraulic pump D. The higher pressure between the load pressure acting on the hydraulic cylinder C and the load pressure acting on the hydraulic motor C′ is sensed as the maximum load pressure by a shuttle valve G, and this maximum load pressure is allowed to act on the unloading pressure control valve F.
- The unloading pressure control valve F is provided to return the discharged oil from the hydraulic pump D to the tank. The amount of the aforementioned discharged oil returned by the unloading pressure control valve F is set by the difference between the maximum load pressure and the discharge pressure of the hydraulic pump D, and by control signals output from a control unit J.
- A computer H connected to the control unit J computes the difference ΔPLS between the discharge pressure of the hydraulic pump D and the load pressure of the hydraulic cylinder C or hydraulic motor C′ based on the functional relation shown in FIG. 28 and the output of sensors a1, a2 and a1′, a2′ for sensing the control input of the operating devices A and A′.
- The function shown in FIG. 28 defines a relation in which the pressure difference ΔPLS increases proportionally until the control input St of the operating device A reaches a set value, and the pressure difference ΔPLS stays at a
value ΔP LS1 when the control input St is at or beyond the set value. - When the control input St is 20%, for example, the pressure difference ΔPLS is computed by the computer H, so a control signal corresponding to a pressure difference ΔPLS2 is output from the control unit J, and the unloading start pressure of the unloading pressure control valve F is set to pressure difference ΔPLS2. As a result, the amount of pressured oil supplied through the pressure compensating valve E′ and flow regulating valve B′ to the hydraulic motor C′ is the amount defined by the pressure difference ΔPLS2.
- FIG. 29 shows the relation between the amount of oil Q supplied to the hydraulic motor C′ and the pressure difference AP before and after the flow regulating valve B′ when the control input St is 20%.
- As shown in FIG. 29, the pressure compensating valve E′ supplies pressured oil in a constant oil amount q2 to the hydraulic motor C′ so that the pressure difference AP of the flow regulating valve B′ is kept at a constant pressure difference ΔPc+ΔPLS (ΔPLS is the pressure loss of the pressure compensating valve E′). However, while the pressure difference ΔP has not yet reached the constant pressure difference ΔPc+ΔPLS (compensated pressure difference), the pressured oil is supplied to the hydraulic motor C′ in the oil amount q1 defined by the unloading start pressure ΔPLS2 of the unloading pressure control valve F.
- Thus, according to this hydraulically operated device, when the control input of the operating device A is set to about 20% for moderate acceleration of the hydraulic motor C′, the amount of oil supplied to the hydraulic motor C′ is limited to the amount of oil q1 defined by the unloading start pressure ΔPLS2, and the hydraulic motor C′ is thus moderately accelerated.
- Furthermore, in the case of the load sensing circuit of a variable delivery pump, when the unloading start pressure of the unloading pressure control valve F is pre-modified, the amount of pressured oil discharged from the hydraulic pump D is increased in advance. The response of the hydraulic cylinder C when operated by the operating unit A is thus better.
- The unloading start pressure of the unloading pressure control valve F is variable. However, the unloading start pressure is set through the computer H and the control unit J. It is accordingly always set after the output from the sensors a1, a2 and a1′, a2′ of the operating devices A and A′, and a resulting problem is the poor response in terms of the hydraulic cylinder C or the hydraulic motor C′. More specifically, when the fluctuations in the load pressure of the hydraulic cylinder C or hydraulic motor C′ are estimated, the unloading start pressure is hopefully pre-modified rapidly irrespective of the control input of the operating devices A and A′. For the reasons described above, however, the unloading start pressure is difficult to modify in advance.
- In view of the foregoing, a second object of the present invention is to provide an unloading pressure control valve allowing the unloading start pressure to be preset so as to improve the response in terms of a hydraulic actuator.
- A pump discharge pressure control means for controlling the displacement volume of a hydraulic pump (discharge volume per revolution) is provided in a hydraulically operated device in which the pressured oil discharged from a variable delivery pump is supplied to a hydraulic actuator such as a hydraulic cylinder by the operation of an operating valve. This pump discharge pressure control means is designed so as to control the displacement volume of a hydraulic pump based on the discharge pressure of a hydraulic pump and the load pressure acting on a hydraulic actuator, so that the aforementioned discharge pressure is greater by a specific pressure than the aforementioned load pressure.
- According to the hydraulically operated device equipped with the pump discharge pressure control means, when the load pressure is increased during the operation of the operating valve, the displacement volume of the hydraulic pump immediately increases to a magnitude corresponding to the load pressure. The actuator is also connected via a pressure compensating valve. A flow rate corresponding to the control input of the operating valve can thus be supplied, irrespective of the magnitude of the load pressure, to the actuator.
- To be supplied at flow rate corresponding to the control input is, in other words, a matter of the action of pressure corresponding to the load. The control input of the operating valve at this time and certain actuator conditions sometimes result in rapid start up with shocks.
- When the aforementioned hydraulic actuator is a hydraulic motor or cylinder driving an operating unit in construction machinery (such as the revolving superstructure, boom, arm, or bucket in the case of a hydraulic shovel, for example), the rapid start up of the aforementioned hydraulic actuator results in lower operating performance, depending on the operating configuration.
- Hydraulically operated devices such as the following have been proposed in patent publications.
- That is, in the hydraulically operated device proposed in Japanese Unexamined Patent Application 9-222101, for example, a bleed valve is connected to the discharge channel of the aforementioned hydraulic pump, and part of the pressured oil discharged by the hydraulic pump is bled through the bleed valve to the tank.
- According to the hydraulically operated device described in this patent publication, the rapid start up of the hydraulic actuator is controlled, resulting in better operating performance.
- However, the bleed valve used in the hydraulically operated device of the aforementioned patent publication bleeds off part of the pressured oil discharged from the hydraulic pump to the tank. In other words, a large amount of the pressured oil that is supposed to be supplied to the hydraulic actuator ends up being returned to the tank when bled off. This results in significant energy loss.
- Other resulting problems are the need for large-scale machines because of the large amounts of pressured oil that are bled off, poor sensitivity, and difficulties in achieving high-precision control.
- In view of the foregoing, a third object of the present invention is to provide a hydraulically operated device that allows energy loss to be minimized to control rapid start up of hydraulic actuators, and that also allows machinery to be made more compact and high-precision control to be achieved.
- Another object of the present invention is to simultaneously achieve the first and second objects.
- Still another object of the present invention is to simultaneously achieve the first and third objects.
- Yet another object of the present invention is to simultaneously achieve the second and third objects.
- And finally another object of the present invention is to simultaneously achieve the first, second, and third objects.
- To achieve the first object, the first of the present inventions is a pressure compensating valve through which passes pressured oil that is fed from a
hydraulic pump 1 to ahydraulic actuator 5, characterized by comprising amain valve 20 that is operated in such a way as to increase the area of the opening between aninlet port 24 and anoutlet port 25 by means of pressure acting on a firstpressure receiving component 21, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a secondpressure receiving component 22 and pressure acting on a thirdpressure receiving component 23, and that is designed to allow the pressure Pa of the pressured oil flowing to theinlet port 24 to act on the firstpressure receiving component 21 and the pressure Pb of theload 5 driven by the pressured oil flowing from theoutlet port 25 to act on the secondpressure receiving component 22; and control pressure producing means 7B for allowing control pressure Pe resulting from a reduction in the pressure Pa of theinlet port 24 to act on the thirdpressure receiving component 23. - The first invention allows the desired pressure compensation characteristics to be obtained by changing the control pressure Pe because the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe.
- Because the control pressure Pe resulting from a reduction in the pressure of the
inlet port 24 is allowed to act on the thirdpressure receiving component 23 of themain valve 20, fluctuations in the control pressure Pe also correspond to fluctuations in the pressure of theinlet port 24. The pressure compensation characteristics are thus unaffected by the pressure fluctuation of theinlet port 24 of themain valve 20. - To achieve the second object described above, the second invention is an unloading pressure control valve for introducing discharged pressured oil from a
hydraulic pump 1 to a tank according to the pressure difference between the discharge pressure PP of thehydraulic pump 1 and the load pressure PLS of ahydraulic actuator 5, characterized by comprising amain valve 20 that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure PP of thehydraulic pump 1 acting on a firstpressure receiving component 123, to operate in the cut-off direction upon load pressure PLS to a secondpressure receiving component 124, and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a thirdpressure receiving component 125; and control pressure producing means 101 for producing the control pressure Pg. - The second invention allows the unloading start pressure to be set by means of the control pressure Pg acting on the third
pressure receiving component 125. The control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from thehydraulic pump 1 can be increased in advance to improve the response in terms of thehydraulic actuator 5. - To achieve the third object described above, the third invention is a hydraulically operated device comprising a plurality of
hydraulic actuators 5 to which pressured oil discharged from avariable delivery pump 1 is supplied viapressure compensating valves 7 anddirectional control valves 4; means for outputting pressure PLS to a loadpressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators; and pump discharge pressure control means for controlling the discharge pressure of thehydraulic pump 1 based on the pressure PLS; wherein the hydraulically operated device is characterized in that avariable bleed valve 11 is located in the loadpressure sensing passage 9. - The third invention allows the amount discharged from the
hydraulic pump 1 to be controlled by bleeding off the pressured oil in the loadpressure sensing passage 9. The amount flowing in the loadpressure sensing channel 9 is generally quite low. The pump pressure is controlled according to the pressure of the loadpressure sensing passage 9, whereas the pressure of the loadpressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from thehydraulic pump 1 can be controlled with greater precision. - To achieve the first and second objects described above, the fourth of the inventions is a hydraulically operated device comprising a pressure compensating valve through which passes pressured oil that is fed from a hydraulic pump1 to a hydraulic actuator 5; and an unloading pressure control valve for introducing discharged pressured oil from the hydraulic pump 1 to a tank according to the pressure difference between the discharge pressure PP of the hydraulic pump 1 and the load pressure PLS of the hydraulic actuator 5; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve 7 itself comprising a pressure compensated main valve 20 that is operated in such a way as to increase the area of the opening between an inlet port 24 and an outlet port 25 by means of pressure acting on a first pressure receiving component 21 for a pressure compensating valve, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component 22 for a pressure compensating valve and pressure acting on a third pressure receiving component 23 for a pressure compensating valve, and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port 24 to act on the first pressure receiving component 21 for a pressure compensating valve and the pressure Pb of the load 5 driven by the pressured oil flowing from the outlet port 25 to act on the second pressure receiving component 22 for a pressure compensating valve, and control pressure producing means 7B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port 24 to act on the third pressure receiving component 23 for a pressure compensating valve; and an unloading pressure control valve 10 itself comprising a main valve 100 for an unloading pressure control valve, that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure PP of the hydraulic pump 1 acting on a first pressure receiving component 123 for an unloading pressure control valve, to operate in the cut-off direction upon load pressure PLS to a second pressure receiving component 124 for an unloading pressure control valve, and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component 125 for an unloading pressure control valve, and control pressure producing means 101 for producing the control pressure Pg.
- According to the fourth invention, the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.
- Because the control pressure Pe resulting from a reduction in the pressure of the
inlet port 24 is allowed to act on the thirdpressure receiving component 23 for a pressure compensating valve in themain valve 20 for a pressure compensating valve, the control pressure Pe also fluctuates according to the fluctuations in the pressure of theinlet port 24. The pressure compensation characteristics are thus unaffected by the pressure fluctuations in theinlet port 24 of themain valve 20. - Furthermore, the unloading start pressure can be set by means of the control pressure Pg acting on the third
pressure receiving component 125 for an unloading pressure control valve. The control pressure Pg is produced by the control pressure producing means. The unloading start pressure can thus be preset by the control pressure producing means, and the amount of pressured oil discharged from thehydraulic pump 1 can be increased in advance to improve the response in terms of thehydraulic actuator 5. - To achieve the first and third objects described above, the fifth of the inventions is a hydraulically operated device comprising a plurality of hydraulic actuators5 to which pressured oil discharged from a variable delivery pump 1 is supplied via pressure compensating valves and directional control valves 4; means 8 for outputting pressure PLS to a load pressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators 5; and pump discharge pressure control means 12 for controlling the discharge pressure of the hydraulic pump 1 based on the pressure PLS; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve 7 itself comprising a main valve 20 that is operated in such a way as to increase the area of the opening between an inlet port 24 and an outlet port 25 by means of pressure acting on a first pressure receiving component 21, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component 22 and pressure acting on a third pressure receiving component 23, and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port 24 to act on the first pressure receiving component 21 and the pressure Pb of the load 5 driven by the pressured oil flowing from the outlet port 25 to act on the second pressure receiving component 22, and control pressure producing means 7B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port 24 to act on the third pressure receiving component 23; and a variable bleed valve 11 is located in the load pressure sensing passage 9.
- According to the fifth invention, the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.
- Because the control pressure Pe resulting from a reduction in the pressure of the
inlet port 24 is allowed to act on the thirdpressure receiving component 23 of themain valve 20, the control pressure Pe also fluctuates according to the fluctuations in the pressure of theinlet port 24. The pressure compensation characteristics are thus unaffected by the pressure fluctuations in theinlet port 24 of themain valve 20. - Furthermore, the amount discharged from the
hydraulic pump 1 can be controlled by bleeding off the pressured oil in the loadpressure sensing passage 9. The amount flowing in the loadpressure sensing channel 9 is generally quite low. The pump pressure is controlled according to the pressure of the loadpressure sensing passage 9, whereas the pressure of the loadpressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from thehydraulic pump 1 can be controlled with greater precision. - To achieve the second and third objects described above, the sixth of the present inventions is a hydraulically operated device comprising a plurality of hydraulic actuators5 to which pressured oil discharged from a variable delivery pump 1 is supplied via pressure compensating valves 7 and directional control valves 4; means 8 for outputting pressure PLS to a load pressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators 5; pump discharge pressure control means 12 for controlling the discharge pressure of the hydraulic pump 1 based on the pressure PLS; and an unloading pressure control valve for introducing discharged pressured oil from the hydraulic pump 1 to a tank according to the pressure difference between the discharge pressure PP of the variable delivery pump 1 and the load pressure PLS of the hydraulic actuators 5; wherein the hydraulically operated device is characterized by comprising an unloading pressure control valve 10 itself comprising a main valve 100 that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure PP of the hydraulic pump 1 acting on a first pressure receiving component 123, to operate in the cut-off direction upon load pressure PLS to a second pressure receiving component 124, and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component 125, and control pressure producing means 101 for producing the control pressure Pg; and a variable bleed valve 11 is located in the load pressure sensing passage 9.
- According to the sixth invention, the unloading start pressure can be set by means of the control pressure Pg acting on the third
pressure receiving component 25. The control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from thehydraulic pump 1 can be increased in advance to improve the response in terms of thehydraulic actuator 5. - The amount discharged from the
hydraulic pump 1 can be controlled by bleeding off the pressured oil in the loadpressure sensing passage 9. The amount flowing in the loadpressure sensing channel 9 is generally quite low. The pump pressure is controlled according to the pressure of the loadpressure sensing passage 9, whereas the pressure of the loadpressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from thehydraulic pump 1 can be controlled with greater precision. - To achieve the first, second, and third objects described above, the seventh of the present inventions is a hydraulically operated device comprising a plurality of hydraulic actuators5 to which pressured oil discharged from a variable delivery pump 1 is supplied via pressure compensating valves and directional control valves 4; means 8 for outputting pressure PLS to a load pressure sensing passage 9 according to the maximum load pressure among the load pressures acting on the actuators 5; pump discharge pressure control means 12 for controlling the discharge pressure of the variable delivery pump 1 based on the pressure PLS; and an unloading pressure control valve for introducing discharged pressured oil from the variable delivery pump 1 to a tank according to the pressure difference between the discharge pressure PP of the variable delivery pump 1 and the load pressure PLS of the hydraulic actuators 5; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve 7 itself comprising a pressure compensated main valve 20 for a pressure compensating valve, that is operated in such a way as to increase the area of the opening between an inlet port 24 and an outlet port 25 by means of pressure acting on a first pressure receiving component 21 for a pressure compensating valve, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component 22 for a pressure compensating valve and pressure acting on a third pressure receiving component 23 for a pressure compensating valve, and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port 24 to act on the first pressure receiving component 21 for a pressure compensating valve and the pressure Pb of the load 5 driven by the pressured oil flowing from the outlet port 25 to act on the second pressure receiving component 22 for a pressure compensating valve, and control pressure producing means 7B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port 24 to act on the third pressure receiving component 23 for a pressure compensating valve; and an unloading pressure control valve 10 itself comprising a main valve 100 for an unloading pressure control valve, that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure PP of the hydraulic pump 1 acting on a first pressure receiving component 123 for an unloading pressure control valve, to operate in the cut-off direction upon load pressure PLS to a second pressure receiving component 124 for an unloading pressure control valve, and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component 125 for an unloading pressure control valve, and control pressure producing means 101 for producing the control pressure Pg; and a variable bleed valve 11 is located in the load pressure sensing passage 9.
- According to the seventh invention, the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.
- Because the control pressure Pe resulting from a reduction in the pressure of the
inlet port 24 is allowed to act on the thirdpressure receiving component 23 for a pressure compensating valve in themain valve 20 for a pressure compensating valve, the control pressure Pe also fluctuates according to fluctuations in the pressure of theinlet port 24. The pressure compensation characteristics are thus unaffected by the fluctuations in theinlet port 24 of themain valve 20 for a pressure compensating valve. - Furthermore, the unloading start pressure can be set by means of the control pressure Pg acting on the third
pressure receiving component 125 for an unloading pressure control valve. The control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from thehydraulic pump 1 can be increased in advance to improve the response in terms of thehydraulic actuators 5. - The amount discharged from the
hydraulic pump 1 can be controlled by bleeding off the pressured oil in the loadpressure sensing passage 9. The amount flowing in the loadpressure sensing channel 9 is generally quite low. The pump pressure is controlled according to the pressure of the loadpressure sensing passage 9, whereas the pressure of the loadpressure sensing passage 9 is the pressure corresponding to the load pressure of the actuators and thus reacts exactly to the fluctuations in the load pressure of the actuators. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from thehydraulic pump 1 can be controlled with greater precision. - FIG. 1 is a circuit diagram of the oil pressure in a hydraulically operated device relating to the present invention;
- FIG. 2 is a circuit diagram of oil pressure, depicting the structure of a pressure compensating valve relating to the present invention;
- FIG. 3 is a longitudinal cross section depicting the attachment of a pressure compensating valve and an operating valve relating to the present invention;
- FIG. 4 is a longitudinal cross section, depicting the structure of a pressure compensating valve relating to the present invention;
- FIG. 5 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 6 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 7 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 8 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 9 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 10 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 11 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 12 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 13 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 14 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 15 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;
- FIG. 16 is a circuit diagram of oil pressure, depicting the structure of an unloading pressure control valve relating to the present invention;
- FIG. 17 is a cross section depicting a specific structure for an unloading pressure control valve relating to the present invention;
- FIG. 18 is a cross section depicting another embodiment of an unloading pressure control valve relating to the present invention;
- FIG. 19 is a circuit diagram of oil pressure in another hydraulic system involving the application of an unloading pressure control valve relating to the present invention;
- FIG. 20 is a circuit diagram of oil pressure, depicting an enlargement of the structure of a variable bleed valve used in the hydraulically operated device of FIG. 1;
- FIG. 21 depicts an embodiment with a variable bleed valve attached to the hydraulically operated device in FIG. 1;
- FIG. 22 is a cross section of line A-A in FIG. 21;
- FIG. 23 is a graph depicting an example of the relation between input and output when a mode set memory means has been set and stored;
- FIG. 24 is a graph depicting another example of the relation between input and output when a mode set memory means has been set and stored;
- FIG. 25 is a circuit diagram of oil pressure, depicting the structure of a conventional hydraulic device equipped with a pressure compensating valve;
- FIG. 26 is a graph depicting the relation between pressure difference and control force;
- FIG. 27 is a circuit diagram of oil pressure in a conventional hydraulically operated device in which an unloading pressure control valve is used;
- FIG. 28 is a graph depicting the relation between the pressure difference and the control input of an operating device; and
- FIG. 29 is a graph depicting the relation between the pressure difference before and after a flow regulating valve and the amount of oil Q supplied to a hydraulic motor.
- Embodiments of the present invention are described in detail below with reference to the attached drawings.
- FIG. 1 depicts an embodiment of a hydraulically operated device relating to the present invention. The hydraulically operated device can be used for a hydraulic shovel, for example.
- The hydraulically operated device comprises a
variable delivery pump 1, auxiliaryhydraulic pump 2, a plurality of closed center operating valves (directional control valves) 4 to which the oil discharged from thehydraulic pump 1 is supplied through anoil passage 3, and a plurality ofhydraulic cylinders 5 corresponding to each operatingvalve 4. - The head oil chambers of the
hydraulic cylinders 5 are connected by means ofoil passages 6 a andpressure compensating valves 7 to theoperating valves 4, and the bottom oil chambers are connected by means of a pressure compensating valve not shown in the figure in anoil passage 6 b to theoperating valves 4. A pressure compensating valve is in fact interposed in theoil passage 6 b, but thus pressure compensating valve has been left out in FIG. 1 to avoid complicating the drawing. - The load pressure P1 of the
hydraulic cylinders 5 connected thereto act on each of theoil passages 6 a. The maximum load pressure among the load pressures P1 acting on theseoil passages 6 a are sensed as the maximum load pressure PLS by ashuttle valve 8, and the sensed maximum load pressure PLS is allowed by means of an oil passage (load pressure sensing passage) 9 to act on thehydraulic pump 1,pressure compensating valves 7, unloadingpressure control valve 10, andvariable bleed valve 11. A fixed throttle 13 is interposed between the tank and theoil passage 9 into which the pressured oil with the maximum load pressure PLS is introduced. - A pump discharge pressure control means12 is attached to the
hydraulic pump 1. The pump discharge pressure control means 12 introduces the discharge pressure PP of thehydraulic pump 1 and the maximum load pressure PLS′ and controls the displacement volume of thehydraulic pump 1 so that the discharge pressure PP is always slightly higher than the maximum load pressure PLS. - The structure of the
pressure compensating valve 7 relating to the present invention is described below with reference to FIG. 2. Thepressure compensating valve 7 is composed of acompensator 7A, a controlpressure producing component 7B, and a pilotpressure supply component 7C. - The
compensator 7A has amain valve 20. Themain valve 20 comprises a firstpressure receiving component 21, a secondpressure receiving component 22, and a thirdpressure receiving component 23. The pressure Pa acting on the firstpressure receiving component 21 acts in such a way as to increase the area of the opening between theinlet port 24 andoutlet port 25. The pressure Pb acting on the secondpressure receiving component 22 and the pressure Pc acting on the thirdpressure receiving component 23 act in such a way as to reduce the area of the opening along with the elastic force of aspring 26. - The
inlet port 24 is connected to the outlet port of the operatingvalve 4 depicted in FIG. 1. The pressure Pa of theinlet port 24 acts on the firstpressure receiving component 21 via anoil passage 27. Theoutlet port 25 is connected to theoil passage 6 a through aload check valve 28. - A
shuttle valve 29 senses the load pressure P1 acting on theoil passage 6 a and the greater pressure Pb among the maximum load pressure PLS, and allows the pressure Pb to act on the secondpressure receiving component 22 of themain valve 20. - The control
pressure producing component 7B has avariable throttle valve 30. The-variable throttle valve 30 is operated in such a way as to reduce the area of the opening between aninlet port 32 and anoutlet port 33 by means of the elastic force of aspring 31. It is also operated in such a way as to increase the area of the opening by means of the elastic force of aspring 35 and the pilot pressure P2 acting on apressure receiving component 34. - In ordinary cases, the
spring 35 is used only in the initial fine tuning of thevariable throttle valve 30, and is not indispensable. The tank port pressure is allowed to act constantly on thespring 31 to ensure that thevariable throttle valve 30 is operated more rapidly. - The
inlet port 32 of thevariable throttle valve 30 is connected to theinlet port 24 of themain valve 20 by way of anoil passage 37 equipped with athrottle 36. The pressure Pe of theinlet port 32 of thevariable throttle valve 30 acts on the thirdpressure receiving component 23 of themain valve 20. - The
outlet port 33 of thevariable throttle valve 30 is connected to theoil passage 6 a by way of anoil passage 40 equipped with acheck valve 39. The pressure Pe acting on the thirdpressure receiving component 23 of themain valve 20 is thus determined by the pressure Pa of theinlet port 24 of themain valve 20, the load pressure P1, and the throttle levels of thethrottle 36 and thevariable throttle valve 30. Pe=Pa when thevariable throttle valve 30 is closed. - The pilot pressure Pd is given as the output pressure of an electromagnetic proportional
pressure control valve 50 located in the pilotpressure supply component 7C. The electromagnetic proportionalpressure control valve 50 introduces the pressured oil discharged from the pilothydraulic pump 2 depicted in FIG. 1 to aninlet port 52. The pressure Pa of this pressured oil is lowered to the pilot pressure Pd by means of the electricity applied to asolenoid 53. The pilot pressure Pd displays a magnitude proportional to the amount of electricity to thesolenoid 53. - When zero electricity is supplied to the
solenoid 53, theoutlet port 55 communicates with thetank port 56 by means of the elastic force of aspring 54, as shown in the figure. The pressure Pd acting on thepressure receiving component 34 of thevariable throttle valve 30 is thus zero. With this, theinlet port 32 andoutlet port 33 of thevariable throttle valve 30 are blocked off from each other by the elastic force of thespring 31, both sides of which are acted upon by the tank port pressure. - The discharge pressure of the pilot
hydraulic pump 2 is held constant by constant pressure means not shown in the figure. - The specific structures of the operating
valve 4 andpressure compensating valve 7 are described below. - As noted above,
pressure compensating valves 7 are interposed not only inoil passages 6 a but also inoil passages 6 b in the hydraulically operated device in FIG. 1. - FIG. 3 depicts an example of the structure of an operating
valve 4 by which pressured oil is selectively supplied to the twopressure compensating valves 7 described above. - The
operating valve 4 has a structure in which abody 60 is provided with aspool 61, pairs of left andright outlet ports 62, pairs of left andright pump ports 63, pairs of left andright actuator ports 64, and pairs of left and right oftank ports 65. - The
spool 61 blocks all of theports 62 through 65 in the center valve state depicted in the figure. When thespool 61 moves left from the center valve state, theoutlet ports 62 on one side communicate with thepump ports 63, and theactuator ports 64 on the other side communicate with thetank ports 65. When thespool 61 moves right from the center valve state, theoutlet ports 65 on the other side communicate with thepump ports 63, and theactuator ports 64 on the first side communicate with thetank ports 65. - The
main valve 20 located in thecompensator 7A of thepressure compensating valve 7 has a valve component 66A interposed between theactuator port 64 and theoutlet port 62 of the operatingvalve 4, and apressing component 67 connected to the valve component 66A. - FIG. 4 depicts an enlargement of the
pressure compensating valve 7. As shown in FIG. 4, thevalve component 66 comprises ahollow component 68 open at the left end, ahole 69 open in the outer peripheral surface through thehollow component 68, and aseat surface 71 that presses into contact with aseat 70 formed in thebody 60. Pressure-receivingsurfaces valve component 66 form the firstpressure receiving component 21 of themain valve 20 depicted in FIG. 2, and thehole 69 of thevalve component 66 forms theoutlet port 25 of themain valve 20. Theentire valve component 66 functions as theload check valve 39 depicted in FIG. 2. - The
pressing component 67 is positioned on an extension of the central axis of thevalve component 66, and comprises apiston 73 that slides to the left and right in asleeve 72 fixed to thebody 60, a slidingelement 74 that slides to the left and right in thepiston 73, and a spring 26 (see FIG. 2) interposed between thesleeve 72 and the slidingelement 74. - An
annular space 75 into which the pressured oil with the maximum 1 load pressure PLS (see FIG. 2) is introduced is formed between thebody 60 and thesleeve 72. The pressured oil with the maximum load pressure PLS introduced into thisannular space 75 flows into a steppedhole 80 in the slidingelement 74 through afine hole 76 located in thesleeve 72, an annular groove 77, ahole 78 located in thepiston 73, and aninlet port 79 located in the slidingelement 74, and acts on the right side of abore 81 located in the steppedhole 80. - Meanwhile, the pressured oil in the
actuator port 64 of the operatingvalve 4 depicted in FIG. 3, that is, the pressured oil with the pressure load P1 flowing through theoil passage 6 a, flows through aninlet port 82 located in the left end of thepiston 73 and into the steppedhole 80 of the slidingelement 74, and acts on the left side of thebore 81. - When the relation between the pressures PLS and P1 is such that PLS is greater than P1, the
bore 81 rotates to the left position of theoutlet port 84, and when PLS is less than P1, thebore 81 rotates to the right position of theoutlet port 84. - As shown in FIGS. 1 and 2, PLS<P1 is a state of transition, where the pressure PLS increases so that PLS=P1.
- The stepped
hole 80 communicates with apressure chamber 83 through theoutlet port 84 and aconvex groove 85 located in the outer peripheral surface thereof. Thus, when PLS>P1, the pressured oil with the maximum load pressure PLS is introduced into thepressure chamber 83, and when PLS<P1, the pressured oil with load pressure P1 is introduced into thepressure chamber 83. - As described above, the stepped
hole 80 and bore 81 have the function of sensing the higher oil pressure between the oil pressure PLS and P1, and of guiding it into thepressure chamber 83. Theshuttle valve 29 depicted in FIG. 2 is composed of the steppedhole 80 and thebore 81. The pressure of the pressured oil introduced into thepressure chamber 83 is pressure Pb depicted in FIG. 2. - The
pressure chamber 83 is a space enclosed by the inner surface of thesleeve 72, the right end surface of thepiston 73, and the outer peripheral surface of the slidingelement 74, where the right end surface of thepiston 73 functions as the secondpressure receiving component 22 depicted in FIG. 2. - The control
pressure producing component 7B is described below. The controlpressure producing component 7B is located to the side of thecompensator 7A, and is equipped with thevariable throttle valve 30 depicted in FIG. 2. - A
spool 88 for changing the flow resistance (throttle level) between theinlet port 32 andoutlet port 33 depicted in FIG. 2 is located in the vertical direction in thebody 87 of thevariable throttle valve 30. - The
spool 88 is such that downwardly directed force (the direction in which the flow resistance increases) is provided by thespring 31, and upwardly directed force (the direction in which the flow resistance decreases) is given by thespring 35 in apressure chamber 90 formed between the spool and an adjustingscrew 89. - The bottom end surface of the
spool 88 facing thepressure chamber 90 forms thepressure receiving component 34 depicted in FIG. 2. - The inner surface of a
concave component 91 located in the left surface of thebody 87 forms apressure chamber 92 along with the right end surface of the slidingelement 74 and the right end surface of thesleeve 72 of thecompensator 7A. The right end surface of the slidingelement 74 facing thepressure chamber 92 forms the secondpressure receiving component 23 of themain valve 20 depicted in FIG. 2. - The
inlet port 32 of thevariable throttle valve 30 communicates through theoil passage 37 equipped with thethrottle 36 to theoutlet port 62 of the operatingvalve 4, that is, to theinlet port 24 of themain valve 20 depicted in FIG. 2, and also communicates-through anoil passage 38 to thepressure chamber 92. Theoutlet port 33 communicates through theoil passage 40 equipped with the check valve 39 (see FIG. 2) to theactuator port 64 of the operatingvalve 4. - The pilot
pressure producing component 7C is located in the top of thebody 87 of the controlpressure producing component 7B. The electromagnetic proportionalpressure control valve 50 forming the pilotpressure producing component 7C comprises aspool 94 arranged in the vertical direction in thebody 93, and asolenoid 53 that presses thespool 94 down against thespring 54. - In this electromagnetic proportional
pressure control valve 50, thespool 94 is driven down by the thrust of thesolenoid 53, allowing the flow resistance to be reduced between theinlet port 52 and theoutlet port 55. - The
outlet port 55 communicates through anoil passage 95 to thepressure chamber 90 of thevariable throttle valve 30. Thespool 94 also is positioned on the axis of thespool 88 of the controlpressure producing component 7B. - The operation of the
pressure compensating valve 7 having the aforementioned structure is described below with reference to FIG. 4. - The pressured oil with the pressure Pa flowing out of the
outlet port 62 of the operatingvalve 4 presses thevalve component 66 to the right by acting on thesurfaces valve component 66 forming the firstpressure receiving component 21 of themain valve 20 depicted in FIG. 2. - Meanwhile, the pressured oil with the load pressure Pb (pressure P1 or PLS) flowing into the
pressure chamber 83 presses thevalve component 66 to the left by acting on the right end surface of the piston 73 (secondpressure receiving component 22 depicted in FIG. 2), and the pressured oil with the control pressure Pe flowing into thepressure chamber 92 presses thevalve component 66 to the left by acting on the right end surface of the sliding element 74 (thirdpressure receiving component 23 depicted in FIG. 2). Thespring 26 also presses thevalve component 66 to the left by means of the slidingelement 74. - The pressure balance in the
main valve 20 can thus be expressed as in the following Eq. (1). - Pa×A 0 =Pe×A 1 +Pb(A 0 −A 1)+F 0 (1)
- A0>A1
- A0: sum of the area of
surfaces valve component 66 - A1: area of right end surface of sliding
element 74 - A0−A1: area of right end surface of
piston 73 - F0: elastic force of
spring 26 - The pressure Pe in Eq. (1) is the control pressure that changes the pressure compensation characteristics of the
pressure compensating valve 7. The control pressure Pe results in pressure Pa when thevariable throttle valve 30 of controlpressure producing component 7B depicted in FIG. 2 is closed. The relation in Eq. (2) below is obtained by substituting Pa into Pe in Eq. (1). - Pa−Pb=F 0/(A 0 −A1) (2)
- As can be seen from this relation, the
pressure compensating valve 7 is operated in such a way that the pressure difference Pa−Pb is constant when Pe=Pa. In other words, pressure compensation is achieved. - Thus, operating both operating
valves 4 depicted in FIG. 1 to bring about the joint operation of thecylinders 5 avoids the drawback of pressured oil becoming concentrated and supplied to only thecylinder 5 with the lighter load. - When the
variable throttle valve 30 of the controlpressure producing component 7B is not closed, the pressured oil passing through the fixedthrottle 36 flows through thevariable throttle valve 30 andcheck valve 39 to thecylinder 5 end. Thus the control pressure Pe obtained by dividing the pressure difference between the pressures Pa and P1 by the throttling ratio between thethrottle 36 andvariable throttle valve 30, in other words, the control pressure Pe resulting from the reduction of the pressure Pa, acts on the thirdpressure receiving component 23 of themain valve 20 in thecompensator 7A. - In this state, the leftward moving force of the sliding
element 74 depicted in FIG. 4 is lower than when Pe=Pa. - Reducing the leftward moving force of the sliding
element 74 is equal to lowering the elastic force Fe of thespring 26 in the Eq. (2). That is because, when the control force Pe is lower than Pa, the pressure difference Pa−Pb is set lower than when Pe=Pa (change in the pressure compensation characteristics). Here, the function of keeping the pressure difference Pa−Pb constant is still maintained, despite the change in the pressure compensation characteristics. - In the case of two or more cylinders with different loads, more pressured oil flows to the one with the lower load under conditions where the control input of the operating
valves 4 is constant. - The control pressure Pe drops as the amount of electricity to the
solenoid 53 of the electromagnetic proportionalpressure control valve 50 increases. Accordingly, when an operating unit in construction machinery, for example (such as the boom, arm, or bucket in hydraulic shovels), is driven bycylinders 5, pressure compensation characteristics suitable for the operating configuration of such an operating unit can be set by controlling the amount of electricity to thesolenoid 53. - The pressure compensation characteristics of
pressure compensating valves 7 for a plurality of cylinders, as shown in FIG. 1, can also be altered, of course. The pressure compensation characteristics of thepressure compensating valves 7 for the series ofcylinders 5 depicted in FIG. 3 can each be varied so as to alter the operating speeds during extension and retraction of thecylinders 5. - In the
pressure compensating valves 7, the pilot pressure Pd no longer acts on thevariable throttle valve 39 of the controlpressure producing component 7B in the event of wire breakage in thesolenoid 53 of the electromagnetic proportionalpressure control valve 50 or in the event of malfunctions of thepilot pump 2 depicted in FIG. 1, for example. In other words, thevariable throttle 39 is no longer capable of throttling operations. - Despite such accidents, however, there is no loss of the pressure compensation characteristics of the
pressure compensating valves 7. Only a fully compensated state results. - That is, when the
variable throttle 39 is closed, the magnitude of the control pressure Pe is changed, making it impossible to change the pressure compensation characteristics. However, since the control pressure Pe is set to Pe=Pa, it is still possible to maintain pressure compensation operations keeping the pressure difference Pa Pb shown in the Eq. (2) at a constant level. - FIG. 5 depicts a second example of the structure of a
pressure compensating valve 7. Thispressure compensating valve 7 differs from thepressure compensating valve 7 in FIG. 4 in that thespring 54 of the electromagnetic proportionalpressure control valve 50 is brought into contact with the top end of thespool 88 of thevariable throttle valve 30 of the controlpressure producing component 7B. - In this
pressure compensating valve 7, when thespool 88 of thevariable throttle valve 30 is operated based on the pilot pressure Pd supplied from the electromagnetic proportionalpressure control valve 50, the operating force is mechanically fed back to thespool 94 of the electromagnetic proportionalpressure control valve 50 through thespring 54. - The operating characteristics (response) of the
spool 88 of thevariable throttle valve 30 are improved, allowing high-precision pressure compensation to be achieved. - FIG. 6 depicts a third example of the structure of the
pressure compensating valve 7. Thispressure compensating valve 7 is such that thevariable throttle valve 30 of the controlpressure producing component 7B and the electromagnetic proportionalpressure control valve 50 have a sharedbody 218, with thesolenoid 53 of the electromagnetic proportionalpressure control valve 50 located on the exterior of thebody 218. This allows the structure to be made more compact and the number of parts to be reduced. - Meanwhile, the control
pressure producing component 7B in thispressure compensating valve 7 forms aflange 88 a having a tapered peripheral surface on thespool 88 of thevariable throttle valve 30, and theflange 88 a is interposed between theinlet port 32 andoutlet port 33 of thevariable throttle valve 30. - When pressured oil with the pressure P1 flows through the
oil passage 40 into theoutlet port 33 of thevariable throttle valve 30 by means of this structure, the top surface of theflange 88 a is placed under pressure by the pressured oil. - The
spool 88 is thus moved down, and the tapered peripheral surface of theflange 88 a presses against the seat surface of thebody 218, so that theinlet port 32 andoutlet port 33 are blocked off from each other. - In this way, the
spool 88 functions as a check valve to prevent the pressured oil with the pressure P1 from flowing toward theinlet port 32. Thebody 218 of thispressure compensating valve 7 thus does not require thecheck valve 39 depicted in FIGS. 4 and 5, making thebody 218 easier to fabricate. - FIG. 7 depicts a fourth example of the structure of the
pressure compensating valve 7. Thispressure compensating valve 7 has a structure in which a joint 102 is attached to anattachment block 219 secured to the top surface of thebody 87 of the controlpressure producing component 7B, and thepressure chamber 90 of thevariable throttle valve 30 in the controlpressure producing component 7B communicates through the joint 102 and piping 95 to theoutlet port 55 of the electromagnetic proportionalpressure control valve 50. - In this
pressure compensating valve 7, the pilot pressure Pd output from the electromagnetic proportionalpressure control valve 50 or the pilot pressure output from a manual pilot valve can be allowed to act on thevariable throttle valve 30 of the controlpressure producing component 7B by way of the joint 102. Thispressure compensating valve 7 is thus suitable for use in cases where the electromagnetic proportionalpressure control valve 50 or pilot valve must be located at a distance from the controlpressure producing component 7B because of restricted space or the like. - The
variable throttle valve 30 of the controlpressure producing component 7B in thispressure compensating valve 7 has a structure similar to that of thevariable throttle valve 30 of thepressure compensating valve 7 depicted in FIG. 4. - FIG. 8 depicts a fifth example of the structure of the
pressure compensating valve 7. Thispressure compensating valve 7 has a structure in which a joint 104 is attached to the exterior of thebody 103 of the controlpressure producing component 7B, and thepressure chamber 90 located in thevariable throttle valve 30 of the controlpressure producing component 7B communicates through the joint 104 and piping 95 to the electromagnetic proportionalpressure control valve 50 or a manual pilot valve not shown in the figure. - The electromagnetic proportional
pressure control valve 50 or pilot valve of thispressure compensating valve 7 can be located apart from the controlpressure producing component 7B. Since the joint 104 is located in thebody 103 of the controlpressure producing component 7B in thispressure compensating valve 7, the machine can be made more compact and the number of parts can be reduced. - The
variable throttle valve 30 of the controlpressure producing component 7B has a structure similar to that of thevariable throttle valve 30 in thepressure compensating valve 7 depicted in FIG. 6. Thebody 103 of the controlpressure producing component 7B in thispressure compensating valve 7 thus requires no check valve in a manner similar to that in thepressure compensating valve 7 depicted in FIG. 6. - FIG. 9 depicts a sixth example of the structure of the
pressure compensating valve 7. Thispressure compensating valve 7 is composed of only thecompensator 7A and the controlpressure producing component 7B. Thecompensator 7A has a structure similar to that of thecompensator 7A depicted in FIG. 4. - The control
pressure producing component 7B is equipped with avariable throttle valve 30 having a structure allowing the magnitude of the throttling to be manually altered. Thisvariable throttle valve 30 has avertical hole 106 in thebody 105, and apoppet type spool 107 is inserted into thisvertical hole 106. The top and bottom of thevertical hole 106 can be rendered communicable and are blocked by the vertical movement of thespool 107. - The top of a
vertical hole 106 communicates through theoil passage 40 to theactuator port 64 of the operatingvalve 4. The bottom of thevertical hole 106 communicates through theoil passage 37 equipped with thethrottle 36 to theoutlet port 62 of the operatingvalve 4, and also communicates through theoil passage 38 to thepressure chamber 92. - An adjusting
screw 108 is threaded into the top of thevertical hole 106, and aspring 109 with weak elastic force is interposed between the adjustingscrew 108 and thespool 107. - In the
variable throttle valve 30 constructed in this manner, the pressured oil with the pressure Pa discharged from theoutlet port 62 of the operatingvalve 4 flows through theoil passage 37 into the bottom of thevertical hole 106. - With this, the
spool 107 is pushed up, and part of the pressured oil with the pressure Pa flows into theoil passage 40 while constricted by thespool 107. The pressure Pe of thepressure chamber 92 is set according to the amount of pressured oil flowing into theoil passage 40, that is, according to the throttle level of thespool 107. - The upward moving stroke of the
spool 107 defining the throttle level of thespool 107 can be adjusted by manually rotating the adjustingscrew 108. Thepressure compensating valve 7 can thus alter the pressure Pe, that is, can alter the pressure compensation characteristics, when thescrew 108 is rotated. - Since the
spool 107 is a poppet valve type, when pressured oil flows from thecylinder 5 into theoil passage 40, thespool 107 is pushed down, blocking off the top and bottom of thevertical hole 106 from each other. In other words, thespool 107 functions as a check valve. - Thus, with this
pressure compensating valve 7, there is no need to provide thebody 105 with thecheck valve 39 depicted in FIG. 4, making thebody 105 easier to fabricate. - FIG. 10 depicts a seventh example of the structure of the
pressure compensating valve 7. Thispressure compensating valve 7 differs from thepressure compensating valve 7 depicted in FIG. 4 in terms of the structure of thecompensator 7A. - That is, the
main valve 20 of thecompensator 7A depicted in FIG. 10 has a spool S comprising the unification of thevalve component 66 and pushingcomponent 67 depicted in FIG. 4. - In this
pressure compensating valve 7, the pressured oil with the maximum load pressure PLS flowing into theannular space 75 flows through ahole 112 located in thesleeve 72 directly into thepressure chamber 83, so the pressure Pb of thepressure chamber 83 results in the maximum load pressure PLS. - The spool S forms a
communication hole 113 along the central axis, thereby allowing theoutlet port 62 of the operatingvalve 4 and thepressure chamber 92 to communicate with each other. As a result, the pressured oil with the pressure Pa flowing from theoutlet port 62 of the operatingvalve 4 flows through the communicatinghole 113 into thepressure chamber 92. In other words, the communicatinghole 113 functions as theoil passage 37 in FIG. 4. - A fixed
throttle 113 a corresponding to the fixedthrottle 36 depicted in FIG. 4 is formed at the end on thepressure chamber 92 side of thecommunication hole 113. - In the
pressure compensating valve 7 having the aforementioned structure, there is no need to provide thebody 60 of thecompensator 7A with theoil passage 37 depicted in FIG. 4, nor is there-any need to provide thebody 87 of the controlpressure producing component 7B with thethrottle 36 depicted in FIG. 4. Thebodies - The
variable throttle valve 30 of the controlpressure producing component 7B has a structure similar to that of thevariable throttle valve 30 depicted in FIG. 4. - FIG. 11 depicts an eighth example of the structure of the
pressure compensating valve 7. The structure of thecompensator 7A in thispressure compensating valve 7 is similar to that of thepressure compensating valve 7 depicted in FIG. 10, and the structures of the controlpressure producing component 7B and pilotpressure producing component 7C are similar to those of thepressure compensating valve 7 depicted in FIG. 6. - Thus, in this
pressure compensating valve 7, the same effects in making thebody 60 and thebody 218 easier to fabricate can be obtained as in thepressure compensating valve 7 depicted in FIG. 10, and the same effects in making a more compact machine, reducing the number of parts, and making it easier to fabricate thebody 100 can be obtained as in thepressure compensating valve 7 depicted in FIG. 6. - FIG. 12 depicts a ninth example of the structure of the
pressure compensating valve 7. The structure of thecompensator 7A in thispressure compensating valve 7 is the same as that of thepressure compensating valve 7 depicted in FIG. 10, while the structure of the controlpressure producing component 7B and the location for attaching the joint 102 are the same as that of thepressure compensating valve 7 depicted in FIG. 7. - In this
pressure compensating valve 7, the same effects in making thebodies pressure compensating valve 7 depicted in FIG. 10, and the same effects in locating the electromagnetic proportionalpressure control valve 50 apart from the controlpressure producing component 7B can be obtained as in thepressure compensating valve 7 depicted in FIG. 7. - FIG. 13 depicts a tenth example of the structure of the
pressure compensating valve 7. The structure of thecompensator 7A of thispressure compensating valve 7 is similar to that of thepressure compensating valve 7 depicted in FIG. 10, and the structure of the controlpressure producing component 7B and the position for attaching the joint 140 are the same as in thepressure compensating valve 7 depicted in FIG. 8. - In this
pressure compensating valve 7, the same effects in making it easier to fabricate thebodies pressure compensating valve 7 depicted in FIG. 10, and the same effects in locating the electromagnetic proportionalpressure control valve 50 apart from the controlpressure producing component 7B can be obtained as in thepressure compensating valve 7 depicted in FIG. 8. - The
variable throttle valve 30 of the controlpressure producing component 7B has a structure similar to that of thevariable throttle valve 30 in thepressure compensating valve 7 depicted in FIG. 6. The same effects in dispensing with the need to provide thebody 103 of the controlpressure producing component 7B with a check valve can be obtained as in thepressure compensating valve 7 depicted in FIG. 6. - FIG. 14 depicts an eleventh example of the structure of the
pressure compensating valve 7. The structure of thecompensator 7A of thispressure compensating valve 7 is similar to that of thepressure compensating valve 7 depicted in FIG. 10, and the structure of the controlpressure producing component 7B is similar to that of thepressure compensating valve 7 depicted in FIG. 9. - In this
pressure compensating valve 7, the same effects in making thebodies pressure compensating valve 7 depicted in FIG. 10. The same effects in being able to manually adjust the throttle level and making thebody 105 easier to fabricate can be obtained as in thepressure compensating valve 7 depicted in FIG. 9. - FIG. 15 depicts a twelfth example of the structure of the
pressure compensating valve 7. The structure of thecompensator 7A of thispressure compensating valve 7 differs from that of thepressure compensating valve 7A depicted in FIG. 4. - The spool S of the
main valve 20 of thecompensator 7A depicted in FIG. 15 is equipped with apiston 116 featuring the unification of thevalve component 66 and thepiston 73 depicted in FIG. 4, and a slidingelement 117 located in thepiston 116. - The
piston 116 and the slidingelement 117 are located along the central axis through the communication holes 118 and 119, respectively. One end of thecommunication hole 119 in the slidingelement 117 communicates through acheck valve 120 to thecommunication hole 118 of thepiston 116, and the other end communicates through athrottle 119 a corresponding to thethrottle 36 depicted in FIG. 2 to thepressure chamber 92. - In the
pressure compensating valve 7 with the aforementioned structure, the pressured oil with the pressure Po supplied from theoutlet port 62 flows into thepressure chamber 92 through thecommunication hole 118, acheck valve 120, aslit 121 formed around thecheck valve 120, aport 122 passing through the peripheral wall of the slidingelement 117, the communicatinghole 119, and thethrottle 119 a. In other words, the communication holes 118 and 119 function as theoil passage 37 depicted in FIG. 2. - Accordingly, there is no need to provide the
body 60 of thecompensator 7A with theoil passage 37 depicted in FIG. 4, and there is no need to provide thebody 87 of the controlpressure producing component 7B with thethrottle 36 depicted in FIG. 4. It is thus easier to fabricate thebodies - Meanwhile, when the pressure P1 of the pressured oil in the
actuator port 64 becomes greater than the pressure Po of the oil pressure in theoutlet port 62, thecheck valve 120 closes. The pressured oil in theactuator port 64 is thus prevented by thecheck valve 120 from flowing into theoutlet port 62. - The
check valve 120 thus has the same function as thecheck valve 39 depicted in FIG. 2. Accordingly, in thispressure compensating valve 7, there is no need to provide thebody 87 of the controlpressure producing component 7B with thecheck valve 39 depicted in FIG. 4, which makes thebody 87 easier to fabricate. - In the
pressure compensating valves 7 described above, theoil passage 40 connected to theoutlet port 33 of thevariable throttle valve 30 was connected to the actuator port 64 (oil passage 6 a) of the operatingvalve 4 depicted in FIG. 3, but thisoil passage 40 may also be connected to thetank port 65. - The structure of the unloading
pressure control valve 10 relating to the present invention is described below with reference to FIG. 16. - FIG. 16 is a circuit diagram of oil pressure, depicting the structure of the unloading
pressure control valve 10. The unloadingpressure control valve 10 is used to return the oil discharged from ahydraulic pump 1 directly to a tank to keep thehydraulic pump 1 in an unloaded state in a hydraulic system comprising, for example, avariable delivery pump 1, an auxiliary hydraulic pump (pilot hydraulic pump) 2, an operatingvalve 4 to which the oil discharged from thehydraulic pump 1 is supplied through anoil passage 3, and a hydraulic cylinder (hydraulic actuator) 5 located opposite the operatingvalve 4. - The unloading
pressure control valve 10 comprises amain valve 100 and an electromagnetic proportionalpressure control valve 101. - The
main valve 100 has a firstpressure receiving component 123, a secondpressure receiving component 124, a thirdpressure receiving component 125, and a fourthpressure receiving component 126. Themain valve 100 sets the throttle level (unloading start pressure) between afirst inlet port 127 andoutlet port 128 by means of the elastic force of aspring 130 and the pressure acting on the firstpressure receiving component 123, secondpressure receiving component 124, thirdpressure receiving component 125, and fourthpressure receiving component 126. - The first
pressure receiving component 123 is connected to thevariable delivery pump 1 along with thefirst inlet port 127, and receives the discharge pressure PP of thehydraulic pump 1. The secondpressure receiving component 124 receives the maximum load pressure PLS by way of athrottle 129. The thirdpressure receiving component 125 receives the control pressure Pg described below. The fourthpressure receiving component 126 is connected to the tank. Themain valve 100 determines the unloading set pressure by means of the elastic force of thespring 130 and the pressure area of the secondpressure receiving component 124 and third pressure receiving component. Themain valve 100 does not require thespring 130. In other words, the unloading start pressure can be set by just the difference between the pressure area of the secondpressure receiving component 124 and the third pressure receiving-component. - The control pressure Pg is given from the electromagnetic proportional
pressure control valve 101. That is, the electromagnetic proportionalpressure control valve 101 introduces the pressured oil discharged from an auxiliaryhydraulic pump 2 through theinlet port 132, and the oil pressure resulting from a reduction in the pressure Pc of this pressured oil is output as the control pressure Pg. The control pressure Pg changes proportionally to the amount of electricity sent to thesolenoid 133. - When zero electricity is supplied to the
solenoid 133, theoutlet port 135 communicates with thetank port 136 by means of the elastic force of aspring 134, as shown in the figure. The control pressure Pg acting on the thirdpressure receiving component 125 of themain valve 100 is thus zero. - The specific structure of the unloading
pressure control valve 10 is described below with reference to FIG. 17. - A sliding
element 145 is slidably inserted into the left side of thevalve body 140 of themain valve 100, and the left end of asleeve 148 is fitted to the right side of thevalve body 140. - The sliding
element 145 has a U-shaped cross section, and is brought into contact on the left end surface with an adjustingscrew 147 threaded into the left end of thevalve body 140. The adjustingscrew 147 is locked by alock nut 148. The interior of the slidingelement 145 communicates through ahole 145 a to the tank. - A
spool 150 has a firstsmall diameter component 151 forming a left half, alarge diameter component 152 forming a central component, and a secondsmall diameter component 153 forming a right half. The left tip of the firstsmall diameter component 151 of thespool 150 is slidably inserted into the slidingelement 145. Thelarge diameter component 152 is slidably inserted into alarge diameter hole 154 in asleeve 146. The secondsmall diameter component 153 is slidably inserted into asmall diameter hole 155 in thesleeve 146. - The
right end surface 150 a of thespool 150 forms the firstpressure receiving component 123 depicted in FIG. 16. Theleft end surface 150 b of thespool 150 forms the fourthpressure receiving component 126. - The
spool 150 is designed so that the cross sectional area of the secondsmall diameter component 153 is a size equal to that obtained by subtracting the cross sectional area of the firstsmall diameter component 151 from the cross sectional area of thelarge diameter component 152. - The right end of the
sleeve 146 is positioned in thevalve body 180 of the operatingvalve 4. Thesleeve 146 forms thefirst inlet port 127 depicted in FIG. 16 by opening the right end. Theinlet port 127 communicates with thepump port 181 of the operatingvalve 4. - Meanwhile, the
sleeve 146 forms theoutlet port 128 depicted in FIG. 16 at a position located slightly to the left of the right end opening. Theoutlet port 128 communicates with thetank port 182 of the operatingvalve 4. - The
sleeve 146 further comprises a loadpressure introduction port 157 and a controlpressure introduction port 158. The loadpressure introduction port 157 introduces pressured oil with the maximum load pressure PLS The controlpressure introduction port 158 introduces control pressure Pg through the electromagnetic proportionalpressure control valve 101. - The load
pressure introduction port 157 communicates through anannular space 159, anoil hole 160, and afine hole 161 to aspring chamber 162. Theannular space 159 is formed between the inner peripheral surface of thesleeve 146 and the outer peripheral surface of the secondsmall diameter component 153 of thespool 150. Theoil hole 60 is formed along the central axis of thespool 150. Thefine hole 161 passes diametrically through thespool 150, forming thethrottle 129 depicted in FIG. 16. - Meanwhile, the control
pressure introduction port 158 communicates with aspace 163 formed between thelarge diameter component 152 of thespool 150 and thesleeve 146. Theright end surface 152 a of the spoollarge diameter component 152 located in thespace 163 forms the thirdpressure receiving component 125 depicted in FIG. 16. - The
spring 130 depicted in FIG. 16 is located in thespring chamber 162. Thespring 130 is interposed between aspring receiver 162 a inserted into the firstsmall diameter component 151 of thespool 150 and the right end surface of the slidingelement 145, and pushes thespool 150 to the right. - While the
spring receiver 162 a is in contact with the left end of thesleeve 146 in the state depicted in the figure, thefirst inlet port 127 andoutlet port 128 are blocked off from each other by the right end of thespool 150. Theleft end surface 152 b of the spoollarge diameter component 152 facing thespring chamber 162 forms the elastic force creating component of thespring 130 as well as the secondpressure receiving component 124 depicted in FIG. 16. - The electromagnetic proportional
pressure control valve 101 of the unloadingpressure control valve 10 is described below. - The electromagnetic proportional
pressure control valve 101 is disposed over thevalve body 140 of themain valve 100. Aspool 167 for allowing theinlet port 132 andoutlet port 135 depicted in FIG. 16 to communicate with each other and to be blocked off from each other is located in thevalve body 166 of the electromagnetic proportionalpressure control valve 101. The top of thevalve body 166 has asolenoid 133 that pushes thespool 167 down against thespring 134. - The
inlet port 132 is connected to the auxiliaryhydraulic pump 2. Theoutlet port 135 communicates through anoil passage 168 to the controlpressure introduction port 158. - The operation of the unloading
pressure control valve 10 having the aforementioned structure is described below. - When the discharge pressure PP of the
hydraulic pump 1 acts on theright end surface 150 a of thespool 150 which is the firstpressure receiving component 123, thespool 150 is pushed to the left (the direction passing through thefirst inlet port 127 and outlet port 128). - Meanwhile, the control pressure Pg supplied from the electromagnetic proportional
pressure control valve 101 acts on theright end surface 152 a of the large diameter component of thespool 150 serving as the thirdpressure receiving component 125, by way of theoil passage 168 and the controlpressure introduction port 158, so that thespool 150 is pushed to the left. - The
spring 130 located in thespring chamber 162 pushes thespool 150 to the right. The load pressure PLS is introduced through the loadpressure introduction port 157,annular space 159,oil hole 160, and fine hole 161 (throttle 129) into thespring chamber 162. The load pressure PLS thus acts on theleft end surface 152 b of the large diameter component of thespool 150 which is the secondpressure receiving component 124, and thespool 150 is pushed to the right. - The balance of force determining the position of the
spool 150 in the unloadingpressure control valve 10 is represented by the following Eq. (3). - P P×A1 =P LS×(A 2 −A 1)+F 0 −Pg×(A 2−A1) (3)
- Where
- A1: area of
right end surface 150 a ofspool 150 - A2: area of
large diameter component 152 ofspool 150 - A3: area of
left end surface 150 b ofspool 150 - F0: elastic force of
spring 130 - As noted above, the relation between area A1, A2, and A3 is A1=(A2−A3). Eq. (3) thus results in Eq. (4) below.
- (P2−P LS)×A 1 =F 0 −Pg×(A 2 −A 1) (4)
- It is evident from the Eq. (4) that a constant pressure difference PP−PLS is obtained irrespective of fluctuations in the load pressure PLS when the control pressure Pg is constant.
- The pressure difference PP−PLS determines the unloading start pressure. The unloading
pressure control valve 10 thus allows the unloading start pressure to be arbitrarily set by controlling the amount of electricity to thesolenoid 133 of the electromagnetic proportionalpressure control valve 101 to change the control pressure Pg. - The
main valve 100 of the unloadingpressure control valve 10 is interposed between thehydraulic pump 1 and the tank. Thus, when the pressure difference PP−PLS reaches the unloading start pressure, the oil discharged from thehydraulic pump 1 is returned to the tank during continuous operation. - When the operating
valves 4 are operated in the center valve position, the pressure difference PP−PLS increases to the unloading start pressure. With this, the oil discharged from thehydraulic pump 1 is returned through the unloadingpressure control valve 10 to the tank, so thehydraulic pump 1 is in an unloaded state. - The electromagnetic proportional
pressure control valve 101 of the unloadingpressure control valve 10 produces pilot control pressure Pg resulting from the reduction of the discharge oil pressure Pc of the auxiliaryhydraulic pump 2. Meanwhile, in themain valve 100, the operating start pressure (unloading start pressure) changes according to the control pressure Pg given by the electromagnetic proportionalpressure control valve 101. - Thus, according to the unloading
pressure control valve 10, control signals to thesolenoid 133 of the electromagnetic proportionalpressure control valve 101 can be changed to set the unloading start pressure to the desired magnitude. - FIG. 18 depicts another embodiment of the unloading pressure control valve relating to the present invention.
- This unloading
pressure control valve 10 comprises anattachment block 185, a piping joint 187, and an oilpressure pilot valve 188. Theattachment block 185 is fixed to the upper surface of thevalve body 140. The piping joint 187 is screwed into a threadedhole 186 located in theattachment block 185, and is thus secured. The oilpressure pilot valve 188 is manually operated. - The threaded
hole 186 passes through the controlpressure introduction port 158. Theinlet port 188 b of the oilpressure pilot valve 188 is connected to the auxiliaryhydraulic pump 2. Theoutlet port 188 a is connected to the piping joint 187. - In this unloading
pressure control valve 10, the pressured oil with the control pressure Pg supplied from the oilpressure pilot valve 188 acts on theright end surface 152 a (third pressure receiving component 125) of the spoollarge diameter component 152 by way of the controlpressure introduction port 158. - This unloading
pressure control valve 10 allows the unloading start pressure to be arbitrarily set according to the control pressure Pg. The oilpressure pilot valve 188 which is the means for producing the control pressure Pg can also be disposed apart from themain valve 100. It can thus be freely disposed, enabling manual remote control of the unloading start pressure, and the like. - In the unloading
pressure control valves 10 depicted in FIGS. 17 and 18, the control pressure Pg acted as the force moving thespool 150 to the left (the direction passing through thefirst inlet port 127 andoutlet port 128 of the main valve 100). - In contrast to the above, it is also possible to allow the control pressure Pg to act as the force moving the
spool 150 to the right. In this case, the pressing force of thespring 130 acts in the direction opposite that described above (the direction in which the spool is pushed to the left). - When the control pressure Pg is allowed to act in the opposite direction as described above, the unloading start pressure increases as the control pressure Pg increases.
- FIG. 19 depicts a hydraulic system featuring the use of two
hydraulic pumps - In this hydraulic system, the
hydraulic pumps corresponding operating valves valve 191 in a convergedflow component 190. A switchingvalve 192 switches between the communication and blockage of pressured oil, with a maximum load pressure PLS-A sensed by oneshuttle valve 8A, and pressured oil with a maximum load pressure PLS-B sensed by anothershuttle valve 8B. - The switching
valves flow component 190 are always simultaneously switched over by means of the pilot pressure Ph. - In the state depicted in the figure, the switching
valves flow component 190 in this case allow the oil discharged by thehydraulic pumps - Meanwhile, when the switching
valves flow component 190 are switched over by the pilot pressure Ph, the oil discharged by thehydraulic pumps pressure control valves - The maximum load pressure PLS-A sensed by the
shuttle valve 8A is the highest among the plurality ofhydraulic cylinders 5 driven by thehydraulic pump 1A. The maximum load pressure PLS-B sensed by theshuttle valve 8B is the highest among the plurality ofhydraulic cylinders 5 driven by thehydraulic pump 1B. - The load pressure PLS-A is supplied to the unloading
pressure control valve 10 and the volume control component (pump discharge pressure control means) 12 of thehydraulic pump 1A. The load pressure PLS-A is also supplied through acheck valve 193A to the load pressure bleedvalve 11. - The load pressure PLS-B is supplied to the unloading
pressure control valve 10 and thevolume control component 12 of thehydraulic pump 1B. The load pressure PLS-B is also supplied through acheck valve 193B to the load pressure bleedvalve 11. - As described above, when the two pump circuits are separated, the switching
valves flow component 190 are switched to a blocking state. - However, even though the switching
valves operating valve 4A is in the center valve state, and theother operating valve 4B is in the operating state, the maximum load pressure PLS-A sensed by theshuttle valve 8A should be zero as long the switchingvalve 192 is operating in an ideal manner. In fact, however, the oil leakage from the switchingvalve 192 results in an increase in the maximum load pressure PLS-A. - In this case, when the maximum load pressure PLS-A increases, the discharge pressure PP of the
hydraulic pump 1A also increases, resulting in the maximum load pressure PLS-A+pump set pressure. - The pressured oil with the maximum load pressure PLS-A is allowed to communicate with the tank during the operation of the one unloading
pressure control valve 10A. That is, the pressured oil with the maximum load pressure PLS-A is introduced from in front of thethrottle 129 through the branched piping into the unloadingpressure control valve 10A, and this pressured oil is also output through athrottle 169 from the unloadingpressure control valve 10A so as to be returned to the tank. This allows the maximum load pressure PLS confined in the piping leading from theshuttle valve 8A to themain valve 20 to escape to the tank, and prevents the discharge pressure PP of thehydraulic pump 1A from increasing. The discharge pressure PP of thehydraulic pump 1B can similarly be prevented from increasing during the operation of the other unloadingpressure control valve 10B. - The structure of the
variable bleed valve 11 relating to the present invention is described below with reference to FIG. 20. - The
variable bleed valve 11 comprises avariable throttle valve 110 and an electromagnetic proportionalpressure control valve 111, as shown in the enlargement in FIG. 20. - The
variable throttle valve 110 is operated so as to increase the area of the opening-between aninlet port 196 and anoutlet port 197 by means of the elastic force of aspring 95 and the pilot pressure Pg acting on apressure receiving component 194, and is operated so as to reduce the area of the opening by means of the elastic force of aspring 198. - The electromagnetic proportional
pressure control valve 111 introduces pressured oil with a standard pressure Pc discharged from the auxiliaryhydraulic pump 2 into theinlet port 199, and the pressure Pc of the pressured oil is reduced to the pilot pressure Pg. The pressured oil with the pilot pressure Pg is allowed to act on thepressure receiving component 194 of thevariable throttle valve 110 by way of theoutlet port 200. The pilot pressure Pg changes proportionally to the amount of electricity to thesolenoid 201. - The
variable bleed valve 11 is connected to acontroller 300. Thecontroller 300 gives a corresponding control signal to thesolenoid 201 of the electromagnetic proportionalpressure control valve 111 based on operation commands such as a command to open the operatingvalve 4 by the operation of an operating lever (not shown in figure). - FIG. 21 depicts the
variable bleed valve 11 while mounted. It may also be seen from FIG. 21 that thevariable bleed valve 11 is provided as a valve block along with a plurality ofoperating valves variable bleed valve 11 is attached by means of asupport block 202 to the operatingvalve 4 located on the outermost side of the plurality ofoperating valves 4 joined in parallel. Thesymbol 4 a indicates the spool of the operatingvalve 4. - FIG. 22 is a cross section of line A-A in FIG. 21. It may be seen from FIG. 22 that the
variable throttle valve 110 is such that thespool 206 is inserted into thespool hole 205 of thevalve body 204. Thespool hole 205 is formed in the vertical direction. - The
spool 206 is interposed between theinlet port 196 andoutlet port 197 of thevariable throttle valve 110. Thespool 206 is such that downward force (the direction in which the area of the opening between theports spring 215. Meanwhile, the upward force (the direction in which the area of the opening between theports spring 195 in thepressure chamber 209 formed between the spool and anadjustment screw 62. - The bottom end surface of the
spool 205 facing thepressure chamber 209 forms thepressure receiving component 194 depicted in FIG. 20. The elastic force of thespring 195 can be fine tuned by the operation of theadjustment screw 217. - The
inlet port 196 communicates with the loadpressure introduction hole 203 through a load pressureintroduction oil passage 210 leading from thevalve body 204 to thesupport block 202. Theoutlet port 197 communicates with the tank through atank oil hole 211 that opens into theattachment surface 204 of thevalve body 204. - The electromagnetic proportional
pressure control valve 111 is disposed on the upper surface of theaforementioned valve body 204. The electromagnetic proportionalpressure control valve 111 comprises a spool 214 and thesolenoid 201. The spool 214 is vertically disposed in thevalve body 213. The spool 214 is disposed coaxially relative to thespool 206 of thevariable throttle valve 110. Thesolenoid 201 pushes the spool 214 down against thespring 215 according to the amount of electricity. - The spool214 is constantly pushed upward by the elasticity of the
spring 215. In this state, theinlet port 199 andoutlet port 200 of the electromagnetic proportionalpressure control valve 111 are blocked off from each other. The standard pressure Pc discharged by the auxiliaryhydraulic pump 2 acts on theinlet port 199. - The
outlet port 200 communicates with thepressure chamber 209 by way of anoil passage 216 located in thevalve body 213 and anoil passage 212 located in thevalve body 204 of thevariable throttle valve 110. Thespring 215 is in contact with the upper tip of thespool 206 of thevariable throttle valve 110. Here, thespring chamber 207 of thevalve body 204 communicates with the tank by way of atank oil hole 208 that opens into theattachment surface 204 a of thevalve body 204. - The operation of the
variable bleed valve 11 is described below. - Pressured oil present in the
oil passage 9 gradually flows through the fixedthrottle 112 into the tank when the operatingvalves 4 are operated in the center valve position. The maximum load pressure PLS acting on the discharge pressure control means 12 thus gradually decreases. When the maximum load pressure PLS decreases to zero, the displacement volume of thehydraulic pump 1 is reduced to the minimum preset volume by the pump discharge pressure control means 12. - When the operating
valve 4 is operated from this state to supply pressured oil to thehydraulic cylinder 5, the maximum load pressure PLS increases. The pressured oil of the maximum load pressure PLS is introduced into theinlet port 196 of thevariable throttle valve 110 through the loadpressure introduction hole 203 connected to theoil passage 9 and the load pressureintroduction oil passage 210. When theinlet port 196 andoutlet port 197 of thevariable throttle valve 110 thus communicate with each other, part of the pressured oil with the maximum load pressure PLS is bled off into the tank through theoutlet port 197. - The amount of the aforementioned pressured oil bled off at this time increases as the area of the opening between the
inlet port 196 andoutlet port 197 increases. The greater the amount that is bled off, the lower the rate of increase of the maximum load pressure PLS. - When the
solenoid 201 of the electromagnetic proportionalpressure control valve 111 is in a noncommunicating state, the spool 214 remains pushed up by thespring 215. Theinlet port 199 andoutlet port 200 of the electromagnetic proportionalpressure control valve 111 are thus blocked off from each other. - In this state, the pilot pressure Pg given from the
outlet port 200 of the electromagnetic proportionalpressure control valve 111 to thepressure chamber 209 of thevariable throttle valve 110 is zero. Thespool 206 of thevariable throttle valve 110 is thus pushed down by thespring 198. - When the
spool 206 is pushed down, theinlet port 196 andoutlet port 197 of thevariable throttle valve 110 are blocked off from each other by thespool 206. The area of the opening between theports variable throttle valve 110 is zero. - When electricity is applied to the
solenoid 201 of the electromagnetic proportionalpressure control valve 111, the spool 214 of the electromagnetic proportionalpressure control valve 111 is pressed down by the thrust of thesolenoid 201, allowing theinlet port 199 andoutlet port 200 to communicate with each other. - With this, the pilot pressure Pg resulting from a reduction in the discharge pressure Pc of the auxiliary
hydraulic pump 2 acts on thepressure chamber 209 of thevariable throttle 20. Thespool 206 of thevariable throttle valve 110 thus moves up against thespring 198. - When the
spool 206 moves up, theinlet port 196 andoutlet port 197 of thevariable throttle valve 110 communicate with each other. As a result, the pressured oil with the maximum load pressure PLS introduced into theinlet port 196 is bled off through theoutlet port 197 into the tank. - The greater the pilot pressure Pg at this time, in other words, the greater the amount of electricity to the
solenoid 201 of the electromagnetic proportionalpressure control valve 111, the greater the amount of pressure bled off by thevariable throttle valve 110. - As is evident from the description above, the
variable bleed valve 11 allows the amount of pressured oil with the maximum load pressure PLS that is bled off to be arbitrarily adjusted by controlling the amount of electricity to the electromagnetic proportionalpressure control valve 111. In other words, the rate of increase in the maximum load pressure PLS in theoil passage 9 can be arbitrarily adjusted by controlling the aforementioned amount of electricity. - Adjusting the amount of pressured oil that is bled off by the
variable throttle valve 110 to zero results in a higher rate of increase in the maximum load pressure PLS acting on the pump discharge pressure control means 12, so the pump discharge pressure control means 12 rapidly increases the displacement volume (discharge oil amount) of thehydraulic pump 1. - As a result, the
hydraulic cylinder 5 starts rapidly at the same time the operatingvalve 4 is operated. - In contrast, when the
variable throttle valve 110 is in bleed off operating mode, the rate of increase in the maximum load pressure PLS acting on the pump discharge pressure control means 12 is lower than when the aforementioned bleed off amount is zero. In this case, the pump discharge pressure control means 12 moderately increases the displacement volume of thehydraulic pump 1, so the start up speed of thehydraulic cylinder 5 decreases. - Accordingly, the
variable bleed valve 11 allows the start up response of thehydraulic cylinder 5 to be adjusted by controlling the amount of electricity to thesolenoid 201 of the electromagnetic proportionalpressure control valve 111. - The amount discharged by the
hydraulic pump 1 is controlled to bleed off the pressured oil in theoil passage 9 for sensing the maximum load pressure PLS serving as the pilot pressure. The amount flowing in the loadpressure sensing channel 9 is generally quite low. The pump pressure is controlled according to the pressure of the loadpressure sensing passage 9, whereas the pressure of the loadpressure sensing passage 9 is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from thehydraulic pump 1 can be controlled with greater precision. - In the aforementioned hydraulically operated device, only bleed off operations actually stop in the event of accidents such as malfunctions of the electromagnetic proportional
pressure control valve 111 which lead to interruption of the pilot pressure Pg. In other words, the operation of thehydraulic cylinder 5 by thehydraulic pump 1 is unaffected even when accidents such as those described above occur. The reliability of the hydraulically operated device can thus be improved. - Moreover, the amount of pressured oil that is bled off can be arbitrarily adjusted by means of control signals output by a
controller 300 described below, making such control easier to manage. Since pressured oil should be supplied by the application of electricity to the electromagnetic proportionalpressure control valve 111 only when bleed off is needed, not only can pressured oil energy loss be further minimized, but electrical energy can also be economized. - Here, the aforementioned hydraulically operated device is equipped with a
controller 300 connected to thevariable bleed valve 11 as described above, and thiscontroller 300 comprises, as shown in FIG. 20, a modesetting memory component 310, a modeselect setting component 320, and a controlsignal output component 330. - The mode
setting memory component 310 sets and stores a plurality of input-output relations according to the operating configuration of thehydraulic cylinder 5. As shown in FIG. 23, for example, three different modes comprising an ordinary mode which is the ordinary operating state, a heavy operating mode requiring considerable force, and a more precise operating mode requiring highly precise manipulations are set and stored in terms of the input-output relations between the open command to the operatingvalve 4 and the control signals to thesolenoid 201 of the electromagnetic proportionalpressure control valve 111, that is, the area of the opening of thevariable throttle valve 110. Although these three input-output relations have the same degree of variation relative to each other, the area of the opening of thevariable throttle valve 110 in terms of open commands to thesame operating valve 4 is preset and stored so as to increase in ascending order from heavy operating mode, to ordinary mode, to precision operating mode. - The mode
select setting component 320 selects and sets one of the three input-output relations set and stored in the modesetting memory component 310. This modeselect setting component 320 selects and sets a corresponding input-output relation according to the operation of a mode select switch not shown in the figure and located in the driver seat of a hydraulic shove, for example. - The control
signal output component 330 converts the open command for the operatingvalve 4 based on the input-output relation selected by the modeselect setting component 320, and the converted control signal is given to thesolenoid 201 of the electromagnetic proportionalpressure control valve 111. - Thus, in the aforementioned hydraulically operated device, a control signal output from the
controller 300 in response to an open command for the operatingvalve 4 can be modified according to the operating configuration of thehydraulic cylinder 5. In other words, when heavy operating mode is selected and set by the modeselect setting component 320, the area of the opening of thevariable throttle valve 110 for open commands to the operatingvalve 4 can be further reduced. Thus, in this heavy operating mode, more pressured oil can be supplied to thehydraulic cylinder 5 and the hydraulic cylinder can be rapidly operated, even though the control input of the operating lever (not shown in figure) is the same. - Meanwhile, when precision operating mode is selected and set, the area of the opening in the
variable throttle valve 110 can be further increased for open commands to the operatingvalve 4. Thus, in precision operating mode, less pressured oil can be supplied to thehydraulic cylinder 5 and the hydraulic cylinder can be moderately operated, even though the control input of the operating lever (not shown in figure) is the same. - The
hydraulic cylinder 5 is provided to drive the operating unit of the hydraulic shovel (such as a boom, arm, or bucket). A hydraulically operated device equipped with thevariable bleed valve 11 can thus provide operating speeds and operating sensitivity for an operating unit that are suitable for the operating configuration of the aforementioned hydraulic shovel. - The plurality of input-output relations set and stored in the mode
setting memory component 310 are not limited to those depicted in FIG. 5. - FIG. 24 is a graph depicting another example of input-output relations set and stored by the mode
setting memory component 310. The input-output relations depicted in FIG. 24 are designed so that the rate of change increases in the order from the heavy operating mode, to ordinary mode, to precision operating mode. The use of this mode setting selection means results in a different proportion of change in the speed by which the pressured oil in the loadpressure sensing passage 9 is bled off into the tank, allowing the operating speeds and operating sensitivity of thehydraulic cylinder 5 to be set with even greater precision according to the operating configuration. - A combination of the input-output relations depicted in FIGS. 23 and 24 can provide input-output relations such as that indicated by the broken lines in FIG. 23 for the heavy operating mode and precision operating mode in relation to ordinary mode. In this case, the input-output relations can be set even more precisely than those depicted in FIG. 24.
- The
variable throttle valve 110 is constructed in such a way as to increase the area of the opening between theinlet port 196 andoutlet port 197 by means of the action of the pilot pressure Pg, but conversely it can also be constructed in such a way as to reduce the area of the aforementioned opening by means of the action of the pilot pressure Pg. - The
variable throttle valve 110 is also constructed in such a way that thespool 206 is pressed in the cut-off direction (downward in FIG. 22) by thespring 198 and thespool 206 is pressed in the communicating direction (upward in FIG. 22) by the pressure in thepressure chamber 209, but it can also be constructed in such a way that the elastic force of thespring 198 and the pressure of thepressure chamber 209 act in directions opposite those described above. - The
variable bleed valve 11 is such that thespool 206 of thevariable throttle valve 110 and the spool 214 of the electromagnetic proportionalpressure control valve 111 are located coaxially, making it possible to achieve more compact shapes with a shorter lateral length. That is, when the lay out of thevariable bleed valve 11, for example, is like that depicted in FIG. 21, a more compact embodiment can be devised because the electromagnetic proportionalpressure control valve 111 can be mounted further inside than thespring case 4 b of the operatingvalve 4, that is, inside the surface defined by thespring case 4 b when a valve block is used in a generally right-angled parallelepiped form. - The
variable bleed valve 11 is such that thespring 215 of the electromagnetic proportionalpressure control valve 111 is in contact with the upper end of thespool 206 of thevariable throttle valve 110. According to this structure, the operating force of thespool 206 is mechanically fed back to the spool 214 of the electromagnetic proportionalpressure control valve 111 through thespring 215 when thespool 206 of thevariable throttle valve 110 is operated. The operating characteristics (response) of thespool 206 of thevariable throttle valve 110 can thus be improved, allowing more precise bleed off operations to be managed. - The
variable bleed valve 11 is also designed to allow the elastic force of thespring 195 of thevariable throttle valve 110 to be fine tuned by rotating theadjustment screw 217. When a plurality ofvariable bleed valves 11 are manufactured, the machining precision of the various parts and the elastic force of thespring 198 used in the individualvariable bleed valves 11 are not uniform. Despite the uneven elastic force of thespring 198, however, it is possible to compensate for the uneven elastic force of thespring 198 by adjusting the elastic force of thespring 195 by rotating theadjustment screw 217.
Claims (7)
1. A pressure compensating valve through which pressured oil fed from a hydraulic pump (1) to a hydraulic actuator (5) passes, comprising:
a main valve (20), that operates in such a way as to increase an area of an opening between an inlet port (24) and an outlet port (25) thereof by means of pressure acting on a first pressure receiving component (21), and operates in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component (22) and pressure acting on a third pressure receiving component (23), and that allows pressure (Pa) of the pressured oil flowing into the inlet port (24) to act on the first pressure receiving component (21) and pressure (Pb) of a load driven by the pressured oil flowing out from the outlet port (25) to act on the second pressure receiving component (22); and
control pressure producing means (7B) for allowing control pressure (Pe) resulting from a reduction in the pressure (Pa) of the inlet port (24) to act on the third pressure receiving component (23).
2. An unloading pressure control valve for introducing a pressured oil discharged from a hydraulic pump (1) to a tank according to pressure difference between discharge pressure (PP) of the hydraulic pump (1) and load pressure (PLS) of a hydraulic actuator (5), comprising:
a main valve (100), that operates toward a communicating direction by means of discharge pressure (PP) of the hydraulic pump (1) acting on a first pressure receiving component (23), to operate toward a cut-off direction by means of load pressure (PLS) acting on a second pressure receiving component (24), and that changes balance of an operating force in each of the directions by means of control pressure (Pg) acting on a third pressure receiving component (25); and
control pressure producing means (101) for producing the control pressure (Pg).
3. A hydraulically operated device comprising:
a plurality of hydraulic actuators (5) to which pressured oil discharged from a variable delivery hydraulic pump (1) is supplied via a pressure compensating valve (7) and a directional control valve (4);
pressure output means (8) for outputting pressure (PLS) to a load pressure sensing passage (9) according to maximum load pressure among load pressures acting on the actuators (5); and
pump discharge pressure control means (12) for controlling discharge pressure of the variable delivery hydraulic pump (1) based on the pressure (PLS) output from the pressure output means (8),
characterized in that a variable bleed valve (11) is provided in the load pressure sensing passage (9).
4. A hydraulically operated device comprising:
a pressure compensating valve through which pressured oil that is fed from a hydraulic pump (1) to a hydraulic actuator (5) passes; and
an unloading pressure control valve for introducing discharged pressured oil from the hydraulic pump (1) to a tank according to the pressure difference between discharge pressure (PP) of the hydraulic pump (1) and load pressure (PLS) of the hydraulic actuator (5),
characterized by further comprising:
a pressure compensating valve (7), that is operated in such a way as to increase area of an opening between an inlet port (24) and an outlet port (25) by means of pressure acting on a first pressure receiving component (21) for the pressure compensating valve, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component (22) for the pressure compensating valve and pressure acting on a third pressure receiving component (23) for the pressure compensating valve, and comprises:
a main valve (20) for the pressure compensating valve for allowing pressure (Pa) of the pressured oil flowing into the inlet port (24) to act on the first pressure receiving component (21) for the pressure compensating valve and pressure (Pb) of a load driven by the pressured oil flowing out from the outlet port (25) to act on the second pressure receiving component (22) for the pressure compensating valve; and
control pressure producing means (7B) for allowing control pressure (Pe) resulting from a reduction in the pressure (Pa) of the inlet port (24) to act on the third pressure receiving component (23) for the pressure compensating valve, and an unloading pressure control valve (10), comprising:
a main valve (100), that operates toward communicating direction by means of discharge pressure (PP) of the hydraulic pump (1) acting on a first pressure receiving component (123) for the unloading pressure control valve, and toward cut-off direction by means of load pressure (PLS) acting on-a second pressure receiving component (124) for an unloading pressure control valve, and changes balance of operating forces in each of the directions by means of control pressure (Pg) acting on a third pressure receiving component (125) for an unloading pressure control valve; and
control pressure producing means (101) for producing the control pressure (Pg).
5. A hydraulically operated device comprising:
a plurality of hydraulic actuators (5) to which pressured oil discharged from a variable delivery hydraulic pump (1) is supplied via a pressure compensating valve (7) and a directional control valve (4);
pressure output means (8) for outputting pressure (PLS) to a load pressure sensing passage (9) according to maximum load pressure among load pressures acting on the actuators (5); and
pump discharge pressure control means (12) for controlling discharge pressure of the variable delivery hydraulic pump (1) based on the pressure (PLS) output from the pressure output means (8),
characterized by further comprising:
a pressure compensating valve (7) comprising:
a main valve (20), that operates in such a way as to increase an area of an opening between an inlet port (24) and an outlet port (25) thereof by means of pressure acting on a first pressure receiving component (21), and operates in such a way as to reduce the area of the opening by means of pressure-acting on a second pressure receiving component (22) and pressure acting on a third pressure receiving component (23), and that allows pressure (Pa) of the pressured oil flowing into the inlet port (24) to act on the first pressure receiving component (21) and pressure (Pb) of a load driven by the pressured oil flowing out from the outlet port (25) to act on the second pressure receiving component (22); and
control pressure producing means (7B) for allowing control pressure (Pe) resulting from a reduction in the pressure (Pa) of the inlet port (24) to act on the third pressure receiving component (23), and
a variable bleed valve (11) provided in the load pressure sensing passage (9).
6. A hydraulically operated device comprising:
a plurality of hydraulic actuators (5) to which pressured oil discharged from a variable delivery hydraulic pump (1) is supplied via a pressure compensating valve (7) and a directional control valve (4);
pressure output means (8) for outputting pressure (PLS) to a load pressure sensing passage (9) according to maximum load pressure among load pressures acting on the actuators (5);
pump discharge pressure control means (12) for controlling discharge pressure of the variable delivery hydraulic pump (1) based on the pressure (PLS) output from the pressure output means (8); and
an unloading pressure control valve (10) for introducing a pressured oil discharged from a hydraulic pump (1) to a tank according to pressure difference between discharge pressure (PP) of the hydraulic pump (1) and load pressure (PLS) of a hydraulic actuator (5),
characterized in that
the unloading pressure control valve (10) comprises:
a main valve (100), that operates toward a communicating direction by means of discharge pressure (PP) of the hydraulic pump (1) acting on a first pressure receiving component (123), to operate toward a cut-off direction by means of load pressure (PLS) acting on a second pressure receiving component (124), and that changes balance of an operating force in each of the directions by means of control pressure (Pg) acting on a third pressure receiving component (125); and
control pressure producing means (101) for producing the control pressure (Pg), and
a variable bleed valve (11) is provided in the load pressure sensing passage (9).
7. A hydraulically operated device comprising:
a plurality of hydraulic actuators (5) to which pressured oil discharged from a variable delivery hydraulic pump (1) is supplied via a pressure compensating valve (7) and a directional control valve (4);
pressure output means (8) for outputting pressure (PLS) to a load pressure sensing passage (9) according to maximum load pressure among load pressures acting on the actuators (5);
pump discharge pressure control means (12) for controlling discharge pressure of the variable delivery hydraulic pump (1) based on the pressure (PLS) output from the pressure output means (8); and
an unloading pressure control valve (10) for introducing a pressured oil discharged from a hydraulic pump (1) to a tank according to pressure difference between discharge pressure (PP) of the hydraulic pump (1) and load pressure (LS) of a hydraulic actuator (5),
characterized by further comprising:
a pressure compensating valve (7), that is operated in such a way as to increase area of an opening between an inlet port (24) and an outlet port (25) by means of pressure acting on a first pressure receiving component (21) for the pressure compensating valve, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component (22) for the pressure compensating valve and pressure acting on a third pressure receiving component (23) for the pressure compensating valve, and comprises:
a main valve (20) for the pressure compensating valve for allowing pressure (Pa) of the pressured oil flowing into the inlet port (24) to act on the first pressure receiving component (21) for the pressure compensating valve and pressure (Pb) of a load driven by the pressured oil flowing out from the outlet port (25) to act on the second pressure receiving component (22) for the pressure compensating valve; and
control pressure producing means (7B) for allowing control pressure (Pe) resulting from a reduction in the pressure (Pa) of the inlet port (24) to act on the third pressure receiving component (23) for the pressure compensating valve;
an unloading pressure control valve (10), comprising:
a main valve (100), that operates toward communicating direction by means of discharge pressure (PP) of the hydraulic pump (1) acting on a first pressure receiving component (123) for the unloading pressure control valve, and toward cut-off direction by means of load pressure (PLS) acting on a second pressure receiving component (124) for an unloading pressure control valve, and changes balance of operating forces in each of the directions by means of control pressure (Pg) acting on a third pressure receiving component (125) for an unloading pressure control valve; and
control pressure producing means (101) for producing the control pressure (Pg.), and
a variable bleed valve (11) provided in the load pressure sensing passage (9).
Priority Applications (1)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
US09/964,491 US20020007631A1 (en) | 1998-03-04 | 2001-09-28 | Pressure compensating valve, unloading pressure control valve and hydraulically operated device |
Applications Claiming Priority (14)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
JP5192498 | 1998-03-04 | ||
JP51,892/1998 | 1998-03-04 | ||
JP5189298 | 1998-03-04 | ||
JP51,913/1998 | 1998-03-04 | ||
JP5191398 | 1998-03-04 | ||
JP51,924/1998 | 1998-03-04 | ||
JP35482398A JPH11316611A (en) | 1998-03-04 | 1998-12-14 | Pressure compensating valve |
JP354,823/1998 | 1998-12-14 | ||
JP5,502/1999 | 1999-01-12 | ||
JP550399A JPH11315806A (en) | 1998-03-04 | 1999-01-12 | Hydraulic driving device |
JP550299A JPH11315805A (en) | 1998-03-04 | 1999-01-12 | Unload valve |
JP5,503/1999 | 1999-01-12 | ||
US09/261,390 US6334308B1 (en) | 1998-03-04 | 1999-03-03 | Pressure compensating valve, unloading pressure control valve and hydraulically operated device |
US09/964,491 US20020007631A1 (en) | 1998-03-04 | 2001-09-28 | Pressure compensating valve, unloading pressure control valve and hydraulically operated device |
Related Parent Applications (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
US09/261,390 Continuation US6334308B1 (en) | 1998-03-04 | 1999-03-03 | Pressure compensating valve, unloading pressure control valve and hydraulically operated device |
Publications (1)
Publication Number | Publication Date |
---|---|
US20020007631A1 true US20020007631A1 (en) | 2002-01-24 |
Family
ID=27547917
Family Applications (2)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
US09/261,390 Expired - Fee Related US6334308B1 (en) | 1998-03-04 | 1999-03-03 | Pressure compensating valve, unloading pressure control valve and hydraulically operated device |
US09/964,491 Abandoned US20020007631A1 (en) | 1998-03-04 | 2001-09-28 | Pressure compensating valve, unloading pressure control valve and hydraulically operated device |
Family Applications Before (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
US09/261,390 Expired - Fee Related US6334308B1 (en) | 1998-03-04 | 1999-03-03 | Pressure compensating valve, unloading pressure control valve and hydraulically operated device |
Country Status (2)
Country | Link |
---|---|
US (2) | US6334308B1 (en) |
DE (1) | DE19909480A1 (en) |
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US20070204607A1 (en) * | 2006-02-27 | 2007-09-06 | Kobelco Construction Machinery Co., Ltd. | Hydraulic circuit of construction machine |
US20090212512A1 (en) * | 2005-05-04 | 2009-08-27 | Markus Hoermann | Method and device for increasing the driving stability of motor vehicles |
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DE3733677A1 (en) * | 1987-10-05 | 1989-04-13 | Rexroth Mannesmann Gmbh | LOAD-INDEPENDENT CONTROL DEVICE FOR HYDRAULIC CONSUMERS |
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1999
- 1999-03-03 US US09/261,390 patent/US6334308B1/en not_active Expired - Fee Related
- 1999-03-04 DE DE19909480A patent/DE19909480A1/en not_active Withdrawn
-
2001
- 2001-09-28 US US09/964,491 patent/US20020007631A1/en not_active Abandoned
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US20070204607A1 (en) * | 2006-02-27 | 2007-09-06 | Kobelco Construction Machinery Co., Ltd. | Hydraulic circuit of construction machine |
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Also Published As
Publication number | Publication date |
---|---|
DE19909480A1 (en) | 1999-09-09 |
US6334308B1 (en) | 2002-01-01 |
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