WO2015080112A1 - 建設機械の油圧駆動装置 - Google Patents
建設機械の油圧駆動装置 Download PDFInfo
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- WO2015080112A1 WO2015080112A1 PCT/JP2014/081146 JP2014081146W WO2015080112A1 WO 2015080112 A1 WO2015080112 A1 WO 2015080112A1 JP 2014081146 W JP2014081146 W JP 2014081146W WO 2015080112 A1 WO2015080112 A1 WO 2015080112A1
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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
- E02F9/2221—Control of flow rate; Load sensing arrangements
- E02F9/2232—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps
- E02F9/2235—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps including an electronic controller
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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
- E02F9/2225—Control of flow rate; Load sensing arrangements using pressure-compensating valves
- E02F9/2228—Control of flow rate; Load sensing arrangements using pressure-compensating valves including an electronic controller
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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
- E02F9/2232—Control of flow rate; Load sensing arrangements using one or more variable displacement 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
- E02F9/2264—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
- E02F9/2278—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
- F15B11/17—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors using two or more 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
- 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
- F15B13/025—Pressure reducing 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
- 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
- F15B13/026—Pressure compensating 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
- 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
- F15B13/06—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with two or more servomotors
- F15B13/08—Assemblies of units, each for the control of a single servomotor only
- F15B13/0803—Modular units
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- E—FIXED CONSTRUCTIONS
- E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
- E02F—DREDGING; SOIL-SHIFTING
- E02F3/00—Dredgers; Soil-shifting machines
- E02F3/04—Dredgers; Soil-shifting machines mechanically-driven
- E02F3/28—Dredgers; Soil-shifting machines mechanically-driven with digging tools mounted on a dipper- or bucket-arm, i.e. there is either one arm or a pair of arms, e.g. dippers, buckets
- E02F3/30—Dredgers; Soil-shifting machines mechanically-driven with digging tools mounted on a dipper- or bucket-arm, i.e. there is either one arm or a pair of arms, e.g. dippers, buckets with a dipper-arm pivoted on a cantilever beam, i.e. boom
- E02F3/32—Dredgers; Soil-shifting machines mechanically-driven with digging tools mounted on a dipper- or bucket-arm, i.e. there is either one arm or a pair of arms, e.g. dippers, buckets with a dipper-arm pivoted on a cantilever beam, i.e. boom working downwardly and towards the machine, e.g. with backhoes
- E02F3/325—Backhoes of the miniature type
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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
- F15B20/00—Safety arrangements for fluid actuator systems; Applications of safety devices in fluid actuator systems; Emergency measures for fluid actuator systems
- F15B20/007—Overload
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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
- F15B2211/20507—Type of prime mover
- F15B2211/20523—Internal combustion engine
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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
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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
- F15B2211/00—Circuits for servomotor systems
- F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
- F15B2211/25—Pressure control functions
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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
- F15B2211/60—Circuit components or control therefor
- F15B2211/665—Methods of control using electronic components
- F15B2211/6652—Control of the pressure source, e.g. control of the swash plate angle
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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
- F15B2211/60—Circuit components or control therefor
- F15B2211/665—Methods of control using electronic components
- F15B2211/6655—Power control, e.g. combined pressure and flow rate control
Definitions
- the present invention relates to a hydraulic drive device for a construction machine such as a hydraulic excavator, and in particular, includes a pump control device (regulator) including at least two variable displacement hydraulic pumps, and one of the hydraulic pumps performing at least torque control.
- a hydraulic drive device for a construction machine having a pump control device (regulator) that has load sensing control and torque control.
- Some hydraulic drive devices for construction machines such as hydraulic excavators are equipped with a regulator that controls the capacity (flow rate) of the hydraulic pump so that the discharge pressure of the hydraulic pump is higher than the maximum load pressure of multiple actuators by the target differential pressure. Widely used, this control is called load sensing control.
- this control is called load sensing control.
- two hydraulic pumps are provided in a hydraulic drive device for a construction machine having a regulator for performing such load sensing control, and two pumps are used to perform load sensing control in each of the two hydraulic pumps. A load sensing system is described.
- the torque of the hydraulic pump so that the absorption torque of the hydraulic pump does not exceed the rated output torque of the prime mover by reducing the capacity of the hydraulic pump as the discharge pressure of the hydraulic pump increases. Control is performed to prevent the prime mover from stopping due to overtorque (engine stall).
- the regulator of one hydraulic pump takes in parameters related to the absorption torque of the other hydraulic pump as well as its own discharge pressure, and performs torque control (total torque control), It is intended to prevent the stoppage of the prime mover and effectively use the rated output torque of the prime mover.
- Patent Document 2 the discharge pressure of one hydraulic pump is led to the regulator of the other hydraulic pump via a pressure reducing valve to perform total torque control.
- the set pressure of the pressure reducing valve is constant, and this set pressure is set to a value simulating the maximum torque for torque control of the regulator of the other hydraulic pump.
- Patent Document 3 in order to perform full torque control on two variable displacement hydraulic pumps, the tilt angle of the other hydraulic pump is detected as the output pressure of the pressure reducing valve, and the output pressure is detected as one hydraulic pressure. Leads to the pump regulator.
- Patent Document 4 the control accuracy of the total torque control is improved by replacing the tilt angle of the other hydraulic pump with the arm length of the swing arm.
- JP 2011-196438 A Japanese Patent No. 3865590 Japanese Patent Publication No. 3-7030 JP-A-7-189916
- the other hydraulic pump is not limited by torque control and is in an operation state in which capacity control is performed by load sensing control
- the absorption torque of the other hydraulic pump is smaller than the maximum torque of torque control.
- the output pressure of the pressure reducing valve simulating the maximum torque is guided to the regulator of one hydraulic pump, and control is performed to reduce the absorption torque of one hydraulic pump more than necessary. For this reason, the total torque control cannot be performed with high accuracy.
- the inclination angle of the other hydraulic pump is detected as the output pressure of the pressure reducing valve, and the output pressure is guided to the regulator of the one hydraulic pump to improve the accuracy of the total torque control.
- the torque of the pump is obtained by the product of the discharge pressure and the capacity, that is, (discharge pressure ⁇ pump capacity) / 2 ⁇ . Leads to one of the two pilot chambers, guides the output pressure of the pressure reducing valve (the discharge pressure proportional to the other hydraulic pump) to the other pilot chamber of the stepped piston, and outputs the sum of the discharge pressure and the discharge amount proportional pressure to the output torque Since the capacity of one of the hydraulic pumps is controlled as a parameter of this, there is a problem that a considerable error occurs between the actually used torque.
- Patent Document 4 the control accuracy of the total torque control is improved by replacing the tilt angle of the other hydraulic pump with the arm length of the swing arm.
- the regulator of Patent Document 4 has a very complicated structure in which the swing arm and the piston provided in the regulator piston slide relative to each other while transmitting force, and have sufficient durability.
- components such as a swing arm and a regulator piston have to be strengthened, and there is a problem that it is difficult to downsize the regulator.
- the space for storing the hydraulic pump is small and it may be difficult to mount.
- the absorption torque of the other hydraulic pump is accurately detected with a pure hydraulic configuration and fed back to the hydraulic pump side, so that all torque control is performed accurately and the rated output torque of the prime mover is effective.
- the present invention provides a prime mover, a variable displacement first hydraulic pump driven by the prime mover, a variable displacement second hydraulic pump driven by the prime mover, A plurality of actuators driven by pressure oil discharged by the first and second hydraulic pumps, and a plurality of flow rate controls for controlling the flow rates of the pressure oil supplied from the first and second hydraulic pumps to the plurality of actuators A valve, a plurality of pressure compensating valves that respectively control the differential pressure across the plurality of flow control valves, a first pump control device that controls a discharge flow rate of the first hydraulic pump, and a discharge flow rate of the second hydraulic pump A first pump control device that controls at least one of a discharge pressure and a capacity of the first hydraulic pump, and an absorption torque of the first hydraulic pump is increased.
- the second pump control device When increasing, it has a first torque control unit that controls the capacity of the first hydraulic pump so that the absorption torque of the first hydraulic pump does not exceed the first maximum torque, and the second pump control device includes: When at least one of the discharge pressure and capacity of the second hydraulic pump increases and the absorption torque of the second hydraulic pump increases, the second hydraulic pump does not exceed the second maximum torque so that the absorption torque of the second hydraulic pump does not exceed the second maximum torque.
- the absorption torque of the second torque control unit that controls the capacity of the hydraulic pump and the second hydraulic pump is smaller than the second maximum torque, the discharge pressure of the second hydraulic pump is discharged by the second hydraulic pump.
- a load sensing control unit for controlling the capacity of the second hydraulic pump so as to be higher than the maximum load pressure of the actuator driven by the pressurized oil by a target differential pressure.
- the first torque control unit is configured such that the discharge pressure of the first hydraulic pump is guided, and the capacity of the second hydraulic pump is decreased and the absorption torque is decreased when the discharge pressure is increased.
- a first torque control actuator that controls the capacity of the first torque, and a first urging means that sets the first maximum torque, wherein the second torque control unit is guided with a discharge pressure of the second hydraulic pump,
- a second torque control actuator for controlling the capacity of the second hydraulic pump so as to decrease the capacity of the second hydraulic pump and decrease the absorption torque when the discharge pressure increases, and a second appendage for setting the second maximum torque.
- the load sensing control unit is configured to reduce the load sensor so that the differential pressure between the discharge pressure of the second hydraulic pump and the maximum load pressure becomes smaller than the target differential pressure.
- the first pump control device further guides the discharge pressure of the second hydraulic pump and the load sensing drive pressure, and the second hydraulic pump restricts the control of the second torque control unit.
- the second hydraulic pump is not limited by the control of the second torque control unit, and the load sensing control unit controls the capacity of the second hydraulic pump.
- the discharge pressure of the second hydraulic pump and the load sensing drive are set so as to simulate the absorption torque of the second hydraulic pump.
- a torque feedback circuit that corrects the discharge pressure of the second hydraulic pump based on the pressure and outputs the corrected pressure as a torque control pressure, and the torque control pressure is derived, and the torque of the first hydraulic pump increases as the torque control pressure increases.
- a third torque control actuator for controlling the capacity of the first hydraulic pump so as to reduce the capacity and reduce the first maximum torque, and the torque feedback circuit guides the discharge pressure of the second hydraulic pump.
- a pressure limiting valve that controls the second torque control unit so as not to exceed a pressure at which control is started, and the variable throttle valve includes the load sensing driving pressure.
- the second hydraulic pump when the second hydraulic pump is not limited by the control of the second torque control unit and the load sensing control unit controls the capacity of the second hydraulic pump (discharge of the second hydraulic pump).
- the pressure in the oil passage between the fixed throttle and the variable throttle valve increases as the discharge pressure of the second hydraulic pump increases and load sensing. It becomes smaller as the driving pressure becomes higher.
- This change in pressure is caused when the discharge pressure of the second hydraulic pump increases when the second hydraulic pump is not limited by the control of the second torque control unit and the load sensing control controls the capacity of the second hydraulic pump. Therefore, it is approximated to a change in the absorption torque of the second hydraulic pump that increases and decreases as the load sensing drive pressure increases (the capacity of the second hydraulic pump decreases).
- the torque control pressure is generated based on the pressure in the oil passage between the fixed throttle and the variable throttle valve, and the change in the torque control pressure is also approximated to the change in the absorption torque of the second hydraulic pump.
- the absorption torque of the second hydraulic pump can be accurately detected with a pure hydraulic configuration, and the torque feedback circuit adjusts the discharge pressure of the second hydraulic pump so as to simulate the absorption torque of the second hydraulic pump. It can be corrected and output as torque control pressure.
- the torque control pressure is guided to the third torque control actuator, and the absorption torque of the second hydraulic pump is fed back to the first hydraulic pump (one hydraulic pump) side, so that the second hydraulic pump becomes the second torque control unit.
- the second hydraulic pump is controlled by the second torque control unit and the load sensing control unit controls the capacity of the second hydraulic pump.
- the first maximum torque set in the first torque control unit of the first hydraulic pump can be reduced by the amount of absorption torque of the second hydraulic pump, and the total torque control can be performed with high accuracy.
- the rated output torque can be used effectively.
- the absorption torque of the second hydraulic pump is detected in a pure hydraulic manner, the first pump control device can be reduced in size and the mountability is improved.
- the torque feedback circuit further includes a pressure reducing valve to which a discharge pressure of the second hydraulic pump is guided as a primary pressure, and between the fixed throttle and the variable throttle valve.
- the pressure of the oil passage is led to the pressure reducing valve as a target control pressure for setting the set pressure of the pressure reducing valve, and the pressure reducing valve is configured to change the first pressure when the discharge pressure of the second hydraulic pump is lower than the set pressure.
- the discharge pressure of the second hydraulic pump is directly output as the secondary pressure, and when the discharge pressure of the second hydraulic pump is higher than the set pressure, the discharge pressure of the second hydraulic pump is reduced to the set pressure and output. Then, the output pressure of the pressure reducing valve is guided to the third torque control actuator as the torque control pressure.
- the pressure in the oil passage between the fixed throttle and the variable throttle valve is not directly used as the torque control pressure, the setting of the fixed throttle and the variable throttle valve for obtaining the necessary target control pressure and the third torque control actuator The responsiveness can be set independently, and the torque feedback circuit for achieving the required performance can be set easily and accurately.
- the discharge pressure of the second hydraulic pump is higher than the set pressure of the pressure reducing valve, the fluctuation of the discharge pressure of the second hydraulic pump is blocked by the pressure reducing valve and does not affect the third torque control actuator, so that the stability of the system is improved. Is secured.
- the pressure limiting valve is a relief valve.
- the absorption torque of the second hydraulic pump can be accurately detected with a pure hydraulic configuration (torque feedback circuit), and the absorption torque is transferred to the first hydraulic pump (one hydraulic pump) side.
- torque feedback circuit torque feedback circuit
- total torque control can be performed with high accuracy, and the rated output torque of the prime mover can be used effectively.
- the absorption torque of the second hydraulic pump is detected in a pure hydraulic manner, the first pump control device can be reduced in size and the mountability is improved. As a result, it is possible to provide an energy efficient, low fuel consumption and practical construction machine.
- 1 is a hydraulic circuit diagram showing an entire hydraulic drive device of a hydraulic excavator (construction machine) according to a first embodiment of the present invention.
- 1 is a hydraulic circuit diagram showing details of a torque feedback circuit of a hydraulic drive device of a hydraulic excavator (construction machine) according to a first embodiment of the present invention.
- 1 is a block diagram illustrating an entire hydraulic drive device of a hydraulic excavator (construction machine) according to a first embodiment of the present invention. It is a figure which shows the relationship between LS drive pressure when a load sensing control piston operate
- FIGS. 1A and 2 are views showing a hydraulic drive device of a hydraulic excavator (construction machine) according to a first embodiment of the present invention
- FIG. 1A is a hydraulic circuit diagram showing the entire hydraulic drive device
- FIG. 2 is a block diagram showing the entire hydraulic drive apparatus
- FIG. 1B is a hydraulic circuit diagram showing details of the torque feedback circuit shown in FIGS. 1A and 2.
- the hydraulic drive device includes a variable displacement first hydraulic pump 1a having first and second two discharge ports P1 and P2, and third and fourth two.
- a variable displacement type second hydraulic pump 1b having two discharge ports P3, P4, a prime mover 2 connected to the first and second hydraulic pumps 1a, 1b and driving the first and second hydraulic pumps 1a, 1b;
- a plurality of actuators 3a driven by the discharge oil of the first and second discharge ports P1, P2 of the first and second hydraulic pumps 1a and the discharge oil of the third and fourth discharge ports P3, P4 of the second hydraulic pump 1b.
- the capacity of the first hydraulic pump 1a and the capacity of the second hydraulic pump 1b are the same.
- the capacity of the first hydraulic pump 1a and the capacity of the second hydraulic pump 1b may be different.
- the first hydraulic pump 1a has a first pump control device (regulator) 5a provided in common with respect to the first and second discharge ports P1, P2.
- the second hydraulic pump 1b includes third and It has the 2nd pump control apparatus (regulator) 5b provided in common with respect to 4th discharge port P3, P4.
- the first hydraulic pump 1a is a split flow type hydraulic pump provided with a single displacement control mechanism (swash plate), and the first pump controller 5a drives the single displacement control mechanism to The displacement (tilt angle of the swash plate) of one hydraulic pump 1a is controlled to control the discharge flow rates of the first and second discharge ports P1, P2.
- the second hydraulic pump 1b is a split flow type hydraulic pump having a single displacement control mechanism (swash plate), and the second pump control device 5b drives the single displacement control mechanism.
- the capacity of the second hydraulic pump 1b tilt angle of the swash plate
- the discharge flow rates of the third and fourth discharge ports P3 and P4 are controlled.
- Each of the first and second hydraulic pumps 1a and 1b may be a combination of two variable displacement hydraulic pumps each having a single discharge port.
- the first hydraulic pump 1a 2 Two capacity control mechanisms (swash plates) of the two hydraulic pumps are driven by the first pump control device 5a, and two capacity control mechanisms (swash plates) of the two hydraulic pumps of the second hydraulic pump 1b are driven by the second pump. What is necessary is just to drive with the control apparatus 5b.
- the prime mover 2 is, for example, a diesel engine.
- the diesel engine includes, for example, an electronic governor, and the rotational speed and torque are controlled by controlling the fuel injection amount.
- the engine speed is set by operating means such as an engine control dial.
- the prime mover 2 may be an electric motor.
- the control valve 4 is connected to a plurality of closed center type flow control valves 6a to 6m and upstream of the flow control valves 6a to 6m, and controls the differential pressure across the meter-in throttle of the flow control valves 6a to 6m.
- the third shuttle valve group 8c for detecting the pressure and the load pressure port of the flow control valves 6j to 6m are connected to the actuators 3d, 3g, 3h and the flow control valve 6m.
- the fourth shuttle valve group 8d that detects the maximum load pressure of the spare actuator when the actuator is connected and the discharge ports P1 and P2 of the first hydraulic pump 1a are connected to the discharge ports P1 and P2, respectively.
- the discharge ports P1 and P2 are opened.
- the discharge oil is returned to the tank and connected to the first and second unload valves 10a and 10b for limiting the increase in discharge pressure, and the discharge ports P3 and P4 of the second hydraulic pump 1b, respectively.
- the third and fourth unload valves 10c and 10d for returning the discharge oil from the discharge ports P3 and P4 to the tank and limiting the increase in discharge pressure, and the first and second discharge ports of the first hydraulic pump 1a.
- the first communication control valve 15a disposed between the discharge oil passages of P1 and P2 and the output oil passages of the first and second shuttle valve groups 8a and 8b, and the second hydraulic pump 1b.
- a second communication control valve 15b disposed between the discharge oil passages of the third and fourth discharge ports P3 and P4 and between the output oil passages of the third and fourth shuttle valve groups 8c and 8d. ing.
- the set pressure of the springs 9a to 9d of the first to fourth unload valves 10a to 10d is set to a pressure that is equal to or slightly higher than a target differential pressure of load sensing control described later.
- control valve 4 is connected to the discharge ports P1 and P2 of the first hydraulic pump 1a, respectively, and the first and second main relief valves functioning as safety valves, and the discharge port of the second hydraulic pump 1b.
- the third and fourth main relief valves are connected to P3 and P4 and function as safety valves.
- the pressure compensation valves 6a to 6f provide a target compensation differential pressure, which is a differential pressure between the discharge pressure of the discharge ports P1 and P2 of the first hydraulic pump 1a and the maximum load pressure detected by the first and second shuttle valve groups 8a and 8b.
- the pressure compensation valves 7g to 7m are configured so that the discharge pressures of the discharge ports P3 and P4 of the second hydraulic pump 1b and the maximum load pressure detected by the third and fourth shuttle valve groups 8c and 8d Is set as the target compensation differential pressure.
- the discharge pressure of the first discharge port P1 is guided to the opening direction operation side, and the highest of the actuators 3a to 3e detected by the first and second shuttle valve groups 8a and 8b.
- the load pressure is guided to the closing direction operation side, and the differential pressure across the meter-in throttle portion of the flow rate control valves 6a to 6c is controlled to be equal to the differential pressure between the two.
- the pressure compensation valves 7d to 7f the discharge pressure of the second discharge port P2 is guided to the opening direction operation side, and the maximum load pressure of the actuators 3a to 3e detected by the first and second shuttle valve groups 8a and 8b is the closing direction. Guided to the operating side, control is performed so that the differential pressure across the meter-in throttle portions of the flow control valves 6d to 6f is equal to the differential pressure between them.
- the discharge pressure of the third discharge port P3 is guided to the opening direction operation side, and the maximum load pressure of the actuators 3d to 3h detected by the third and fourth shuttle valve groups 8c and 8d is the closing direction. Guided to the operating side, control is performed so that the differential pressure across the meter-in throttle portion of the flow control valves 6g to 6i is equal to the differential pressure between the two.
- the discharge pressure of the fourth discharge port P4 is guided to the opening direction operation side, and the maximum load pressure of the actuators 3d to 3h detected by the third and fourth shuttle valve groups 8c and 8d is the closing direction.
- control is performed so that the differential pressure across the meter-in throttle of the flow rate control valves 6j to 6m is equal to the differential pressure between the two.
- the flow rate corresponding to the opening area ratio of the flow control valve is controlled regardless of the load pressure of the actuator.
- the differential pressure across the meter-in throttle portion of the flow control valve is adjusted according to the degree of saturation. It is possible to reduce and secure good composite operability.
- the plurality of actuators 3a to 3d are, for example, an arm cylinder of a hydraulic excavator, a bucket cylinder, a turning motor, and a left traveling motor, respectively, and the plurality of actuators 3e to 3h are, for example, a right traveling motor, a swing cylinder, a blade cylinder, Boom cylinder.
- the arm cylinder 3a has flow control valves 6a and 6e and a pressure compensation valve 7a so that the discharge oils of both the first and second discharge ports P1 and P2 of the first hydraulic pump 1a are joined and supplied.
- 7e are connected to the first and second discharge ports P1, P2, and the boom cylinder 3h is supplied by supplying the discharge oil from both the third and fourth discharge ports P3, P4 of the second hydraulic pump 1b.
- the third and fourth discharge ports P3 and P4 are connected via the flow rate control valves 6h and 6l and the pressure compensation valves 7h and 7l.
- the left travel motor 3d includes a second discharge port P2 which is a discharge port on one side of the first and second discharge ports P1 and P2 of the first hydraulic pump 1a, and a third and fourth discharge of the second hydraulic pump 1b.
- the second and fourth flow control valves 6f and 6j and the pressure compensation valves 7f and 7j are used so that the discharge oil from the fourth discharge port P4, which is a discharge port on one side of the ports P3 and P4, is joined and supplied.
- the travel right travel motor 3e connected to the discharge ports P2, P4 includes a first discharge port P1 which is a discharge port on the other side of the first and second discharge ports P1, P2 of the first hydraulic pump 1a, and a first discharge port P1.
- the flow rate control valves 6c and 6g and the pressure are controlled so that the discharge oil from the third discharge port P3 which is the discharge port on the other side of the third and fourth discharge ports P3 and P4 of the hydraulic pump 1b is joined and supplied.
- First through compensation valves 7c, 7g Beauty is connected to a third discharge port P1, P3.
- the bucket cylinder 3b is connected to the first discharge port P1 via the flow control valve 6b and the pressure compensation valve 7b so that the discharge oil from the first discharge port P1 of the first hydraulic pump 1a is supplied,
- the motor 3c is connected to the second discharge port P2 via the flow control valve 6d and the pressure compensation valve 7d so that the discharge oil from the second discharge port P2 of the first hydraulic pump 1a is supplied.
- the swing cylinder 3f is connected to the third discharge port P3 via the flow rate control valve 6i and the pressure compensation valve 7i so that the discharge oil of the third discharge port P3 of the second hydraulic pump 1b is supplied, and the blade cylinder 3g
- the second hydraulic pump 1b is connected to the fourth discharge port P4 via the flow control valve 6k and the pressure compensation valve 7k so that the discharge oil from the fourth discharge port P4 of the second hydraulic pump 1b is supplied.
- the flow control valve 6m and the pressure compensation valve 7m are spare (accessories). For example, when the bucket 308 is replaced with a crusher, the opening / closing cylinder of the crusher is connected to the fourth through the flow control valve 6m and the pressure compensation valve 7m. Connected to the discharge port P4.
- the first communication control valve 15a is not in the combined operation of simultaneously driving the travel motors 3d, 3e and at least one of the other actuators (boom cylinder 3a, bucket cylinder 3b, swing motor 3c) related to the first hydraulic pump 1a ( (Hereinafter referred to as other than the travel combined operation) is at the upper cut-off position in the figure, and the lower side is illustrated during the combined operation for simultaneously driving the travel motors 3d and 3e and at least one of the other actuators (hereinafter referred to as the travel combined operation). Switch to the communication position.
- the second communication control valve 15b is not in the combined operation of simultaneously driving the travel motors 3d, 3e and at least one of the other actuators (swing cylinder 3f, blade cylinder 3g, boom cylinder 3h) related to the second hydraulic pump 1b ( (Hereinafter referred to as other than the travel combined operation) is at the upper cut-off position in the figure, and the lower side is illustrated during the combined operation for simultaneously driving the travel motors 3d and 3e and at least one of the other actuators (hereinafter referred to as the travel combined operation). Switch to the communication position.
- the first hydraulic pump 1a blocks the communication of the discharge oil passages of the first and second discharge ports P1 and P2, and the lower communication in the figure.
- the discharge oil passages of the first and second discharge ports P1, P2 of the first hydraulic pump 1a are communicated.
- the second communication control valve 15b When the second communication control valve 15b is at the upper cutoff position in the figure, the communication of the discharge oil passages of the third and fourth discharge ports P3 and P4 of the second hydraulic pump 1b is cut off. When switched to the communication position on the side, the discharge oil passages of the third and fourth discharge ports P3 and P4 of the second hydraulic pump 1b are connected.
- the first communication control valve 15a has a built-in shuttle valve.
- the output oil path of the first shuttle valve group 8a and the output oil path of the second shuttle valve group 8b When the communication is cut off and the output oil passages of the first and second shuttle valve groups 8a and 8b are communicated with the downstream sides of the first and second shuttle valve groups 8a and 8b, respectively, the first and second shuttle valves are switched.
- the output oil passages of the groups 8a and 8b are communicated with each other via a shuttle valve, and the highest load pressure on the high pressure side is led out to the downstream side.
- the second communication control valve 15b has a built-in shuttle valve.
- the output oil passage of the third shuttle valve group 8c and the output oil passage of the fourth shuttle valve group 8d are connected.
- the respective output oil passages of the third and fourth shuttle valve groups 8c and 8d are communicated with the downstream sides thereof and switched to the lower communication position in the figure, the third and fourth shuttle valves
- the output oil passages of the groups 8c and 8d are communicated with each other via a shuttle valve, and the highest load pressure on the high pressure side is led out to the downstream side.
- the first communication control valve 15a When the first communication control valve 15a is in the upper shut-off position in the figure, on the first discharge port P1 side of the first hydraulic pump 1a, the maximum load pressure of the actuators 3a, 3b, 3e detected by the first shuttle valve group 8a Is led to the first unload valve 10a and the pressure compensation valves 7a to 7c, and the first unload valve 10a restricts the rise of the discharge pressure of the first discharge port P1 based on the maximum load pressure, and the pressure compensation valve 7a ⁇ 7c controls the differential pressure across the meter-in throttle of the flow control valves 6a-6c.
- the maximum load pressure of the actuators 3a, 3c, 3d detected by the second shuttle valve group 8b is guided to the second unload valve 10b and the pressure compensation valves 7d-7f.
- the second unload valve 10b limits the increase in the discharge pressure of the second discharge port P2, and the pressure compensation valves 7d to 7f are different from each other in the meter-in throttle portion of the flow control valves 6d to 6f. Control the pressure.
- the actuators 3a to 3 detected by the first and second shuttle valve groups 8a and 8b on the first discharge port P1 side of the first hydraulic pump 1a When the first communication control valve 15a is switched to the lower communication position in the figure, the actuators 3a to 3 detected by the first and second shuttle valve groups 8a and 8b on the first discharge port P1 side of the first hydraulic pump 1a.
- the maximum load pressure of 3e is guided to the first unload valve 10a and the pressure compensation valves 7a to 7c, and the first unload valve 10a limits the increase in the discharge pressure of the first discharge port P1 based on the maximum load pressure.
- the pressure compensation valves 7a to 7c control the differential pressure across the meter-in throttle portions of the flow control valves 6a to 6c.
- the maximum load pressure of the actuators 3a to 3e detected by the first and second shuttle valve groups 8a and 8b is the second unload valve 10b and the pressure compensation valve.
- the second unload valve 10b restricts the rise of the discharge pressure of the second discharge port P2 based on the maximum load pressure
- the pressure compensation valves 7d to 7f are meter-in of the flow control valves 6d to 6f. Controls the differential pressure across the throttle.
- the maximum load pressure of the actuators 3e, 3f, 3h detected by the third shuttle valve group 8c is detected on the third discharge port P3 side of the second hydraulic pump 1b. Is guided to the third unloading valve 10c and the pressure compensation valves 7g to 7i, and based on the maximum load pressure, the third unloading valve 10c limits the increase in the discharge pressure of the third discharge port P3, and the pressure compensation valve 7g ⁇ 7i controls the differential pressure across the meter-in throttle of the flow control valves 6g-6i.
- the maximum load pressure of the actuators 3d, 3g, 3h detected by the fourth shuttle valve group 8d is guided to the fourth unload valve 10d and the pressure compensation valves 7j to 7m.
- the fourth unload valve 10d restricts the rise of the discharge pressure of the fourth discharge port P4, and the pressure compensation valves 7j to 7m are different from each other in the meter-in throttle portion of the flow control valves 6j to 6m. Control the pressure.
- the actuators 3d to 3d detected by the third and fourth shuttle valve groups 8c and 8d on the third discharge port P3 side of the second hydraulic pump 1b When the second communication control valve 15b is switched to the lower communication position in the figure, the actuators 3d to 3d detected by the third and fourth shuttle valve groups 8c and 8d on the third discharge port P3 side of the second hydraulic pump 1b.
- the maximum load pressure of 3 h is led to the third unload valve 10c and the pressure compensation valves 7g to 7i, and the third unload valve 10c limits the increase in the discharge pressure of the third discharge port P3 based on the maximum load pressure.
- the pressure compensation valves 7g to 7i control the differential pressure across the meter-in throttle of the flow control valves 6g to 6i.
- the highest load pressures of the actuators 3d to 3h detected by the third and fourth shuttle valve groups 8c and 8d are the fourth unload valve 10d and the pressure compensation valve. 7j to 7m, the fourth unload valve 10d limits the rise of the discharge pressure of the fourth discharge port P4 based on the maximum load pressure, and the pressure compensation valves 7j to 7m are meter-in of the flow control valves 6j to 6m. Controls the differential pressure across the throttle.
- the discharge pressures of the first and second discharge ports P1 and P2 of the first hydraulic pump 1a are discharged from the first and second discharge ports P1 and P2 among the plurality of actuators 3a to 3h.
- a first load sensing control unit 12a for controlling the tilt angle (capacity) of the swash plate of the first hydraulic pump 1a so as to be higher by a predetermined pressure than the maximum load pressure of the actuators 3a to 3e driven by oil;
- a first torque control unit 13a for limiting and controlling the tilt angle (capacity) of the swash plate of the first hydraulic pump 1a so that the absorption torque of the hydraulic pump 1a does not exceed a predetermined value.
- the discharge pressures of the third and fourth discharge ports P3 and P4 of the second hydraulic pump 1b are discharged from the third and fourth discharge ports P3 and P4 among the plurality of actuators 3a to 3h.
- a second load sensing control unit 12b for controlling the tilt angle (capacity) of the swash plate of the second hydraulic pump 1b so as to be higher by a predetermined pressure than the maximum load pressure of the actuators 3d to 3h driven by oil;
- a second torque control unit 13b for limiting and controlling the tilt angle (capacity) of the swash plate of the second hydraulic pump 1b so that the absorption torque of the hydraulic pump 1b does not exceed a predetermined value.
- the first load sensing control unit 12a selects the load sensing control valves 16a and 16b that generate load sensing driving pressure (hereinafter referred to as LS driving pressure) and the low pressure side of the LS driving pressure generated by the load sensing control valves 16a and 16b.
- the low pressure selection valve 21a to be output and the LS driving pressure selected and output by the low pressure selection valve 21a are guided, and the load for changing the tilt angle of the swash plate of the first hydraulic pump 1a according to the LS driving pressure.
- a sensing control piston (load sensing control actuator) 17a a sensing control piston (load sensing control actuator) 17a.
- the second load sensing control unit 12b selects the load sensing control valves 16c and 16d that generate load sensing driving pressure (hereinafter referred to as LS driving pressure) and the low pressure side of the LS driving pressure generated by the load sensing control valves 16c and 16d.
- the low pressure selection valve 21b to be output and the LS driving pressure selected and output by the low pressure selection valve 21b are guided, and the load for changing the tilt angle of the swash plate of the second hydraulic pump 1b according to the LS driving pressure.
- a sensing control piston (load sensing control actuator) 17b a sensing control piston (load sensing control actuator) 17b.
- the control valve 16a is positioned opposite to the spring 16a1 for setting a target differential pressure for load sensing control, and the pressure received by the discharge pressure of the first discharge port P1.
- a portion 16a2 and a pressure receiving portion 16a3 located on the same side as the spring 16a1 are provided.
- the maximum load pressure of the actuators 3a to 3e detected by the first and second shuttle valve groups 8a and 8b is applied to the pressure receiving portion 16a3 of the control valve 16a.
- the control valve 16a includes the discharge pressure of the first discharge port P1 guided to the pressure receiving portion 16a2, the maximum load pressure of the actuators 3a, 3b, 3e or the actuators 3a to 3e guided to the pressure receiving portion 16a3, and the biasing force of the spring 16a1.
- the LS driving pressure is increased / decreased.
- the control valve 16a moves to the left in the figure and communicates the secondary port with the hydraulic pressure source (first discharge port P1) to increase the LS drive pressure, leading to the high pressure side of the first discharge port P1 guided to the pressure receiving portion 16a2.
- the control valve 16a moves to the right in the figure and becomes secondary.
- the LS driving pressure is lowered by connecting the port to the tank.
- the hydraulic pressure source that communicates with the secondary port when the control valve 16a moves to the left in the figure may be a pilot hydraulic pressure source that is formed in the discharge oil passage of the pilot pump and generates a constant pilot pressure.
- the control valve 16b is located on the same side as the spring 16b1 and a spring 16b1 that sets a target differential pressure for load sensing control, is opposed to the spring 16b1, and is guided by the discharge pressure of the second discharge port P2. And a pressure receiving portion 16b3 located at the same position.
- the first communication control valve 15a is in the upper shut-off position in the figure, the maximum load pressure of the actuators 3a, 3c, 3d detected by the second shuttle valve group 8b is guided to the pressure receiving portion 16b3 of the control valve 16b.
- the maximum load pressure of the actuators 3a to 3e detected by the first and second shuttle valve groups 8a and 8b is applied to the pressure receiving portion 16a3 of the control valve 16b.
- the control valve 16b includes the discharge pressure of the second discharge port P2 guided to the pressure receiving portion 16b2, the maximum load pressure of the actuators 3a, 3c, 3d or the actuators 3a to 3e guided to the pressure receiving portion 16b3, and the biasing force of the spring 16b1.
- the LS drive pressure is increased or decreased in the same manner as the control valve 16a.
- the low pressure selection valve 21a selects the low pressure side of the LS drive pressure generated by the load sensing control valves 16a and 16b and outputs it to the load sensing control piston 17a.
- the load sensing control piston 17a changes the tilt angle of the swash plate of the first hydraulic pump 1a based on the LS driving pressure, and increases or decreases the discharge flow rates of the first and second discharge ports P1, P2.
- control valve 16c is positioned opposite to the spring 16c1 for setting a target differential pressure for load sensing control and the pressure received by the discharge pressure of the third discharge port P3.
- a portion 16c2 and a pressure receiving portion 16c3 located on the same side as the spring 16c1 are provided.
- the control valve 16c includes the discharge pressure of the third discharge port P3 guided to the pressure receiving portion 16c2, the maximum load pressure of the actuators 3e, 3f, 3h or the actuators 3d to 3h guided to the pressure receiving portion 16c3, and the biasing force of the spring 16c1.
- the LS drive pressure is increased or decreased in the same manner as the control valve 16a.
- the control valve 16d is located on the same side as the spring 16d1 and a spring 16d1 that sets a target differential pressure for load sensing control, is positioned facing the spring 16d1, and is guided by the discharge pressure of the fourth discharge port P4. And a pressure receiving portion 16d3 located at the same position.
- the second communication control valve 15b is in the upper shut-off position in the figure, the maximum load pressure of the actuators 3d, 3g, 3h detected by the fourth shuttle valve group 8d is guided to the pressure receiving portion 16d3 of the control valve 16d.
- the control valve 16d includes a discharge pressure of the fourth discharge port P4 guided to the pressure receiving portion 16d2, a maximum load pressure of the actuators 3d, 3g, 3h or the actuators 3d to 3h guided to the pressure receiving portion 16d3, an urging force of the spring 16d1
- the LS drive pressure is increased or decreased in the same manner as the control valve 16a.
- the low pressure selection valve 21b selects the low pressure side of the LS drive pressure generated by the load sensing control valves 16c and 16d and outputs it to the load sensing control piston 17b.
- the load sensing control piston 17b changes the tilt angle of the swash plate of the second hydraulic pump 1b based on the LS drive pressure, and increases or decreases the discharge flow rates of the third and fourth discharge ports P3 and P4.
- FIG. 3 is a diagram showing a relationship between the LS driving pressure when the load sensing control pistons 17a and 17b are operated and the tilt angles of the swash plates of the first and second hydraulic pumps 1a and 1b.
- LS driving pressures acting on the load sensing control pistons 17a and 17b are indicated by Px1 and px2, and tilt angles of the swash plates of the first and second hydraulic pumps 1a and 1b are indicated by q1 and q2.
- the load sensing control piston 17a reduces the tilt angle q1 of the swash plate of the first hydraulic pump 1a to discharge the first and second discharge ports P1, P2.
- the tilt angle q1 of the swash plate of the first hydraulic pump 1a is increased to increase the discharge flow rates of the first and second discharge ports P1, P2.
- the first load sensing control unit 12a drives the discharge pressure on the high pressure side of the first and second discharge ports P1, P2 of the first hydraulic pump 1a by the discharge oil of the first and second discharge ports P1, P2.
- the tilt angle (capacity) of the swash plate of the first hydraulic pump 1a is controlled to be higher than the maximum load pressure of the actuators 3a to 3e by a predetermined pressure.
- K is the rate of change of the tilt angle q1 of the swash plate of the first hydraulic pump 1a with respect to the LS drive pressure Px1, and the spring constants of springs S3 and S4 described later and the tilt angle q2 of the second hydraulic pump 1b ( (Capacity).
- the load sensing control piston 17b also changes the tilt angle q2 of the swash plate of the second hydraulic pump 1b according to the increase / decrease of the LS driving pressure Px2, and the third and third of the second hydraulic pump 1b.
- the second pressure is set so that the discharge pressure on the high pressure side of the fourth discharge ports P3, P4 is higher than the maximum load pressure of the actuators 3d-3h driven by the discharge oil of the third and fourth discharge ports P3, P4 by a predetermined pressure.
- the tilt angle (capacity) of the swash plate of the hydraulic pump 1b is controlled.
- the target differential pressure of the load sensing control set by the springs 16a1 and 16b1 and the springs 16c1 and 16d1 is about 2 MPa, for example.
- the first torque control unit 13a includes a first torque control piston (first torque control actuator) 18a into which the discharge pressure of the first discharge port P1 is introduced, and a second discharge port P2.
- the second torque control unit 13b includes a third torque control piston (second torque control actuator) 18b into which the discharge pressure of the third discharge port P3 is introduced, and a fourth torque into which the discharge pressure of the fourth discharge port P4 is introduced.
- the first torque control unit 13a includes the discharge pressures of the third and fourth discharge ports P3 and P4 of the second hydraulic pump 1b and the LS drive pressure acting on the load sensing control piston 17b of the second load sensing control unit 12b.
- T2max second maximum torque
- the second hydraulic pump 1b is operated by the second torque control unit 13b.
- the second load sensing control unit 12b controls the capacity of the second hydraulic pump 1b without being limited by the control (when the pressure is lower than the starting pressure Pb of the constant absorption torque control of the second hydraulic pump 1b described later).
- the second oil is based on the discharge pressures and the LS drive pressures of the third and fourth discharge ports P3 and P4 so as to simulate the absorption torque of the second hydraulic pump 1b.
- the output pressure of the torque feedback circuit 30 that corrects the discharge pressure of the torque control piston (third torque control actuator) 31a and the fourth discharge port P4 of the second hydraulic pump 1b is led, and the output pressure increases as the output pressure increases.
- 1st reduction torque which reduces the inclination angle (capacity) of the swash plate of 1 hydraulic pump 1a, and reduces the maximum torque T1max set by spring S1, S2.
- the control piston and a (third torque control actuator) 31b The control piston and a (third torque control actuator) 31b.
- FIG. 4A is a torque control diagram of the first torque control unit 13a
- FIG. 4B is a torque control diagram of the second torque control unit 13b.
- the vertical axes are the tilt angles (capacities) q1 and q2, and if the vertical axes are replaced with the discharge flow rates Q1, Q2; Q3, Q4, these become horsepower control diagrams.
- the horizontal axis represents the pump discharge pressure.
- the average discharge pressure of the first and second discharge ports P1, P2 (P1p + P2p / 2)
- FIG. 4B the average discharge of the third and fourth discharge ports P3, P4. Pressure (P3p + P4p / 2).
- the average discharge pressure of the first and second discharge ports P1 and P2 increases.
- the first torque control unit 13a does not operate while the average discharge pressure is equal to or lower than the pressure at the start end of the characteristic line TP1a (torque control start pressure) Pa.
- the tilt angle (capacity) q1 of the swash plate of the first hydraulic pump 1a is controlled by the first load sensing control unit 12a without being restricted by the control of the first torque control unit 13a.
- the maximum tilt angle q1max of the first hydraulic pump 1a can be increased according to the operation amount (required flow rate).
- the first torque control unit 13a operates, and the average discharge pressure As the pressure increases, the absorption torque constant control (or constant horsepower control) is performed so as to reduce the maximum tilt angle (maximum capacity) of the first hydraulic pump 1a along the characteristic lines TP1a and TP1b. In this case, the first load sensing control unit 12a cannot increase the tilt angle of the first hydraulic pump 1a beyond the tilt angle defined by the characteristic lines TP1a and TP1b.
- the characteristic lines TP1a and TP1b are set to approximate a constant absorption torque curve (hyperbola) TP1 by two springs S1 and S2.
- the first torque control unit 13a performs a constant absorption torque control (or a constant horsepower control) so that the absorption torque of the first hydraulic pump 1a does not exceed the maximum torque T1max. I do.
- the maximum torque T1max is set to be slightly smaller than the rated output torque TER of the engine 2.
- the maximum torque T2max is set in the second torque controller 13b by the springs S3 and S4 regardless of the operating state of the first hydraulic pump 1a.
- TP2a and TP2b are characteristic lines of the springs S3 and S4 that set the maximum torque T1max.
- the second torque control unit 13b When the discharge oil of the second hydraulic pump 1b is supplied to any of the actuators 3d to 3h related to the second hydraulic pump 1b and the average discharge pressure of the third and fourth discharge ports P3 and P4 rises, this average discharge pressure Is equal to or lower than the pressure at the start of the characteristic line TP2a (torque control start pressure) Pb, the second torque control unit 13b does not operate.
- the tilt angle (capacity) q2 of the swash plate of the second hydraulic pump 1b is not limited by the control of the second torque control unit 13b, and is controlled by the second load sensing control unit 12b.
- the maximum tilt angle q2max of the second hydraulic pump 1b can be increased according to the operation amount (required flow rate).
- the second torque control unit 13b When the average discharge pressure of the third and fourth discharge ports P3 and P4 exceeds Pb with the swash plate of the second hydraulic pump 1b being at the maximum tilt angle q2max, the second torque control unit 13b operates and the average discharge pressure The absorption torque constant control is performed so that the maximum tilt angle (maximum capacity) of the second hydraulic pump 1b is reduced along the characteristic lines TP2a and TP2b as the pressure increases. In this case, the second load sensing control unit 12b cannot increase the tilt angle of the second hydraulic pump 1b beyond the tilt angle defined by the characteristic lines TP2a and TP2b.
- the characteristic lines TP2a and TP2b are set to approximate a constant absorption torque curve (hyperbola) TP2 by two springs S3 and S4.
- the second torque control unit 13b performs a constant absorption torque control (or a constant horsepower control) so that the absorption torque of the second hydraulic pump 1b does not exceed the maximum torque T2max. I do.
- the maximum torque T2max is smaller than the maximum torque T1max set in the first torque control unit 13a and is set to about 1 ⁇ 2 of the rated output torque TER of the engine 2.
- the torque feedback circuit 30 corrects and outputs the discharge pressures of the third and fourth discharge ports P3 and P4 of the second hydraulic pump 1b so as to simulate the absorption torque of the second hydraulic pump 1b.
- the first and second reduced torque control pistons 31a and 31b decrease the maximum torque T1max set in the first torque control unit 13a as the output pressure of the torque feedback circuit 30 increases.
- two arrows R1 and R2 indicate the effect that the first and second torque reduction control pistons 31a and 31b reduce the maximum torque T1max.
- the discharge pressure of the third and fourth discharge ports P3, P4 of the second hydraulic pump 1b increases, the absorption torque of the second hydraulic pump 1b at that time is T2 smaller than the maximum torque T2max, and the torque feedback circuit 30
- the torque feedback pistons 32a and 32b decrease the maximum torque T1max to T1max ⁇ T2s as indicated by an arrow R1 in FIG. 4A.
- the torque feedback pistons 32a and 32b are indicated by an arrow R2 in FIG. 4A. As shown, the maximum torque T1max is reduced to T1max ⁇ T2maxs.
- the maximum torque T1max set in the first torque control unit 13a is slightly smaller than the rated output torque TER of the engine 2 as described above, and the discharge oil of the second hydraulic pump 1b is not supplied to the actuators 3d to 3h.
- the first torque control unit 13a has the maximum absorption torque of the first hydraulic pump 1a.
- absorption torque constant control or constant horsepower control
- the absorption torque of the first hydraulic pump 1a is controlled so as not to exceed the rated output torque TER of the engine 2.
- the engine 2 can be prevented from being stopped (engine stall) while the rated output torque TER of the engine 2 is effectively used to the maximum.
- the torque is as described above.
- the feedback pistons 32a and 32b reduce the maximum torque T1max to T1max ⁇ T2s or T1max ⁇ T2maxs as indicated by an arrow X in FIG. 4A.
- the first hydraulic pump 1a and the second hydraulic pump can be used in the combined operation of simultaneously driving any one of the actuators 3a to 3e related to the first hydraulic pump 1a and any one of the actuators 3d to 3h related to the second hydraulic pump 1b.
- Total torque control is performed so that the total absorption torque of 1b does not exceed the rated output torque TER of the engine 2.
- the engine 2 is stopped while the rated output torque TER of the engine 2 is used to the maximum extent possible. (Engine stall) can be prevented.
- FIG. 1B is a diagram showing details of the torque feedback circuit 30.
- the torque feedback circuit 30 corrects and outputs the discharge pressure of the third discharge port P3 of the second hydraulic pump 1b so as to simulate the absorption torque of the second hydraulic pump 1b; and And a second torque feedback circuit unit 30b that corrects and outputs the discharge pressure of the fourth discharge port P4 of the second hydraulic pump 1b so as to simulate the absorption torque of the second hydraulic pump 1b.
- the first torque feedback circuit unit 30a has a first torque pressure reducing valve 32a to which the discharge pressure of the third discharge port P3 is guided as a primary pressure, and a target control pressure for setting a set pressure of the first torque pressure reducing valve 32a.
- a first voltage dividing circuit 33a for generating the first pressure reducing valve 32a when the discharge pressure of the third discharge port P3 is lower than the set pressure, the discharge pressure of the third discharge port P3 is directly changed to the secondary pressure.
- the discharge pressure of the third discharge port P3 is higher than the set pressure, the discharge pressure of the third discharge port P3 is reduced to the set pressure (target control pressure) and output, and the output pressure (two (Next pressure) is guided to the first reduced torque control piston 31a as the torque control pressure.
- the first voltage dividing circuit 33a includes a first partial pressure restrictor 34a through which the discharge pressure of the third discharge port P3 is guided, a first partial pressure valve 35a located on the downstream side of the first partial pressure restrictor 34a, A first relief valve (pressure limiting valve) connected to the first oil passage 36a between the pressure restricting portion 34a and the first pressure dividing valve 35a so as to prevent the pressure in the first oil passage 36a from exceeding the set pressure (relief pressure). 37a.
- the first partial pressure restrictor 34a is a fixed restrictor and has a certain opening area.
- the first partial pressure valve 35a is a variable throttle valve to which an LS driving pressure Px2 acting on the load sensing control piston 17b of the second load sensing control unit 12b is guided and whose opening area is changed according to the LS driving pressure Px2.
- the LS driving pressure Px2 is a tank pressure
- the opening area of the first voltage dividing valve 35a is zero (fully closed)
- the opening area of the first voltage dividing valve 35a increases.
- the opening area of the first partial pressure valve 35a becomes maximum (fully open).
- the target control pressure generated in the first oil passage 36a between the first partial pressure restrictor 34a and the first partial pressure valve 35a in accordance with the change in the opening area of the first partial pressure valve 35a is the first relief valve 37a.
- the pressure continuously changes from the set pressure to the tank pressure (zero), and the torque control pressure generated by the first torque pressure reducing valve 32a also changes continuously according to the change in the target control pressure.
- the set pressure of the first relief valve 37a is set equal to Pb in accordance with the torque control start pressure Pb (FIG. 4B) of the second torque control unit 13b.
- the second torque feedback circuit unit 30b is configured similarly to the first torque feedback circuit unit 30a. That is, the second torque feedback circuit unit 30b has a second torque pressure reducing valve 32b to which the discharge pressure of the fourth discharge port P4 is guided as a primary pressure, and a target control for setting the set pressure of the second torque pressure reducing valve 32b.
- a second voltage dividing circuit 33b for generating pressure, and the second torque pressure reducing valve 32b maintains the discharge pressure of the fourth discharge port P4 as it is when the discharge pressure of the fourth discharge port P4 is lower than the set pressure.
- the discharge pressure of the fourth discharge port P4 is reduced to the set pressure (target control pressure) and output. (Secondary pressure) is guided to the second reduced torque control piston 31b as the torque control pressure.
- the second voltage dividing circuit 33b includes a second partial pressure restrictor 34b through which the discharge pressure of the fourth discharge port P4 is guided, a second partial pressure valve 35b positioned downstream of the second partial pressure restrictor 34b, A second relief valve (pressure limiting valve) is connected to the second oil passage 36b between the pressure restricting portion 34b and the second pressure dividing valve 35b, and prevents the pressure in the second oil passage 36b from exceeding the set pressure (relief pressure). 37b.
- the second partial pressure restrictor 34b is a fixed restrictor and has a certain opening area.
- the second partial pressure valve 35b is a variable throttle valve to which an LS driving pressure Px2 acting on the load sensing control piston 17b of the second load sensing control unit 12b is guided and whose opening area is changed according to the LS driving pressure Px2.
- the LS driving pressure Px2 is a tank pressure
- the opening area of the second voltage dividing valve 35b is zero (fully closed)
- the opening area of the second voltage dividing valve 35b increases, and LS
- the drive pressure Px2 rises above a predetermined pressure, the opening area of the second partial pressure valve 35b becomes maximum (fully open).
- the target control pressure generated in the second oil passage 36b between the second partial pressure restrictor 34b and the second partial pressure valve 35b according to the change in the opening area of the second partial pressure valve 35b is the second relief valve 37b.
- the pressure continuously changes from the set pressure to the tank pressure (zero), and the torque control pressure generated by the second torque pressure reducing valve 32b also changes continuously according to the change in the target control pressure.
- the set pressure of the second relief valve 37b is set equal to Pb in accordance with the torque control start pressure Pb (FIG. 4B) of the second torque control unit 13b.
- FIG. 5A is a diagram showing the relationship between the LS drive pressure Px2 and the opening areas of the first and second voltage dividing valves 35a and 35b
- FIG. 5B is the opening area of the first and second voltage dividing valves 35a and 35b and the target.
- FIG. 5C is a diagram showing the relationship between the discharge pressure of the third and fourth discharge ports and the target control pressure when the LS drive pressure Px2 changes
- FIG. 5D is a diagram showing the relationship with the control pressure.
- FIG. 6 is a diagram showing the relationship between the discharge pressure of the third and fourth discharge ports and the torque control pressure when the LS drive pressure Px2 changes.
- AP3 and AP4 are opening areas of the first and second pressure dividing valves 35a and 35b
- P3tref and P4tref are target control pressures generated in the first and second oil passages 36a and 36b
- P3p and P4p Is the discharge pressure of the third and fourth discharge ports
- P3t and P4t are torque control pressures generated by the first and second torque pressure reducing valves 32a and 32b.
- the opening areas AP3 and AP4 of the first and second pressure dividing valves 35a and 35b are as follows.
- the opening areas AP3 and AP4 of the first and second pressure dividing valves 35a and 35b increase and the LS drive pressure Px2 increases to a predetermined pressure Px2a or more.
- the opening areas AP3 and AP4 of the first and second voltage dividing valves 35a and 35b are maximum APmax (fully open).
- the pressures of the first and second oil passages 36a and 36b are the third and second pressures. Equal to the discharge pressures P3p and P4p of the 4 discharge ports. However, the pressure in the first and second oil passages 36a and 36b cannot be higher than the set pressure of the first and second relief valves 37a and 37b.
- the target control pressures P3tref and P4tref decrease, and the first and second pressure dividing valves 35a and 35b.
- the target control pressures P3tref and P4tref become the tank pressure (zero).
- the target control pressures P3tref and P4tref Becomes equal to the discharge pressure of the third and fourth discharge ports.
- the target control pressures P3tref and P4tref also increase at the same value as the discharge pressures at the third and fourth discharge ports.
- the slope of the straight line representing the rate of increase of the target control pressures P3tref and P4tref at this time is 1.
- the target control pressures P3tref and P4tref are constant at the set pressures of the first and second relief valves 37a and 37b. It becomes.
- the first and second pressure dividing valves 35a and 35b open areas AP3 and AP4 increase accordingly, and the first and second discharge ports increase in discharge pressure.
- the target control pressures P3tref and P4tref increase at a smaller rate (small linear slope) than when the second partial pressure valves 35a and 35b open areas AP3 and AP4 are zero (fully closed).
- the rate of increase (inclination of the straight line) of the target control pressures P3tref and P4tref decreases, and the target control pressures P3tref and P4tref obtained with the same discharge pressures at the third and fourth discharge ports decrease.
- the first and second partial pressure valves 35a, 35b open areas AP3, AP4 become maximum APmax (fully open), and the target control pressures P3tref, P4tref become tank pressure (zero).
- the torque control pressures P3t and P4t are the same as the target control pressures P3tref and P4tref.
- the torque control pressures P3t and P4t are the same as the discharge pressures of the third and fourth discharge ports, and the torque control pressures P3t and P4t are increased as the LS drive pressure increases. And the torque control pressures P3t and P4t obtained with the same discharge pressures at the third and fourth discharge ports are reduced.
- the torque control pressures P3t and P4t generated by the torque feedback circuit units 30a and 30b as described above have characteristics that simulate the absorption torque of the second hydraulic pump 1b.
- the absorption torques ⁇ 3 and ⁇ 4 are It is calculated by the following formula.
- ⁇ 3 (P3p ⁇ q2) / 2 ⁇ (1)
- ⁇ 4 (P4p ⁇ q2) / 2 ⁇ (2)
- P3p and P4p are the discharge pressures of the third and fourth discharge ports P3 and P4, and q2 is the tilt angle of the second hydraulic pump 1b.
- the tilt angle of the second hydraulic pump 1b is controlled by the second load sensing control unit 12b.
- the swash plate of the second hydraulic pump 1b receives the LS drive pressure Px2 and the springs S3 and S4, and the tilt angle q2 is expressed by the following equation.
- K q2max-K x Px2 (3)
- K is a constant determined from the relationship between the spring constants of the springs S3 and S4 and the tilt angle q2 (capacity) of the second hydraulic pump 1b, and is a value corresponding to the slope K shown in FIG.
- the torque control pressures P3t and P4t have characteristics that simulate the absorption torque of the second hydraulic pump 1b
- the torque control pressures P3t and P4t are generated in the first and second reduced torque control pistons 31a and 31b by applying the torque control pressures P3t and P4t. It is necessary that the urging force to be a value proportional to the absorption torques ⁇ 3 and ⁇ 4 of the third and fourth discharge ports P3 and P4. For this purpose, the following relationship must be established.
- A is the pressure receiving area of the first and second torque reduction pistons 31a and 31b, and C is a proportionality constant.
- FIG. 6 is a diagram showing the relationship between the discharge pressures P3p, P4p, the torque control pressures P3t, P4t, and the LS drive pressure Px2 of the third and fourth discharge ports expressed by the expressions (6) and (7). .
- the torque control pressures P3t and P4t are the discharge pressures P3p and P3p of the third and fourth discharge ports, respectively. Same as P4p.
- the value of (1- (K ⁇ Px2 / D)) which is the slope of a straight line representing the rate of increase of the torque control pressures P3t and P4t, decreases, and the third and fourth values are the same.
- the torque control pressures P3t and P4t obtained by the discharge pressures P3p and P4p at the discharge port are lowered.
- the absorption torque constant control (or constant horsepower control) of the second torque control unit 13b starts, and the second hydraulic pump
- the absorption torque of 1b is constant. Therefore, the torque control pressures P3t and P4t may be made constant at the torque control start pressure Pb.
- the rate of increase of the torque control pressures P3t and P4t when the discharge pressures P3p and P4p of the third and fourth discharge ports shown in FIG. Is equal to the increase rate (straight line slope) of the torque control pressures P3t and P4t when the discharge pressures P3p and P4p of the third and fourth discharge ports shown in FIG.
- the torque control pressure P3t, P4t reaches the torque control start pressure Pb that is the set pressure of the first and second relief valves 37a, 37b, the set pressure (Pb) is constant. It becomes.
- the torque control pressures P3t and P4t generated by the torque feedback circuit units 30a and 30b have characteristics simulating the absorption torque of the second hydraulic pump 1b, and the torque feedback circuit units 30a and 30b have the second hydraulic pressure.
- the pump 1b is controlled by the second torque control unit 13b and operates at the maximum torque T2max (second maximum torque)
- the second hydraulic pump 1b is not limited by the control of the second torque control unit 13b.
- the absorption of the main pump 202 It has a function of correcting and outputting the discharge pressure of the main pump 202 so as to simulate torque.
- Fig. 7 shows the external appearance of the hydraulic excavator.
- the hydraulic excavator includes an upper swing body 300, a lower traveling body 301, and a front work machine 302, and the upper swing body 300 is mounted on the lower travel body 301 so as to be rotatable.
- the upper swing body 300 is connected to the tip portion of the upper swing body 300 via a swing post 303 so as to be rotatable in the vertical and horizontal directions.
- the lower traveling body 301 includes left and right crawler belts 310 and 311, and a soil removal blade 305 that can move up and down in front of the track frame 304.
- the upper swing body 300 includes a cabin (operator's cab) 300a.
- a front work machine and operating lever devices 309a and 309b for turning (only one is shown) and operating lever / pedal devices 309c and 309d for traveling (one side) Operation means such as only shown) are provided.
- the front work machine 302 is configured by pin-coupling a boom 306, an arm 307, and a bucket 308.
- the upper swing body 300 is driven to rotate by the swing motor 3c with respect to the lower traveling body 301, and the front work machine 302 rotates in the horizontal direction by rotating the swing post 303 by the swing cylinder 3f (see FIG. 1A).
- the left and right crawler belts 310 and 311 of the lower traveling body 301 are rotationally driven by the left and right traveling motors 3d and 3e, and the blade 305 is vertically driven by the blade cylinder 3g.
- the boom 306, the arm 307, and the bucket 308 rotate in the vertical direction by expanding and contracting the boom cylinder 3h, the arm cylinder 3a, and the bucket cylinder 3b, respectively.
- the flow control valve 6b or the flow control valve 6d is switched by operating the respective operation lever, and the discharge port P1 or P2 on one side is switched. Discharged oil is supplied to the bucket cylinder 3b or the swing motor 3c. Also at this time, the discharge flow rates of the first and second discharge ports P1, P2 are controlled by the load sensing control of the first load sensing control unit 12a and the constant absorption torque control of the first torque control unit 13a. The discharge oil from the discharge port P2 or P1 on the side where pressure oil is not supplied to the bucket cylinder 3b or the swing motor 3c is returned to the tank via the unload valve 10b or 10a.
- the flow control valve 6i or the flow control valve 6k is switched by operating each operation lever, and the discharge port P3 or P4 on one side is switched. Is supplied to the swing cylinder 3f or the blade cylinder 3g. Also at this time, the discharge flow rates of the third and fourth discharge ports P3 and P4 are controlled by the load sensing control of the second load sensing control unit 12b and the constant absorption torque control of the second torque control unit 13b. The discharge oil from the discharge port P4 or P3 on the side where pressure oil is not supplied to the swing cylinder 3f or the blade cylinder 3g is returned to the tank via the unload valve 10d or 10c.
- the constant absorption torque control of the first torque controller 13a the total torque control shown in FIG. 4A is performed.
- the discharge flow rate of the first and second discharge ports P1, P2 and the discharge flow rate of the third and fourth discharge ports P3, P4 are controlled.
- the total torque control shown in FIG. 4A is performed.
- the oil discharged from the first discharge port P1 on the side where the flow control valves 6a to 6c are closed is returned to the tank via the unload valve 10a.
- the excess flow rate of the discharge oil at the discharge port on the side where the required flow rate is low or the discharge oil at the discharge port on the side where the flow rate control valve is closed is returned to the tank via the unload valve.
- the load pressure (maximum load pressure) of the actuator on the first discharge port P1 side detected by the first shuttle valve group 208a is guided to the pressure compensation valves 7a to 7c and the first unload valve 210a, and the second shuttle.
- the load pressure (maximum load pressure) of the actuator on the second discharge port P2 side detected by the valve group 208b is guided to the pressure compensation valves 7d to 7f and the second unload valve 210b, and the first discharge port P1 side and the second discharge pressure are detected.
- the pressure compensation valve and the unload valve are controlled separately on the discharge port P2 side.
- the pressure rise at the discharge port is restricted based on the low load pressure by the unload valve on the discharge port side.
- the pressure loss of the unloading valve when returning to the state is reduced, and operation with less energy loss becomes possible.
- the first and second discharge ports P1, P and the third and fourth Even if there is a difference in the discharge flow rates at the discharge ports P3 and P4, the supply flow rate of the travel motor 3d on the left side and the supply flow rate of the travel motor 3e on the right side are the same, and the vehicle body does not meander and can travel straight. it can.
- the traveling The supply flow rate to the left travel motor 3d and the supply flow rate to the right travel motor 3e are as follows.
- the flow rate control valves 6f and 6j and the flow rate control valves 6c and 6g and the flow rate control valves 6a and 6e are switched.
- the first communication control valve 215a is switched to the lower communication position in the figure.
- the oil discharged from the first and second discharge ports P1 and P2 is supplied from the first hydraulic pump 1a side to the travel motor 3d on the left side, and discharged from the second hydraulic pump 1b side to the fourth discharge port P4.
- Oil is supplied, and the oil discharged from the first and second discharge ports P1 and P2 is supplied from the first hydraulic pump 1a side to the travel motor 3e on the right side of the travel, and is supplied from the second hydraulic pump 1b side to the third discharge port. P3 discharge oil is supplied. The remaining pressure oil supplied to the traveling motors 3d and 3e of the first and second discharge ports P1 and P2 is supplied to the arm cylinder 3a.
- the actuators 3a to 3 detected by the first and second shuttle valve groups 208a and 208b since the first communication control valve 215a switches to the lower communication position in the figure, the actuators 3a to 3 detected by the first and second shuttle valve groups 208a and 208b.
- the maximum load pressure of 3e is the load sensing control valves 216a and 216b, the pressure compensation valves 7a to 7c and 7d to 7f, and the first unload valves 210a and 210b. Then, load sensing control and pressure compensation valve and unload valve control are performed.
- the maximum load pressure is separately applied to the third discharge port P3 side and the fourth discharge port P4 side.
- the left traveling motor 3d is supplied with the discharge oil from the first and second discharge ports P1 and P2 from the first hydraulic pump 1a side, and supplied from the second hydraulic pump 1b side to the fourth discharge port P4.
- the oil discharged from the first and second discharge ports P1, P2 is supplied to the right traveling motor 3e from the first hydraulic pump 1a side, and discharged from the second hydraulic pump 1b side to the third discharge port P3. Oil is supplied. Accordingly, even in the traveling combined operation, the supply flow rate of the travel left travel motor 3d and the supply flow rate of the travel right travel motor 3e become the same, and the vehicle body can travel straight without being meandering.
- the discharge flow rate of the first discharge port P1 is Q1
- the discharge flow rate of the second discharge port P2 is Q2
- the discharge flow rate of the third discharge port P3 is Q3
- the discharge flow rate of the fourth discharge port P4 is Q4
- the flow rate of pressure oil supplied to the travel motor 3d is Qd
- the flow rate of pressure oil supplied to the travel motor 3e on the right side is Qe
- the flow rate of pressure oil supplied to the boom cylinder 3a which is an actuator other than the travel motor.
- the flow rate Qa of the pressure oil supplied to the boom cylinder 3a is subtracted from the combined flow rate Q1 + Q2 of the discharge oil from the first and second discharge ports P1, P2 from the first hydraulic pump 1a side to the left and right traveling motors 3d, 3e. Further, half of Q1 + Q2-Qa is supplied. The reason why it becomes 1/2 of Q1 + Q2-Qa is that the stroke amount (opening area) of the flow control valve 6f and the stroke amount (opening area-required flow rate) of the flow control valve 6c are the same.
- the operation example of the traveling combined operation is a case where the traveling motors 3d and 3e and the arm cylinder 3a are driven simultaneously.
- an actuator bucket cylinder 3b, swing motor 3c driven by pressure oil discharged only from the first discharge port P1 or the second discharge port P2 of the first hydraulic pump 1a, or
- actuators tilt cylinder 3f, blade cylinder 3g driven by pressure oil discharged only from the third discharge port P3 or the fourth discharge port P4 of the second hydraulic pump 1b.
- the vehicle body does not meander and can travel straight.
- the first to fourth shuttle valve groups 208a to 208d, the first and second communication control valves 15a and 15b, the load sensing control valves 216a to 216d, and the low pressure selection valves 221a and 221b are provided.
- the first and second communication control valves 15a and 15b are configured to communicate and block both the discharge port and the output oil path of the maximum load pressure
- the first and second communication control valves 15a and 15b communicate the discharge port and The circuit configuration other than that may be the same as that of the first embodiment. Even in this case, the first and second communication control valves 15a and 15b are switched to the communication position at the time of the traveling combined operation, so that the effect of ensuring the straight traveling performance can be obtained.
- FIG. 8 shows, as a comparative example, a hydraulic system in the case where the full-torque control technique described in Patent Document 2 is incorporated in the two-pump load sensing system including the first and second hydraulic pumps 1a and 1b shown in FIG. FIG. In the figure, members equivalent to those shown in FIG.
- the pressure reducing valves 41a and 41b reduce and output the discharge pressures of the third and fourth discharge ports of the second hydraulic pump 1b so that the secondary pressure (torque control pressure) does not exceed the set pressure.
- the set pressure of the pressure reducing valves 41a and 41b is a value corresponding to the maximum torque T2max set by the springs S3 and S4 of the torque control unit of the second hydraulic pump 1b (starting pressure Pb of constant absorption torque control shown in FIG. 4B). It is set as follows.
- FIG. 9 is a diagram showing the total torque control of the comparative example shown in FIG.
- the second hydraulic pump 1b when the discharge pressures of the third and fourth discharge ports of the second hydraulic pump are equal to or higher than the start pressure of the constant absorption torque control, the second hydraulic pump 1b is under the constant absorption torque control.
- the pressure reducing valves 41a and 41b reduce the discharge pressures of the third and fourth discharge ports of the second hydraulic pump to a pressure corresponding to the maximum torque T2max to reduce the torque control piston 31a of the first hydraulic pump 1a.
- the total torque control is performed on the first hydraulic pump 1a side by reducing the maximum torque from T1max by T2max.
- the second hydraulic pump 1b is not under the constant absorption torque control, and the second In some cases, the hydraulic pump 1b is controlled to a tilt angle smaller than the tilt limited by the constant absorption torque control by the load sensing control. In this case, the absorption torque of the second hydraulic pump 1b assumed at the pressure corresponding to the maximum torque T2max is larger than the actual absorption torque of the second hydraulic pump 1b.
- FIG. 10 is a diagram showing total torque control according to the present embodiment.
- the torque feedback circuit 30 includes the second hydraulic pump when the second hydraulic pump 1b operates under the maximum torque T2max (second maximum torque) under the control of the second torque control unit 13b.
- T2max second maximum torque
- the second load sensing control unit 12b controls the capacity of the second hydraulic pump 1b (starting pressure Pb of the constant absorption torque control of the second hydraulic pump 1b)
- the discharge pressures of the third and fourth discharge ports P3 and P4 of the second hydraulic pump 1b are corrected and output so as to simulate the absorption torque of the second hydraulic pump 1b.
- the first and second reduced torque control pistons 31a and 31b are set to the first torque control unit 13a as the output pressure of the torque feedback circuit 30 increases. Reduce the torque T1max.
- the discharge pressure of the third and fourth discharge ports P3 and P4 of the second hydraulic pump 1b increases, and the absorption torque of the second hydraulic pump 1b at that time is T2 which is smaller than the maximum torque T2max.
- the absorption torque simulated by the torque feedback circuit 30 is T2s ( ⁇ T2)
- the torque feedback pistons 32a and 32b reduce the maximum torque T1max to T1max-T2s as shown by the arrows in FIG. Total torque control is performed with this maximum torque T1max-T2s.
- the maximum torque does not decrease more than necessary, and the engine 2 can be prevented from being stopped (engine stall) while the rated output torque TER of the engine 2 is effectively used to the maximum.
- the absorption torque of the second hydraulic pump 1b can be accurately detected with a pure hydraulic configuration (torque feedback circuit 30), and the absorption torque is detected with the first hydraulic pump.
- torque feedback circuit 30 torque feedback circuit 30
- the total torque control can be performed with high accuracy, and the rated output torque TER of the prime mover 2 can be used effectively.
- the first pump control device 5a can be reduced in size, and the mountability of the hydraulic pump including the pump control device is improved. As a result, it is possible to provide an energy efficient, low fuel consumption and practical construction machine.
- the first and second partial pressure restrictors (fixed restrictors) 34a and 34b and the first and second partial pressure valves (variable restrictors) 35a and 35b are connected to each other.
- the target control pressure formed in the second oil passages 36a and 36b and the torque control pressure output by the first and second pressure reducing valves 32a and 32b are the same value, and the first and second oil passages 36a, It is also possible to use the pressure formed in 36b directly as the torque control pressure.
- the first and second The partial pressure restrictors (fixed restrictors) 34a and 34b become resistors and it is difficult to supply a sufficient amount of pressure oil to the third torque control actuators 32a and 32b, and the responsiveness of the third torque control actuators 32a and 32b is low. It can get worse.
- the first and second oil passages 36a between the first and second partial pressure restrictors (fixed restrictors) 34a and 34b and the first and second partial pressure valves (variable restrictors) 35a and 35b. , 36b is set as the target control pressure to the first and second pressure reducing valves 32a, 32b to set the set pressure of the first and second pressure reducing valves 32a, 32b, and the first pressure is determined from the discharge pressure of the second hydraulic pump 1b. Since the torque control pressure is generated by the second pressure reducing valves 32a and 32b, a flow rate when the third torque control actuators 32a and 32b are driven by the torque control pressure is secured, and the third torque control actuators 32a and 32b are secured. Responsiveness when driving can be improved.
- the pressures of the first and second oil passages 36a and 36b between the first and second partial pressure restrictors (fixed restrictors) 34a and 34b and the first and twenty-second partial pressure valves (variable restrictors) 35a and 35b. are not directly used as torque control pressures, the first and second partial pressure restrictors (fixed restrictors) 34a and 34b and the first and twenty second partial pressure valves (variable restrictor) 35a for obtaining a necessary target control pressure.
- 35b and the responsiveness of the third torque control actuators 32a, 32b can be independently set, and the torque feedback circuit 30 can be easily and accurately set to exhibit necessary performance. .
- the discharge pressure of the second hydraulic pump 1b is higher than the set pressure of the first and second pressure reducing valves 32a and 32b, the discharge pressure fluctuation of the second hydraulic pump 1b is changed to the first and second pressure reducing valves 32a and 32b. And the third torque control actuators 32a and 32b are not affected, so that the stability of the system is ensured.
- first and second hydraulic pumps are split flow type hydraulic pumps having the first and second discharge ports P1, P2 and the third and fourth discharge ports P3, P4
- both or one of the first and second hydraulic pumps may be a single flow type hydraulic pump having a single discharge port.
- the first and second hydraulic pumps are single flow type hydraulic pumps, it is only necessary to have one each of the torque feedback circuit 30 and one torque reduction control piston to which the torque control pressure is guided.
- 4A and 4B represents the pressure of a single discharge port (discharge pressure of the hydraulic pump).
- the first and second partial pressure restrictors (fixed restrictors) 34a and 34b and the first and second partial pressure valves (variable restrictors) 35a and 35b are connected to each other. Since the target control pressure formed in the first and second oil passages 36a and 36b and the torque control pressure output by the first and second pressure reducing valves 32a and 32b are the same value, the first and second oil passages The pressure formed in 36a and 36b may be directly guided to the reduced torque control actuators 31a and 31b as the torque control pressure.
- the first and second partial pressure restrictors (fixed restrictors) 34a and 34b and the first and second partial pressure valves (variable restrictors) 35a and 35b are provided.
- the first and second relief valves 37a and 37b are provided so that the pressure in the first and second oil passages 36a and 36b does not become the set pressure (torque start pressure Pb) or more
- a pressure reducing valve is used instead of the relief valve. Also good.
- the same function can be obtained by setting the set pressure of the pressure reducing valve to the torque start pressure Pb and using the output pressure of the pressure reducing valve as the target control pressures P35ref and P4tref.
- the 1st pump control apparatus 5a shall have the 1st load sensing control part 12a and the 1st torque control part 18a
- the 1st load sensing control part 12a in the 1st pump control apparatus 5a is not essential
- Other control methods such as so-called positive control or negative control can be used as long as the capacity of the first hydraulic pump can be controlled according to the operation amount of the operation lever (the opening area of the flow control valve ⁇ the required flow rate). May be.
- the load sensing system of the above embodiment is an example, and the load sensing system can be variously modified.
- a differential pressure reducing valve that outputs the pump discharge pressure and the maximum load pressure as absolute pressure is provided, the output pressure is guided to the pressure compensation valve, the target compensation differential pressure is set, and the LS control valve is provided.
- the target differential pressure for load sensing control is set, the pump discharge pressure and the maximum load pressure may be guided to the pressure control valve and the LS control valve through separate oil passages.
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Abstract
Description
図1A、図1B及び図2は、本発明の第1の実施の形態に係わる油圧ショベル(建設機械)の油圧駆動装置を示す図であり、図1Aは油圧駆動装置全体を示す油圧回路図、図2は油圧駆動装置全体を示すブロック図である。図1Bは、図1A及び図2に示すトルクフィードバック回路の詳細を示す油圧回路図である。
τ4=(P4p×q2)/2π…(2)
前述したように、P3p,P4pは第3及び第4吐出ポートP3,P4の吐出圧であり、q2は第2油圧ポンプ1bのの傾転角である。
ここで、KはバネS3,S4のバネ定数と第2油圧ポンプ1bの傾転角q2(容量)の関係から決定される定数であり、図3に示した傾きKに相当する値である。
τ4=C(A×P4t)…(5)
ここで、Aは第1及び第2減トルク制御ピストン31a,31bの受圧面積であり、Cは比例定数である。
τ4=(P4p×(q2max-K×Px2))/2π=C(A×P4t)
変形すると、次の式になる。
P4t=((P4p×(q2max-K×Px2))/2π)/C×A
D=2π/C×Aと置き換えると、次の式になる。
P4t=D(P4p×(q2max-K×Px2))
D×q2maxが1となるようにAとCの値を設定すると、次の式になる。
P4t=P4p×(1-(K×Px2/D))…(7)
図6は、(6)式及び(7)式で表される第3及び第4吐出ポートの吐出圧P3p,P4pとトルク制御圧力P3t,P4tとLS駆動圧力Px2との関係を示す図である。
~動作~
次に、本実施の形態の動作を説明する。
<<第1油圧ポンプ1a側アクチュエータの単独駆動>>
第1油圧ポンプ1a側に接続されたアクチュエータの1つ、例えばアームシリンダ3aを単独で駆動してアーム動作を行うときは、アーム用の操作レバーを操作すると流量制御弁6a,6eが切り換わり、アームシリンダ3aに第1及び第2吐出ポートP1,P2の吐出油が合流して供給される。また、このとき前述したように、第1ロードセンシング制御部12aのロードセンシング制御と第1トルク制御部13aの吸収トルク一定制御により第1及び第2吐出ポートP1,P2の吐出流量が制御される。
第2油圧ポンプ1b側に接続されたアクチュエータの1つ、例えばブームシリンダ3hを単独で駆動してブーム動作を行うときは、ブーム用の操作レバーを操作すると流量制御弁6h,6lが切り換わり、ブームシリンダ3hに第3及び第4吐出ポートP3,P4の吐出油が合流して供給される。また、このとき前述したように、第2ロードセンシング制御部12bのロードセンシング制御と第2トルク制御部13bの吸収トルク一定制御により第3及び第4吐出ポートP3,P4の吐出流量が制御される。
<<アームシリンダとブームシリンダの同時駆動>>
アームシリンダ3aとブームシリンダ3hを同時に駆動してアーム307とブーム306の複合動作を行うときは、アーム用の操作レバーとブーム用の操作レバーを操作すると流量制御弁6a,6eと流量制御弁6h,6lが切り換わり、アームシリンダ3aに第1及び第2吐出ポートP1,P2の吐出油が合流して供給され、ブームシリンダ3hに第3及び第4吐出ポートP3,P4の吐出油が合流して供給される。また、第1油圧ポンプ1a側と第2油圧ポンプ1b側のそれぞれで、前述したように、第1及び第2ロードセンシング制御部12a,12bのロードセンシング制御と第1及び第2トルク制御部13a,13bの吸収トルク一定制御により、第1及び第2吐出ポートP1,P2の吐出流量と第3及び第4吐出ポートP3,P4の吐出流量が制御される。また、第1トルク制御部13aの吸収トルク一定制御では、図4Aに示した全トルク制御が行われる。
旋回モータ3cとブームシリンダ3hとを同時に駆動して上部旋回体300(旋回)とブーム306の複合動作を行うときは、旋回用の操作レバーとブーム用の操作レバーを操作するとと流量制御弁6dと流量制御弁6h,6lが切り換わり、旋回モータ3cに第2吐出ポートP2の吐出油が供給され、ブームシリンダ3hに第3及び第4吐出ポートP3,P4の吐出油が合流して供給される。また、第1油圧ポンプ1a側と第2油圧ポンプ1b側のそれぞれで、前述したように、第1及び第2ロードセンシング制御部12a,12bのロードセンシング制御と第1及び第2トルク制御部13a,13bの吸収トルク一定制御により、第1及び第2吐出ポートP1,P2の吐出流量と第3及び第4吐出ポートP3,P4の吐出流量が制御される。また、第1トルク制御部13aの吸収トルク一定制御では、図4Aに示した全トルク制御が行われる。流量制御弁6a~6cが閉じられている側の第1吐出ポートP1の吐出油はアンロード弁10aを介してタンクに戻される。
第1油圧ポンプ1aの第1及び第2吐出ポートP1,P2のみに接続されるアクチュエータ(アームシリンダ3a、バケットシリンダ3b、旋回モータ3c)の少なくとも1つと、第2油圧ポンプ1bの第3及び第4吐出ポートP3,P4のみに接続されるアクチュエータ(スイングシリンダ3f、ブレードシリンダ3g、ブームシリンダ3h)の少なくとも1つを同時に駆動する上記以外の複合動作においても、上記と同様に、ロードセンシング制御と吸収トルク一定制御により、第1及び第2吐出ポートP1,P2の吐出流量と第3及び第4吐出ポートP3,P4の吐出流量が制御され、第1トルク制御部13aの吸収トルク一定制御では、図4Aに示した全トルク制御が行われる。流量制御弁が閉じられている側の吐出ポートの吐出油は対応するアンロード弁を介してタンクに戻される。
第1油圧ポンプ1aの第1吐出ポートP1に接続されるアクチュエータ(アームシリンダ3a、バケットシリンダ3b、走行右の走行モータ3e)の少なくとも1つと、第1油圧ポンプ1aの第2吐出ポートP2に接続されるアクチュエータ(アームシリンダ3a,旋回モータ3c、走行左の走行モータ3d)の少なくとも1つを同時に駆動する複合動作では、アームシリンダ3aを単独で駆動するアーム動作の場合と同様、第1ロードセンシング制御部12aのロードセンシング制御と第1トルク制御部13aの吸収トルク一定制御により第1及び第2吐出ポートP1,P2の吐出流量が制御される。また、要求流量の少ない側の吐出ポートの吐出油の余剰流量或いは流量制御弁が閉じられている側の吐出ポートの吐出油はアンロード弁を介してタンクに戻される。このとき、第1シャトル弁群208aによって検出された第1吐出ポートP1側のアクチュエータの負荷圧(最高負荷圧)が圧力補償弁7a~7cと第1アンロード弁210aに導かれ、第2シャトル弁群208bによって検出された第2吐出ポートP2側のアクチュエータの負荷圧(最高負荷圧)が圧力補償弁7d~7fと第2アンロード弁210bに導かれ、第1吐出ポートP1側と第2吐出ポートP2側とで別々に圧力補償弁とアンロード弁の制御が行われる。これにより低負荷圧側の吐出ポートの余剰流量がタンクに戻るとき、その吐出ポートの圧力は当該吐出ポート側のアンロード弁によって低い負荷圧に基づいて圧力上昇が制限されるため、余剰流量がタンクに戻るときのアンロード弁の圧損が低減し、エネルギーロスの少ない運転が可能となる。
第2油圧ポンプ1b側の2つのアクチュエータを同時に駆動する複合動作においても、上述した第1油圧ポンプ1a側の2つのアクチュエータを同時に駆動する複合動作の場合と同様、第2ロードセンシング制御部12bのロードセンシング制御と第2トルク制御部13bの吸収トルク一定制御により第3及び第4吐出ポートP3,P4の吐出流量が制御される。また、要求流量の少ない側の吐出ポートの吐出油の余剰流量或いは流量制御弁が閉じられている側の吐出ポートの吐出油はアンロード弁を介してタンクに戻され、このときのアンロード弁の圧損が低減し、エネルギーロスの少ない運転が可能となる。
走行左の走行モータ3dと走行右の走行モータ3eを駆動して走行動作を行うときは、左右の走行用操作レバー或いはペダルを操作すると流量制御弁6f,6jと流量制御弁6c,6gが切り換わり、走行左の走行モータ3dに第1油圧ポンプ1aの第2吐出ポートP2の吐出油と第2油圧ポンプ1bの第4吐出ポートP4の吐出油が合流して供給され、走行右の走行モータ3eに第1油圧ポンプ1aの第1吐出ポートP1の吐出油と第2油圧ポンプ1bの第3吐出ポートP3の吐出油が合流して供給される。このため、仮に、第1油圧ポンプ1aの斜板の傾転角と第2油圧ポンプ1bの斜板の傾転角が相違し、第1及び第2吐出ポートP1,Pと第3及び第4吐出ポートP3,P4で吐出流量の相違が発生したとしても、走行左の走行モータ3dの供給流量と走行右の走行モータ3eの供給流量は同じとなり、車体は蛇行せず、直進走行することができる。
走行右の供給流量:Q1+Q3
ここで、Q1=Q2(同一斜板のため)、Q3=Q4(同一斜板のため)の関係にある。したがって、仮にQ1=Q2≠Q3=Q4となったとしても、
Q2+Q4=Q1+Q3
の関係は成り立ち、走行左の走行モータ3dの供給流量と走行右の走行モータ3eの供給流量は同じとなる。
走行モータ3d,3eと他のアクチュエータの少なくとも1つ、例えばアームシリンダ3aとを同時に駆動する走行複合動作を行う場合について説明する。
に導かれ、ロードセンシング制御と圧力補償弁及びアンロード弁の制御が行われる。一方、第2油圧ポンプ1b側においては、第2連通制御弁215bは図示上側の遮断位置に保持されているため、第3吐出ポートP3側と第4吐出ポートP4側とで別々に最高負荷圧が検出され、それぞれの最高負荷圧が対応するロードセンシング制御弁216c,216dと圧力補償弁7g~7i,7j~7m及び第3及び第4アンロード弁210c,210dに導かれ、ロードセンシング制御と圧力補償弁及びアンロード弁の制御が行われる。
走行左の供給流量Qe=(Q1+Q2-Qa)/2+(Q3+Q4)/2
すなわち、Qd=Qeであり、車体は蛇行せず、直進走行することができる。
次に、本実施の形態により得られる効果について説明する。
以上の実施の形態では、第1及び第2油圧ポンプが第1及び第2吐出ポートP1,P2及び第3及び第4吐出ポートP3,P4を有するスプリットフロータイプの油圧ポンプである場合について説明したが、第1及び第2油圧ポンプの両方或いは一方は単一の吐出ポートを有するシングルフロータイプの油圧ポンプであってもよい。第1及び第2油圧ポンプがシングルフロータイプの油圧ポンプである場合、トルクフィードバック回路30の回路部とトルク制御圧力が導かれる減トルク制御ピストンはそれぞれ1つづつあればよい。また、図4A及び図4Bの横軸は単一の吐出ポートの圧力(油圧ポンプの吐出圧)となる。
1b 第2油圧ポンプ
2 原動機(ディーゼルエンジン)
3a~3h アクチュエータ
3a アームシリンダ
3d 走行左の走行モータ
3e 走行右の走行モータ
3h ブームシリンダ
4 コントロールバルブ
5a 第1ポンプ制御装置
5b 第2ポンプ制御装置
6a~6m 流量制御弁
7a~7m 圧力補償弁
8a 第1シャトル弁群
8b 第2シャトル弁群
8c 第3シャトル弁群
8d 第4シャトル弁群
9a~9d バネ
10a~10d アンロード弁
12a 第1ロードセンシング制御部
12b 第2ロードセンシング制御部
13a 第1トルク制御部
13b 第2トルク制御部
15a 第1連通制御弁
15b 第2連通制御弁
16a~16d ロードセンシング制御弁
17a,17b ロードセンシング制御ピストン(ロードセンシング制御アクチュエータ)
18a 第1トルク制御ピストン(第1トルク制御アクチュエータ)
19a 第2トルク制御ピストン(第1トルク制御アクチュエータ)
18b 第3トルク制御ピストン(第2トルク制御アクチュエータ)
19b 第4トルク制御ピストン(第2トルク制御アクチュエータ)
21a,21b 低圧選択弁
30 トルクフィードバック回路
30a 第1トルクフィードバック回路部
30b 第2トルクフィードバック回路部
31a 第1減トルク制御ピストン(第3トルク制御アクチュエータ)
31b 第2減トルク制御ピストン(第3トルク制御アクチュエータ)
32a 第1トルク減圧弁
32b 第2トルク減圧弁
33a 第1分圧回路
33b 第2分圧回路
34a 第1分圧絞り部
34b 第2分圧絞り部
35a 第1分圧弁
35b 第1分圧弁
36a 第1油路
36b 第2油路
37a 第1リリーフ弁(圧力制限弁)
37b 第2リリーフ弁(圧力制限弁)
P1,P2 第1及び第2吐出ポート
P3,P4 第3及び第4吐出ポート
S1,S2 バネ
S3,S4 バネ
Claims (3)
- 原動機と、
前記原動機により駆動される可変容量型の第1油圧ポンプと、
前記原動機により駆動される可変容量型の第2油圧ポンプと、
前記第1及び第2油圧ポンプにより吐出された圧油により駆動される複数のアクチュエータと、
前記第1及び第2油圧ポンプから前記複数のアクチュエータに供給される圧油の流量を制御する複数の流量制御弁と、
前記複数の流量制御弁の前後差圧をそれぞれ制御する複数の圧力補償弁と、
前記第1油圧ポンプの吐出流量を制御する第1ポンプ制御装置と、
前記第2油圧ポンプの吐出流量を制御する第2ポンプ制御装置とを備え、
前記第1ポンプ制御装置は、
前記第1油圧ポンプの吐出圧と容量の少なくとも一方が増大し、前記第1油圧ポンプの吸収トルクが増大するとき、前記第1油圧ポンプの吸収トルクが第1最大トルクを超えないように前記第1油圧ポンプの容量を制御する第1トルク制御部を有し、
前記第2ポンプ制御装置は、
前記第2油圧ポンプの吐出圧と容量の少なくとも一方が増大し、前記第2油圧ポンプの吸収トルクが増大するとき、前記第2油圧ポンプの吸収トルクが第2最大トルクを超えないように前記第2油圧ポンプの容量を制御する第2トルク制御部と、
前記第2油圧ポンプの吸収トルクが前記第2最大トルクよりも小さいとき、前記第2油圧ポンプの吐出圧が前記第2油圧ポンプにより吐出された圧油により駆動されるアクチュエータの最高負荷圧より目標差圧だけ高くなるよう前記第2油圧ポンプの容量を制御するロードセンシング制御部とを有する建設機械の油圧駆動装置において、
前記第1トルク制御部は、前記第1油圧ポンプの吐出圧が導かれ、前記吐出圧の上昇時に前記第2油圧ポンプの容量を減少させ吸収トルクが減少するよう前記第1油圧ポンプの容量を制御する第1トルク制御アクチュエータと、前記第1最大トルクを設定する第1付勢手段とを有し、
前記第2トルク制御部は、前記第2油圧ポンプの吐出圧が導かれ、前記吐出圧の上昇時に前記第2油圧ポンプの容量を減少させ吸収トルクが減少するよう前記第2油圧ポンプの容量を制御する第2トルク制御アクチュエータと、前記第2最大トルクを設定する第2付勢手段とを有し、
前記ロードセンシング制御部は、
前記第2油圧ポンプの吐出圧と前記最高負荷圧との差圧が前記目標差圧よりも小さくなるにしたがって低くなるようロードセンシング駆動圧力を変化させる制御弁と、前記ロードセンシング駆動圧力が低くなるにしたがって前記第2油圧ポンプの容量を増加し吐出流量が増加するよう前記第2油圧ポンプの容量を制御するロードセンシング制御アクチュエータとを有し、
前記第1ポンプ制御装置は、更に、
前記第2油圧ポンプの吐出圧と前記ロードセンシング駆動圧力とが導かれ、前記第2油圧ポンプが前記第2トルク制御部の制御の制限を受け、前記第2最大トルクで動作するときと、前記第2油圧ポンプが前記第2トルク制御部の制御の制限を受けず、前記ロードセンシング制御部が前記第2油圧ポンプの容量を制御するときのいずれの場合にも前記第2油圧ポンプの吸収トルクを模擬した特性となるよう、前記第2油圧ポンプの吐出圧と前記ロードセンシング駆動圧力に基づいて前記第2油圧ポンプの吐出圧を補正し、トルク制御圧力として出力するトルクフィードバック回路と、
前記トルク制御圧力が導かれ、前記トルク制御圧力が高くなるにしたがって前記第1油圧ポンプの容量を減少させ前記第1最大トルクが減少するよう前記第1油圧ポンプの容量を制御する第3トルク制御アクチュエータとを有し、
前記トルクフィードバック回路は、
前記第2油圧ポンプの吐出圧が導かれる固定絞りと、
この固定絞りの下流側に位置し、下流側がタンクに接続された可変絞り弁と、
前記固定絞りと前記可変絞り弁との間の油路に接続され、前記油路の圧力を前記第2トルク制御部の制御を開始する圧力以上にならないように制御する圧力制限弁とを有し、
前記可変絞り弁は、前記ロードセンシング駆動圧力が最低圧力にあるときは全閉し、前記ロードセンシング駆動圧力が高くなるにしたがって開口面積が大きくなるよう構成され、
前記トルクフィードバック回路は、前記固定絞りと前記可変絞り弁との間の油路の圧力に基づいて前記トルク制御圧力を生成し、このトルク制御圧力が前記第3トルク制御アクチュエータに導かれることを特徴とする建設機械の油圧駆動装置。 - 請求項1記載の建設機械の油圧駆動装置において、
前記トルクフィードバック回路は、前記第2油圧ポンプの吐出圧が一次圧として導かれる減圧弁を更に備え、
前記固定絞りと前記可変絞り弁との間の油路の圧力が前記減圧弁のセット圧を設定する目標制御圧力として前記減圧弁に導かれ、
前記減圧弁は、前記第2油圧ポンプの吐出圧が前記セット圧よりも低いときは、前記第2油圧ポンプの吐出圧をそのまま二次圧力として出力し、前記第2油圧ポンプの吐出圧が前記セット圧よりも高いときは、前記第2油圧ポンプの吐出圧を前記セット圧に減圧して出力し、前記減圧弁の出力圧が前記トルク制御圧力として前記第3トルク制御アクチュエータに導かれることを特徴とする建設機械の油圧駆動装置。 - 請求項1又は2記載の建設機械の油圧駆動装置において、
前記圧力制限弁はリリーフ弁であることを特徴とする建設機械の油圧駆動装置。
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CN201480046560.9A CN105473872B (zh) | 2013-11-28 | 2014-11-26 | 工程机械的液压驱动装置 |
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WO2012093703A1 (ja) * | 2011-01-06 | 2012-07-12 | 日立建機株式会社 | 履帯式走行装置を備えた作業機の油圧駆動装置 |
JP6194259B2 (ja) * | 2014-01-31 | 2017-09-06 | Kyb株式会社 | 作業機の制御システム |
JP6510396B2 (ja) * | 2015-12-28 | 2019-05-08 | 日立建機株式会社 | 作業機械 |
CN107158693A (zh) * | 2017-07-13 | 2017-09-15 | 谷子赫 | 六自由度游戏模拟器 |
KR102133312B1 (ko) * | 2017-09-08 | 2020-07-13 | 히다찌 겐끼 가부시키가이샤 | 유압 구동 장치 |
CN109707688B (zh) * | 2018-12-29 | 2020-08-18 | 中国煤炭科工集团太原研究院有限公司 | 一种具有前置压力补偿器的流量抗饱负载敏感多路阀 |
US11753800B2 (en) * | 2020-03-27 | 2023-09-12 | Hitachi Construction Machinery Tierra Co., Ltd. | Hydraulic drive system for construction machine |
JP7471901B2 (ja) * | 2020-04-28 | 2024-04-22 | ナブテスコ株式会社 | 流体圧駆動装置 |
US11680381B2 (en) | 2021-01-07 | 2023-06-20 | Caterpillar Underground Mining Pty. Ltd. | Variable system pressure based on implement position |
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