WO2003001067A1 - Hydraulic driving unit for working machine, and method of hydraulic drive - Google Patents

Hydraulic driving unit for working machine, and method of hydraulic drive Download PDF

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Publication number
WO2003001067A1
WO2003001067A1 PCT/JP2002/006138 JP0206138W WO03001067A1 WO 2003001067 A1 WO2003001067 A1 WO 2003001067A1 JP 0206138 W JP0206138 W JP 0206138W WO 03001067 A1 WO03001067 A1 WO 03001067A1
Authority
WO
WIPO (PCT)
Prior art keywords
hydraulic pump
hydraulic
pump
displacement
pressure
Prior art date
Application number
PCT/JP2002/006138
Other languages
French (fr)
Japanese (ja)
Inventor
Hirokazu Shimomura
Tomohiko Yasuda
Original Assignee
Hitachi Construction Machinery Co., Ltd.
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Hitachi Construction Machinery Co., Ltd. filed Critical Hitachi Construction Machinery Co., Ltd.
Priority to AU2002313244A priority Critical patent/AU2002313244B2/en
Priority to US10/344,120 priority patent/US7048515B2/en
Priority to DE60238983T priority patent/DE60238983D1/en
Priority to JP2003507430A priority patent/JP4077789B2/en
Priority to KR1020037002428A priority patent/KR100540772B1/en
Priority to EP02738772A priority patent/EP1398512B1/en
Publication of WO2003001067A1 publication Critical patent/WO2003001067A1/en

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Classifications

    • 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
    • F15B11/161—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load
    • F15B11/165—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load for adjusting the pump output or bypass in response to demand
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/06—Control using electricity
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/06—Control using electricity
    • F04B49/065—Control using electricity and making use of computers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/08—Regulating by delivery pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2201/00—Pump parameters
    • F04B2201/12—Parameters of driving or driven means
    • F04B2201/1204—Position of a rotating inclined plate
    • F04B2201/12041—Angular position
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2201/00—Pump parameters
    • F04B2201/12—Parameters of driving or driven means
    • F04B2201/1205—Position of a non-rotating inclined plate
    • F04B2201/12051—Angular position
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2203/00—Motor parameters
    • F04B2203/06—Motor parameters of internal combustion engines
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2203/00—Motor parameters
    • F04B2203/06—Motor parameters of internal combustion engines
    • F04B2203/0603—Torque
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2205/00—Fluid parameters
    • F04B2205/05—Pressure after the pump outlet
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2205/00—Fluid parameters
    • F04B2205/09—Flow through the pump
    • 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
    • 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
    • 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
    • F15B2211/255—Flow control functions
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/30525—Directional control valves, e.g. 4/3-directional control valve
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/3056—Assemblies of multiple valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/31—Directional control characterised by the positions of the valve element
    • F15B2211/3105—Neutral or centre positions
    • F15B2211/3116—Neutral or centre positions the pump port being open in the centre position, e.g. so-called open centre
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/315—Directional control characterised by the connections of the valve or valves in the circuit
    • F15B2211/3157—Directional control characterised by the connections of the valve or valves in the circuit being connected to a pressure source, an output member and a return line
    • F15B2211/31576—Directional control characterised by the connections of the valve or valves in the circuit being connected to a pressure source, an output member and a return line having a single pressure source and a single output member
    • 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
    • F15B2211/50—Pressure control
    • F15B2211/505—Pressure control characterised by the type of pressure control means
    • F15B2211/50509—Pressure control characterised by the type of pressure control means the pressure control means controlling a pressure upstream of the pressure control means
    • F15B2211/50536—Pressure control characterised by the type of pressure control means the pressure control means controlling a pressure upstream of the pressure control means using unloading valves controlling the supply pressure by diverting fluid to the return line
    • 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
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/6306—Electronic controllers using input signals representing a pressure
    • F15B2211/6309—Electronic controllers using input signals representing a pressure the pressure being a pressure source supply pressure
    • 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
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/6306—Electronic controllers using input signals representing a pressure
    • F15B2211/6316—Electronic controllers using input signals representing a pressure the pressure being a pilot pressure
    • 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
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/6333—Electronic controllers using input signals representing a state of the pressure source, e.g. swash plate angle
    • 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
    • F15B2211/635—Circuits providing pilot pressure to pilot pressure-controlled fluid circuit elements
    • F15B2211/6355—Circuits providing pilot pressure to pilot pressure-controlled fluid circuit elements having valve means
    • 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
    • F15B2211/6651—Control of the prime mover, e.g. control of the output torque or rotational speed
    • 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
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/71—Multiple output members, e.g. multiple hydraulic motors or cylinders

Definitions

  • the present invention relates to an engine provided in a working machine such as a hydraulic excavator and having a fuel injection control device capable of controlling a governor region to an isochronous characteristic or a reverse dollar characteristic, and a variable displacement hydraulic driven by the engine.
  • the present invention relates to a hydraulic drive device and a hydraulic drive method for a working machine including a pump.
  • a variable displacement hydraulic pump driven by the engine a regulator for controlling the displacement of the hydraulic pump, and a pressure discharged from the hydraulic pump are generally used.
  • a plurality of hydraulic actuators driven by oil, a pressure detector that detects the discharge pressure of the hydraulic pump and outputs a discharge pressure signal, and a discharge pressure signal output from the pressure detector is input and regulated.
  • a controller for outputting a control signal for controlling the displacement of the hydraulic pump during the race.
  • the engine output characteristic is such that the engine speed increases as the engine output torque (engine load) decreases in the governor region where the mechanical governor is controlled. It has droop characteristics. Such droop characteristics are caused by the inertia of the flywheel included in the mechanical governor.
  • the discharge pressure of the hydraulic pump will be low during empty load operation after loading earth and sand in a bucket, and the engine load will be reduced and the engine speed will increase.
  • the discharge flow rate of the hydraulic pump increases, the flow rate supplied to the hydraulic actuator increases, and the hydraulic pump discharges relatively quickly. Evening speed can be obtained. As a result, the work speed in unloading operation is increased, and work efficiency can be improved.
  • the governor region has an isochronous characteristic. Or have a fuel injection control device that can be controlled to reverse dollar characteristics
  • the isochronous characteristic of the engine control is independent of the lightness of the engine load, that is, the engine output torque. Regardless of the decrease, the engine speed is maintained constant in the governor region.
  • the reverse droop characteristic is a characteristic in which the engine speed decreases as the engine output torque (engine load) decreases.
  • Such conventional technology eliminates the influence of the inertia of a flywheel, such as a mechanical governor, and achieves low fuel consumption and low noise compared to a work machine equipped with an engine that has a mechanical two-wheel governor. it can. Disclosure of the invention
  • a working machine equipped with an engine that performs isochronous control or reverse droop control has the advantage of realizing low fuel consumption and low noise, but the engine speed does not increase even when the engine is lightly loaded. Therefore, there may be a problem in work.
  • the work equipment is a hydraulic excavator, and the empty load operation is performed and the engine load is light, the discharge flow rate of the hydraulic pump does not increase because the engine speed does not increase.
  • the engine load may be light.
  • the operating speed of the hydraulic actuator does not increase as in the case of a work machine with a mechanical two-strength engine with a governor-type engine, so the operation feeling may be uncomfortable.
  • An object of the present invention is to provide at least a part of the governor region with isochronous characteristics and reverse
  • a hydraulic drive system equipped with an engine having a fuel injection control device that can control any of the loop characteristics work that can increase the discharge flow rate of the hydraulic pump as the engine load becomes lighter even in the governor region
  • a hydraulic drive device and a hydraulic drive method for a machine work that can increase the discharge flow rate of the hydraulic pump as the engine load becomes lighter even in the governor region.
  • the present invention provides a fuel injection control device capable of controlling at least a part of a governor region to one of a isochronous characteristic, a reverse droop characteristic, and a combination of the isochronous characteristic and the reverse droop characteristic.
  • a hydraulic pump for a working machine comprising: an engine having a hydraulic pump driven by the engine; and a plurality of hydraulic factories driven by hydraulic oil discharged from the hydraulic pump.
  • Pump absorption torque control means for controlling the displacement of the hydraulic pump so that the displacement of the hydraulic pump does not exceed a value determined by a preset pump absorption torque curve when the discharge pressure of the pump exceeds the first predetermined pressure;
  • a flow rate correction control means for controlling so that the displacement volume of the hydraulic pump in accordance with decreases from the second predetermined pressure is increased.
  • the engine when the engine load during operation is heavy and the discharge pressure of the hydraulic pump is higher than the first predetermined pressure, the engine is controlled by the pump absorption torque control (pump absorption horsepower control).
  • the output horsepower can be used effectively.
  • the flow rate correction control means pushes the hydraulic pump in accordance with a decrease in the pump discharge pressure.
  • the displacement is controlled so as to increase, so that the discharge flow rate of the hydraulic pump can be increased in the governor region even if the engine speed does not increase due to the isochronous characteristic or the reverse droop characteristic. Overnight speed can be increased.
  • the present invention provides a fuel injection system in which at least a part of the governor region can be controlled to any of isochronous characteristics, reverse droop characteristics, or a combination of isochronous characteristics and reverse droop characteristics.
  • An engine having a control device; a variable displacement hydraulic pump driven by the engine; and a plurality of hydraulic actuators driven by pressure oil discharged from the hydraulic pump.
  • the hydraulic pump for controlling the displacement of the hydraulic pump, a pressure detector for detecting a discharge pressure of the hydraulic pump, and a pressure detector for detecting the discharge pressure of the hydraulic pump.
  • a pump absorption torque control means for controlling the regulation so that the displacement of the hydraulic pump does not exceed a value determined by a preset pump absorption torque curve when the discharge pressure exceeds a first predetermined pressure; and the hydraulic pump.
  • the flow rate correction control means for controlling the regulator so that the displacement of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure.
  • the pump output torque control (pump absorption horsepower control) effectively utilizes the output horsepower of the engine and the pump discharge flow rate when the engine is lightly loaded. Increase control becomes possible, and the hydraulic actuator speed can be increased when the engine is lightly loaded.
  • the second predetermined pressure is equal to the first predetermined pressure.
  • the flow rate correction control means functions immediately, and the displacement of the hydraulic pump can be increased.
  • control canceling means for invalidating the increase control of the displacement of the hydraulic pump by the flow rate correction control means.
  • the fuel injection control device is capable of controlling at least a part of the governor region to have an isochronous characteristic
  • the control release means is preferably provided with a traveling mode switch. Includes at least one of a lifting mode switch and a leveling mode switch.
  • the hydraulic actuator speed is set to a constant speed regardless of the increase or decrease of the engine load. Traveling operation, lifting work, and ground preparation work can be performed.
  • the flow rate correction control controls the displacement of the hydraulic pump so that the discharge flow rate of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure.
  • the discharge flow rate of the hydraulic pump can be increased in the governor region even if the engine speed does not increase due to the isochronous characteristic or the reverse droop characteristic.
  • the fuel injection control device can control at least a part of the governor region to have a reverse droop characteristic
  • the flow rate correction control means First means for controlling the displacement of the hydraulic pump so that the discharge flow rate of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure; and (2) Select the second means for controlling the displacement of the hydraulic pump so that the discharge flow rate of the hydraulic pump is kept constant when the pressure decreases from a predetermined pressure, and select one of the first means and the second means Selection means to perform the selection.
  • the discharge flow rate of the hydraulic pump is controlled to increase when the first means is selected, and the discharge flow rate of the hydraulic pump is kept constant when the second means is selected.
  • the flow is controlled in accordance with the work content.
  • the flow rate correction control means further includes a third means for invalidating the increase control of the displacement of the hydraulic pump, and the selection means includes: One of the first means, the second means, and the third means is selected.
  • the control for increasing the displacement of the hydraulic pump is invalidated, and the flow rate can be controlled according to the work content.
  • the pump absorption torque control means includes a target displacement for controlling a pump absorption torque based on a discharge pressure of the hydraulic pump and a pump absorption torque curve.
  • the pump absorption torque control means and the flow rate correction control means can be computerized.
  • the pump absorption torque control means limits the maximum value of the displacement of the hydraulic pump to a value not more than a value determined by the pump absorption torque curve.
  • the flow rate correction control means is means for controlling so that the maximum value of the displacement of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from a second predetermined pressure.
  • pump absorption torque control pump absorption horsepower control
  • pump discharge flow rate when the engine is lightly loaded. If the required flow rate in the factory is small, the displacement of the hydraulic pump is controlled accordingly to obtain a desired factory speed.
  • (11) Further, in the above (1) or (2), further comprising a first calculating means for calculating a first target displacement according to a required flow rate of the plurality of hydraulic factories, wherein the pump absorption torque control The means calculates a second target displacement for pump absorption torque control from the discharge pressure of the hydraulic pump and the pump absorption torque curve, and when the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure.
  • the corrected second target displacement becomes the target displacement for control
  • the displacement of the hydraulic pump becomes the corrected second target displacement.
  • pump absorption torque control pump absorption horsepower control
  • the first target displacement becomes the target displacement for control, so that the displacement of the hydraulic pump is equal to the first target displacement. It is controlled according to the required flow rate based on the volume, and a desired factor overnight speed can be obtained.
  • the present invention controls at least a part of the governor region to any one of the isochronous characteristic, the reverse droop characteristic, and the characteristic combining the isochronous characteristic and the reverse droop characteristic.
  • a work machine comprising: an engine having a possible fuel injection control device; a variable displacement hydraulic pump driven by the engine; and a plurality of hydraulic factories driven by hydraulic oil discharged from the hydraulic pump.
  • the hydraulic driving method of (1) when the discharge pressure of the hydraulic pump exceeds the first predetermined pressure, the displacement of the hydraulic pump is adjusted so that the displacement of the hydraulic pump does not exceed a value determined by a predetermined pump absorption torque curve.
  • the discharge pressure of the hydraulic pump Control shall be performed so that the displacement of the hydraulic pump increases as the pressure decreases from the constant pressure.
  • the discharge flow rate of the hydraulic pump can be increased in the governor region even if the engine speed does not increase due to the isochronous characteristic or the reverse droop characteristic.
  • the fuel injection control device is capable of controlling at least a part of the governor region to a reverse-drip characteristic, and the discharge pressure of the hydraulic pump
  • the displacement of the hydraulic pump is increased so that the discharge flow rate of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure.
  • FIG. 1 is a diagram showing an entire system including a hydraulic circuit of a hydraulic drive device of a working machine according to a first embodiment of the present invention.
  • FIG. 2 is a diagram showing an external appearance of a hydraulic shovel on which the hydraulic drive device according to the present embodiment is mounted.
  • FIG. 3 is a characteristic diagram showing a relationship between a rotation speed and an output torque of an engine having an electronic governor that performs isochronous control.
  • FIG. 4 is a diagram showing the details of the structure of the Reggiore.
  • FIG. 5 is a diagram showing a relationship between a control current signal supplied to the electromagnetic proportional pressure reducing valve in the regulation and a tilt angle of the hydraulic pump.
  • FIG. 6 is a functional block diagram showing the arithmetic functions of the work implement controller.
  • FIG. 7 is a diagram showing the relationship between the pump discharge pressure used in the second target tilt angle calculation unit of the work machine controller and the second target tilt.
  • FIG. 8 is a diagram illustrating a relationship between a pump discharge pressure and a pump tilt angle correction value used in the tilt angle correction value calculation unit of the work machine controller.
  • FIG. 9 is a diagram illustrating a relationship between the pump discharge pressure corrected by the adding unit and the second target pump displacement.
  • FIG. 1 OA is a diagram showing the relationship between the pump discharge pressure P and the pump displacement 0 according to the related art having a mechanical governor engine for controlling the governor region to a dollar-gap characteristic
  • FIG. FIG. 6 is a diagram showing a relationship between a pump discharge pressure and a pump discharge flow rate according to a conventional technique.
  • FIG. 11A is a diagram showing the relationship between the pump discharge pressure P and the pump displacement 0 according to the present embodiment and a conventional technology having an engine that controls the governor region to the isochronous characteristic
  • FIG. It is a figure which shows the relationship between the pump discharge pressure and the pump discharge flow rate by the same prior art and this embodiment.
  • FIG. 12 is a characteristic diagram showing the relationship between the rotation speed and the output torque of an engine having an electronic governor that controls the reverse droop characteristic according to the second embodiment of the present invention.
  • FIG. 13 is a functional block diagram showing an arithmetic function of a work implement controller according to the second embodiment.
  • FIG. 14 is a diagram illustrating the relationship between the pump discharge pressure and the pump tilt angle correction value used in the tilt angle correction value calculation unit of the work machine controller.
  • FIG. 15 is a diagram illustrating a relationship between the discharge pressure signal corrected by the adding unit and the second target displacement.
  • FIG. 16A is a diagram showing the relationship between the pump discharge pressure P and the pump tilt 0 according to the conventional technology having an engine that controls the governor region to have the reverse droop characteristic
  • FIG. FIG. 4 is a diagram illustrating a relationship between a pump discharge pressure and a pump discharge flow rate.
  • FIG. 17A is a diagram showing the relationship between the pump discharge pressure P and the pump tilt ⁇ ⁇ ⁇ according to the second embodiment
  • FIG. 17B is a diagram showing the pump discharge pressure and the pump displacement according to the second embodiment.
  • FIG. 4 is a diagram showing a relationship with a discharge flow rate.
  • FIG. 18 is a characteristic diagram showing the relationship between the rotation speed and output torque of an engine having an electronic governor that performs control combining the isochronous characteristic and the reverse droop characteristic according to the third embodiment of the present invention. is there.
  • FIG. 19 is a diagram showing the relationship between the pump discharge pressure and the pump tilt angle correction value used in the tilt angle correction value calculation unit of the work machine controller.
  • FIG. 20 is a diagram illustrating a relationship between the discharge pressure signal corrected by the adding unit and the second target displacement.
  • FIG. 1 is a diagram showing an entire system including a hydraulic circuit of a hydraulic drive device for a working machine according to an embodiment of the present invention.
  • the hydraulic drive device is provided in a working machine, for example, a hydraulic excavator.
  • a working machine for example, a hydraulic excavator.
  • an engine 1 and an electronic governor 1 2 constituting a fuel injection control device of the engine 1 are provided.
  • an engine controller 13 The electronic governor 12 and the engine controller 13 are capable of controlling the governor region to have an isochronous characteristic.
  • the electronic governor 12 is controlled by the engine controller 13 and injects fuel into the engine 1.
  • This type of fuel injection control device is known, for example, from Japanese Patent Application Laid-Open No. H10-159599.
  • the hydraulic drive device includes, for example, a swash plate type variable displacement hydraulic pump 2 driven by an engine 1 and a displacement volume (swash plate) of the hydraulic pump 2. And a plurality of hydraulic actuators, such as a hydraulic cylinder 3, a hydraulic motor 4, a hydraulic cylinder 5, 6, etc., driven by hydraulic oil discharged from the hydraulic pump 2.
  • the pilot pressure for switching the directional control valves 7 to 10, the main relief valve 11 and the directional control valves 7 to 10 for controlling the flow of the pressure oil supplied to these hydraulic actuators is controlled.
  • a pressure detector 14 that detects the discharge pressure of the hydraulic pump 2 and outputs a discharge pressure signal P, and a tilt angle of the swash plate of the hydraulic pump 2
  • the displacement angle detector 15 that detects the displacement and outputs the displacement angle signal ⁇
  • the mode selection switch 17 that can output the control unlock signal F, and the operating lever device 50,.
  • a signal control valve 53 that has a combination of a shuttle valve that inputs pilot pressure and selects and outputs one of the pilot pressures, and a pilot pressure signal D that is detected by detecting the pilot pressure output from the signal control valve 53
  • the output pressure detector 55, the discharge pressure signal P output from the pressure detector 14 and the tilt angle signal output from the tilt angle detector 15 and the control output from the mode selection switch 17 Release signal F, Pilot output from pressure detector 5 5
  • a work machine controller 18 that inputs a pressure signal D and outputs a control current signal R for controlling the displacement volume to a regulator 16.
  • FIG. 2 shows an external view of a hydraulic shovel on which the hydraulic drive device according to the present embodiment is mounted.
  • the hydraulic excavator has a lower traveling body 102, an upper revolving body 103, and a front work machine 104, and the upper revolving body 103 is mounted on the lower traveling body 102 so as to be pivotable,
  • the front work machine 104 is attached to the front of the upper swing body 103 so as to be vertically movable.
  • the upper revolving structure 103 is provided with an engine room 105 and a driver's cab 106.
  • the front work machine 104 is a multi-joint structure having a boom 108, an arm 109, and a bucket 110.
  • the lower traveling body 102, the upper revolving body 103, and the front work machine 104 are respectively left and right traveling motors 111 (only one is shown), rotating motors 112, and boom cylinders 111 as an actuator.
  • arm cylinder 1 1 4 bucket cylinder 1 1 5
  • the lower traveling body 102 travels by rotation of the left and right traveling motors 1 1
  • the upper revolving body 1 0 3 is the rotating motor 1 1
  • the boom 1108 of the front work machine 104 rotates vertically by the expansion and contraction of the boom cylinder 113
  • the arm cylinder 109 rotates by the expansion and contraction of the arm cylinder 114.
  • the bucket 110 rotates up and down and forward and backward by expansion and contraction of the bucket cylinder 115.
  • the hydraulic cylinders 3, 5, and 6 are a boom cylinder 113, an arm cylinder 111, and a bucket cylinder 115, and the hydraulic motor 4 is a rotating motor—112. .
  • the operation lever devices 50,... And the mode selection switch 17 are arranged in the operator cab 106, and the engine 1 and the hydraulic pump 2 are arranged in the engine room 105.
  • Hydraulic equipment and electronic equipment such as the directional control valves 7 to 10, the engine controller 13, and the work equipment controller 18, are installed at appropriate places on the upper swing body 103.
  • Fig. 3 shows the relationship between the rotational speed N of the engine 1 and the output torque Te by the fuel injection control device (using the electronic governor 12 and the engine controller 13) that implements isochronous control.
  • the output torque characteristic of the engine 1 is divided into a characteristic of a governor region 33 (isochronous characteristic) represented by a straight line 32 and a characteristic of a full load region represented by a curve 30.
  • the governor region 33 is an output region where the governor opening is 100% or less
  • the full load region is an output region where the governor opening is 100%.
  • the broken line 31 shows the characteristic (droop characteristic) in the governor region of the conventional mechanical governor engine for comparison. Since the two-force lug governor has a structure in which the amount of fuel injection is adjusted by the balance between the flywheel and the spring, the governor region of the mechanical governor engine is as shown by the broken line 31 1.
  • the engine has a droop characteristic in which the engine speed N increases as Te decreases.
  • the engine speed N is kept at the rated speed NO by the electronic governor 12 irrespective of the decrease in the engine output torque Te. It has isochronous characteristics for performing isochronous control to maintain the above. By this isochronous control, lower fuel consumption and lower noise can be realized as compared to a working machine equipped with a mechanical governor type engine.
  • Fig. 4 shows the details of the regiyure overnight.
  • the regulator 16 controls the tilt angle of the hydraulic pump 2 according to the control current signal R output from the work machine controller 18 so as to match the target pump tilt angle indicated by the control current signal R. It has an electromagnetic proportional pressure reducing valve 60, a servo valve 61, and a servo piston 62.
  • the proportional solenoid pressure reducing valve 60 receives the control current signal R from the work implement controller 18 and receives the control current
  • the control pressure is output in proportion to the signal R
  • the servo valve 61 operates by the control pressure to control the position of the servo piston 62
  • the servo piston 62 drives the swash plate 2a of the hydraulic pump 2. Control its tilt angle.
  • the discharge pressure of the hydraulic pump 2 is guided to the input port of the servo valve 61 via the check valve 63, and is always acting on the small diameter chamber 62a of the servo piston 62 via the passage 54. .
  • the discharge pressure of the pilot pump 66 is led to the input port of the electromagnetic proportional pressure reducing valve 60, and the pressure is reduced to the control pressure by operating the electromagnetic proportional pressure reducing valve 60.
  • This control pressure acts on the pilot piston 61 a of the servo valve 61 through the passage 67.
  • the discharge pressure of the hydraulic pump 2 is lower than the discharge pressure of the pilot pump 66, the discharge pressure of the pilot pump 66 is guided to the input port of the servo valve 61 via the check valve 69 as servo assist pressure.
  • FIG. 5 shows the control current signal R given to the electromagnetic proportional pressure reducing valve 60 and the tilt angle of the swash plate 2a of the hydraulic pump 2 (hereinafter, simply referred to as the tilt angle of the hydraulic pump 2 or the pump tilt).
  • the discharge pressure of the self-pump 2 also acts on the small diameter chamber 62 a of the servo piston 62 through the passage 54, but the piston 62 moves to the right in the figure due to the area difference. I do.
  • the feedback lever 71 rotates counterclockwise in the figure about the pin 72 as a fulcrum. Since the end of the feedback lever 71 is connected to the sleeve 61d by the pin 73, the sleeve 61d moves leftward in the figure.
  • the movement of the servo piston 62 is performed until the notch in the opening of the sleeve 61d and the spool 61b is closed, and when it is completely closed, the servo piston 61 stops.
  • the tilt angle of the hydraulic pump 2 becomes the minimum position, and the discharge flow rate of the hydraulic pump 2 becomes the minimum.
  • the control current signal R becomes larger than R1 and the proportional solenoid pressure reducing valve 60 operates
  • the control pressure corresponding to the operation amount of the solenoid proportional pressure reducing valve 60 passes through the passage 67 and the pilot piston of the servo valve 61. Acts on 6 1a and moves the spool 6 1b rightward in the figure to a position where it balances the force of the spring 6 1c.
  • the large-diameter chamber 6 2 b of the servo piston 62 is connected to the tank 75 via a passage inside the spool 61 b.
  • the feedback lever 71 rotates clockwise about the pin 72, and the sleeve 61d of the servo valve 61 moves rightward in the figure.
  • the movement of the servo piston 62 is performed until the notch in the opening of the sleeve 6Id and the spool 61b closes, and when it is completely closed, the servo piston 61 stops.
  • the tilt angle of the hydraulic pump 2 increases, and the discharge flow rate of the hydraulic pump 2 increases.
  • the increase in the discharge flow rate of the hydraulic pump 2 is proportional to the increase in the control pressure, that is, the increase in the control current signal R.
  • the spool 61b of the servo valve 61 returns to the left in the figure to a position where the spool 61b balances the force of the spring 61c.
  • the discharge pressure of the hydraulic pump 2 (or the discharge pressure of the pilot pump 66) passes through the sleeve 61d of the servo valve 62, the spool 61b, and acts on the large-diameter chamber 62b of the servo piston 62 to reduce the diameter.
  • the servo piston 52 moves to the right in the figure due to the area difference with the chamber 62a.
  • the feedback lever 71 rotates counterclockwise in the figure with the pin 72 as a fulcrum, and the sleeve 61d of the servo valve 61 moves leftward in the figure.
  • the movement of the servo piston 62 is performed until the notch in the opening of the sleeve 6Id and the spool 61b is closed, and when it is completely closed, the servo piston 61 stops.
  • the tilt angle of the pump 2 becomes smaller, and the discharge flow rate of the hydraulic pump 2 Decreases.
  • the decrease in the discharge flow rate of the hydraulic pump 2 is proportional to the decrease in the control pressure, that is, the decrease in the control current signal R.
  • FIG. 6 is a functional block diagram showing details of the mode selection switch 17 and an arithmetic function of the work implement controller 18.
  • the mode selection switch 17 includes, for example, a traveling mode switch 17a, a load mode switch 17b, and a terrain mode switch 17c, and one of these switches 17a to l7c is operated by an operator. Outputs control release signal F when operated.
  • the work implement controller 18 includes a first target pump tilt angle calculating section 81, a second target pump tilt angle calculating section 82, a tilt angle correction value calculating section 83, and a switching section 84. , An addition unit 85, a minimum value selection unit 86, a subtraction unit 87, and a control current calculation unit 88.
  • the first target pump tilt angle calculation unit 81 receives the pilot pressure signal D from the pressure detector 55, refers to this table in a table stored in the memory, and indicates the signal D at that time. Calculate the first target displacement of the hydraulic pump 2 corresponding to the pilot pressure.
  • This first target tilt is a target tilt of positive control according to the lever operation amount (required flow rate) of the operating lever device 50,... (See FIG. 1), and the pilot pressure increases in the memory table. Therefore, the relationship between the two is set so that the first target displacement also increases.
  • the second target pump displacement angle calculation unit 82 receives the discharge pressure signal P of the hydraulic pump 2 from the pressure detector 14 and refers to this to a table stored in the memory.
  • the second target displacement ⁇ T of the hydraulic pump 2 corresponding to the pump discharge pressure indicated by P (hereinafter, for convenience, the same sign P as the signal) is calculated.
  • the second target displacement ⁇ T is a limit value for controlling the torque of the hydraulic pump 2
  • the table of the memory stores the pump discharge pressure based on the pump absorption torque curve as shown in FIG.
  • the relationship between P and the second target tilt (limit value) of the hydraulic pump 2 is set.
  • reference numeral 20 denotes a pump absorption torque curve, which is set to match the curve 21 of the output torque Te (see FIG. 3) at a predetermined rotation speed (for example, the rated rotation speed NO) of the engine 1. .
  • the pump discharge pressure P is in the range of P1 or more, the second target pump displacement changes along its pump absorption torque curve 20 and the pump discharge As the pressure P increases, the second target pump displacement ⁇ T decreases.
  • the second target pump displacement ⁇ T is the first maximum displacement ⁇ maxl. If the discharge pressure P is lower than P1, the second target pump displacement ⁇ The tilt is kept at the first maximum tilt 0 maxl.
  • the first maximum tilt 0 maxl is based on the design specifications of the hydraulic excavator, for example, the above-described swing motor 1 1 2, boom cylinder 1 1 3, arm cylinder 1 1 4, bucket cylinder 1 1 5 (hydraulic cylinders 3, 4, This value is determined by design specifications such as the operating speed of 6 and the hydraulic motor 4). In other words, the first maximum displacement 0 maxl is set so that the pump discharge flow rate obtained thereby gives a desired operating speed for those factories.
  • P min is the minimum discharge pressure of the hydraulic pump 2
  • Pmax is the maximum discharge pressure of the hydraulic pump 2.
  • the maximum discharge pressure P max corresponds to the set pressure of the main relief valve 11 (see Fig. 1).
  • a range 23 between the minimum discharge pressure Pmin and the pressure P1 is a region corresponding to the governor region 33 described above.
  • the absorption torque of the hydraulic pump 2 is represented by the product of the discharge pressure of the hydraulic pump 2 and the displacement (tilt angle) of the hydraulic pump 2. Therefore, the second target tilt 0 T corresponding to the pump discharge pressure P is calculated from the pump absorption torque curve 20, and the tilt angle of the hydraulic pump 2 is controlled to be the second target pump tilt 0 T. This means that the hydraulic pump 2 tilts so that the product of the pump discharge pressure P and the second target pump tilt (absorbing torque of the hydraulic pump 2) is maintained at the pump absorbing torque (constant value) represented by the curve 20.
  • the tilt angle correction value calculation unit 83 receives the discharge pressure signal P of the hydraulic pump 2 from the pressure detector 14 and refers to the table to a table stored in the memory. Calculates the correction value S of the second target displacement of the hydraulic pump 2 corresponding to the pump discharge pressure indicated by (hereinafter similarly denoted by the same reference symbol P as the signal).
  • the correction value S is controlled by the isochronous control to increase the displacement angle of the hydraulic pump 2 as the engine load becomes lighter, even when the engine speed in the governor region 33 (FIG. 3) is constant, and the discharge flow rate is reduced. This is for correcting the tilt angle of the hydraulic pump 2 so that it increases, and the table in the memory stores the correction value when the pump discharge pressure P is equal to or higher than P1, as shown in FIG.
  • S 0 and the discharge pressure P becomes smaller than P1
  • the relationship between the discharge pressure P and the correction value S is set so that the correction value S increases linearly proportionally as the discharge pressure P decreases. ing.
  • the switching unit 84 opens when the control release signal F is output from the mode selection switch 17 to invalidate the correction value S of the target pump displacement.
  • the addition unit 85 corrects the target pump displacement calculated by the tilt angle correction value calculation unit 83 to the second target displacement of the hydraulic pump 2 calculated by the second target pump displacement angle calculation unit 82.
  • the value S is added, and the corrected second target tilt is calculated.
  • FIG. 9 shows the relationship between the discharge pressure P corrected by the adding unit 85 and the second target displacement ⁇ T.
  • the characteristic line 19 shown in FIG. 7 is corrected as shown by the characteristic line 22, and the correction is made as the pump discharge pressure P decreases from P1 to Pmin.
  • the obtained second target tilt ⁇ T linearly increases from the first maximum tilt ⁇ maxl to the second maximum tilt ⁇ ma 2 (second first maximum tilt 0 maxl + S max).
  • the second maximum tilt 0 max2 is set, for example, corresponding to the structural maximum tilt of the hydraulic pump 2 (pump performance limit).
  • the minimum value selector 86 is configured to calculate the first target tilt> D of the hydraulic pump 2 calculated by the first target pump tilt angle calculator 81 and the second target tilt ⁇ corrected by the adder 85.
  • the smaller of T is selected and set as the target tilt Sc for controlling the hydraulic pump 2.
  • the corrected second target tilt is calculated. 6 T is output as the target pump displacement 0 c for control, and the target pump displacement 0 c for control is limited to the corrected second target displacement ⁇ T or less.
  • the subtraction unit 87 calculates the deviation between the control target pump displacement 0 c and the displacement angle signal 0 output from the displacement angle detector 15, and the control current calculation unit 88 performs, for example, an integral control operation.
  • the control current signal R is calculated from the difference ⁇ 0 by calculation.
  • the tilt angle signal 0 is controlled so as to coincide with the control target pump tilt 0c.
  • the hydraulic pump 2 When the engine 1 is started, the hydraulic pump 2 is driven, and one of the operation lever devices 50,... Is operated, the hydraulic oil discharged from the hydraulic pump 2 is applied to the corresponding one of the directional control valves 7 to 10.
  • the hydraulic excavator is supplied to the hydraulic cylinders 3, 5, 6, or the hydraulic motor 4 via the hydraulic excavator.
  • the front work machine 104 of the hydraulic excavator shown in FIG. 2 is driven to perform excavation work of earth and sand.
  • the first target displacement of the hydraulic pump 2 corresponding to the pilot pressure signal D output from the pressure detector 55 is calculated by the first target pump displacement angle calculation unit 81.
  • the second target pump tilt angle calculating section 82 the second target tilt of the hydraulic pump 2 corresponding to the discharge pressure signal P of the hydraulic pump 2 output from the pressure detector 14 is calculated.
  • the tilt angle correction value calculating section 83 calculates a target tilt correction value S of the hydraulic pump 2 corresponding to the discharge pressure signal P of the hydraulic pump 2 output from the pressure detector 14.
  • the minimum value selection unit 86 calculates in the first target pump tilt angle calculation unit 81.
  • the first target tilt of the hydraulic pump 2 is selected as the target tilt ⁇ c for control, and the subtraction unit 87 and the control current calculation unit 88 adjust the tilt angle signal 0 to the target tilt 0c.
  • the control current signal R is calculated, and the control current signal R is output to the electromagnetic proportional pressure reducing valve 60 of the regulator 16.
  • Is discharged in proportion to the product of The discharge flow rate is a flow rate corresponding to the lever operation amount of the operation lever device, and the discharge flow rate is supplied to a corresponding one of the hydraulic cylinders 3, 5, 6, or the hydraulic motor 4, and the actuator is operated by the operation lever. It is driven at a speed corresponding to the operation amount of the device.
  • the minimum value selection unit 86 calculates the second target pump tilt angle calculation unit 82.
  • the second target tilt of the hydraulic pump 2 is selected as the target tilt 0 c for control, and the control current signal R calculated from the target tilt 0 c and the tilt angle signal 0 is used as a control signal.
  • Pump absorption torque control Control of the tilt angle of the hydraulic pump 2 based on the pump absorption torque curve 20 is called pump absorption torque control, and control of the discharge flow rate of the hydraulic pump 2 is called pump absorption horsepower control.
  • the discharge pressure P of the hydraulic pump 2 decreases from P2.
  • the tilt angle of the hydraulic pump 2 is increased by the correction value S 1 compared to the first maximum tilt 0 maxl which is the tilt angle when the discharge pressure of the hydraulic pump 2 is at P 1, Accordingly, the discharge flow rate of the hydraulic pump 2 also increases.
  • the correction value S increases linearly proportionally as the discharge pressure P becomes lower than P1.
  • the corrected second target displacement ⁇ T is linearly proportional to the first maximum displacement ⁇ m as the pump discharge pressure P decreases from P 1 as shown by the characteristic line 22.
  • the operating speed of the hydraulic cylinders 3, 5, 6, and hydraulic motor 4 can be increased accordingly.
  • the characteristic indicated by the characteristic line 22 substantially matches the dollar-pull characteristic line 31 of the mechanical governor shown in FIG.
  • FIGS. 10A and 10B show the relationship between the pump discharge pressure P and the pump displacement 0 and the pump discharge pressure and the pump according to the related art having a mechanical governor type engine that controls the governor region to a dollar-gap characteristic. This shows the relationship with the discharge flow rate.
  • the engine speed N increases as the pump discharge pressure P decreases from P 1, so that even if the pump displacement 0 is constant, the engine speed N With the increase, the pump discharge flow rate Q increases as shown by the broken line 26. As a result, the flow rate supplied to the hydraulic actuator is increased, and the work speed in unloading operation is increased, and work efficiency can be improved.
  • FIGS. 11A and 11B show the relationship between the pump discharge pressure P and the pump tilt 0 according to this embodiment and the related art having an engine for controlling the governor region to have isochronous characteristics, and the relationship between the pump discharge pressure and the pump discharge. This shows the relationship with the flow rate.
  • the engine speed N is constant at the rated speed NO irrespective of the decrease in the engine output torque Te as indicated by the straight line 32 in FIG. Therefore, Pmi equivalent to governor region 3 3 In the range 23 between n and P1, if the pump tilt 0 is constant as shown by the chain line 27, the pump discharge flow rate Q will also be as shown by the chain line 28 in FIG. It is constant.
  • the pump tilt ⁇ is represented by a straight line 35 corresponding to the characteristic line 22 in FIG.
  • the operation or the operation that does not require the increase control of the discharge flow rate of the hydraulic pump 2 when the engine is lightly loaded includes a traveling operation, a hanging load operation, and a leveling operation.
  • the operator operates a corresponding one of the switches 17 a to 17 c of the mode selection switch 17.
  • the control release signal F is output from the mode selection switch 17 to the work implement controller 18, the switching section 84 is opened, and the correction value S of the target pump displacement is invalidated.
  • the control for increasing the discharge flow rate of the hydraulic pump 2 based on the correction value S of the rotation angle correction value calculation unit 83 is not performed.
  • the traveling mode switch 17a of the mode selection switch 17 described above is configured to operate when a signal from the detecting means for detecting the operation of the traveling operation lever is input to the work implement controller 18. It may be. The same applies to the other mode switches 17b and 17c.
  • the pump discharge flow rate Q is gradually increased as the engine load becomes lighter. be able to. That is, it is possible to increase the pump discharge flow rate substantially equal to the flow rate increase due to the droop characteristic in the mechanical governor. As a result, it is possible to increase the speed of the hydraulic actuator at a light load of the engine, and to improve the work efficiency at a light load such as unloading work.
  • the operation of a working machine equipped with a mechanical governor engine Good operation feeling can be given to the operator who is used to the operation.
  • the correction value S by the turning angle correction value calculation unit 83 is invalidated.
  • FIGS. 12 to 17B A second embodiment of the present invention will be described with reference to FIGS. 12 to 17B.
  • the present invention is applied to a hydraulic drive device including an engine having a fuel injection control device capable of controlling a governor region to a reverse droop characteristic.
  • the overall configuration of the hydraulic drive device according to the present embodiment is substantially the same as that of the first embodiment shown in FIG. 1 except for the following points.
  • the fuel injection control device including the electronic governor 12 and the engine controller 13 shown in FIG. 1 can control the governor region to have a reverse droop characteristic.
  • the engine 1 is controlled so that the rotational speed of the engine 1 decreases as the engine output torque Te (engine load) decreases.
  • FIG. 12 shows the relationship between the rotational speed N of the engine 1 controlled by the reverse droop characteristic and the output torque Te.
  • the engine speed N decreases as the engine output torque Te (engine load) decreases. Due to the reverse droop characteristic, the engine speed at light load can be further reduced, and further lower fuel consumption and lower noise can be realized as compared with the dollar characteristic characteristic isochronous characteristic.
  • FIG. 13 is a functional block diagram illustrating the arithmetic functions of the work implement controller 18 according to the present embodiment.
  • the work implement controller 18 includes a first target pump tilt angle calculating section 81, a second target pump tilt angle calculating section 82, a first tilt angle correction value calculating section 83A, and a second Tilting angle correction value calculation section 83 B, 0 setting section 83 C, switching section 84 A, addition section 85, minimum value selection section 86, subtraction section 87, control current calculation Unit 8 has the functions of 8.
  • the first and second tilt angle correction value calculators 8 3 A and 8 3 B receive the discharge pressure signal P of the hydraulic pump 2 from the pressure detector 14 and store this in the memory, respectively. Referring to the table, a correction value S for the second target tilt of the hydraulic pump 2 is calculated.
  • the relationship between the discharge pressure P and the correction value Sa is set so that when the discharge pressure P becomes smaller than P1, the capture value Sa increases linearly in proportion to the discharge pressure P.
  • the 0 setting unit 83C outputs 0 as the correction value S.
  • the mode selection switch 17A is a dial type and has first, second and third switching positions.
  • the switching unit 84A selects the correction value Sa calculated by the first tilt angle correction value calculation unit 83A as shown in the figure.
  • the second target displacement 0 T of the hydraulic pump 2 calculated by the second target pump displacement angle calculator 82 is selected by the switching unit 84A. Then, the corrected second target tilt is calculated by adding the corrected value S.
  • FIG. 15 shows the relationship between the pump discharge pressure P corrected by the adding unit 85 and the second target displacement 0.
  • the switching unit 84A selects the correction value Sa calculated by the first tilt angle correction value calculation unit 83A
  • the characteristic line 19 in the range 34 corresponding to the governor region 33 is corrected as shown by the characteristic line 40.
  • the fourth maximum displacement emax4 is set according to, for example, the maximum displacement (pump performance limit) in the structure of the hydraulic pump 2.
  • the switching unit 84A selects the correction value Sb calculated by the second tilt angle correction value calculation unit 83B, the characteristic line 19 in the range 34 corresponding to the governor region 33 becomes like the characteristic line 41.
  • the characteristic indicated by the characteristic line 40 is apparently almost identical to the droop characteristic line 31 of the mechanical governor shown in FIG. 12, and the characteristic indicated by the characteristic line 41 is the same as the characteristic line 32 of the isochronous control shown in FIG. They almost match in appearance.
  • the engine 1 is controlled to the reverse-drag characteristic, and the discharge flow rate increase control of the hydraulic pump 2 is performed by either the correction value Sa or the correction value Sb. Except for this point, it is substantially the same as the first embodiment.
  • the correction value Sa calculated by the first tilt angle correction value calculation unit 83A is selected, the tilt angle of the hydraulic pump 2 is increased by the characteristic line 40 shown in FIG.
  • discharge flow rate increase control is performed, the mode switch 17A is switched to the second position, and the correction value Sb calculated by the second tilt angle correction value calculator 83b is selected.
  • the control for increasing the tilt angle of the hydraulic pump 2 is performed according to the characteristic line 41 shown in FIG.
  • Figures 16A and 16B show the relationship between the pump discharge pressure P and the pump displacement 0 and the relationship between the pump discharge pressure and the pump discharge flow rate according to the conventional technology having an engine that controls the governor region to the reverse droop characteristic. Show.
  • the calculation function of the work implement controller does not include the tilt angle correction value calculation unit 83, the switching unit 84, and the addition unit 85 shown in FIG. 6, it corresponds to the governor region 33.
  • the pump displacement 0 is constant as shown by the straight line 25.
  • the engine speed N decreases as the engine output torque (engine load) Te decreases, as indicated by the straight line 34 in FIG. Therefore, in the range 23 between P min and P 1, the engine speed N decreases as the pump discharge pressure P decreases from P 1, so that even if the pump displacement 0 is constant, the engine speed N Due to the decrease, the pump discharge flow rate Q decreases as shown by the broken line 44.
  • the flow rate supplied to the hydraulic actuator is reduced, and the working speed in the unloading operation is lower than that in the case of the isochronous control.
  • FIG. 17A and FIG. 17B show the relationship between the pump discharge pressure P and the pump tilt angle and the relationship between the pump discharge pressure and the pump discharge flow rate according to the present embodiment.
  • the pump displacement 0 changes as a straight line 45 corresponding to the characteristic line 40 in FIG.
  • the flow rate Q changes as shown by the straight line 46 as the pump tilt S increases. That is, even if the engine speed N decreases due to the reverse droop characteristic, the pump discharge flow rate Q increases linearly as the pump discharge pressure P decreases from P1. As a result, similarly to the prior art shown in FIGS.
  • the flow rate supplied to the hydraulic actuator is increased, the working speed in the unloading operation is increased, and the working efficiency can be improved.
  • the correction value Sb calculated by the second tilt angle correction value calculation unit 83B is selected and the characteristic line 19 shown in Fig. 15 is corrected to the characteristic line 41, the governor region 3 In the range 23 between Pmin and P1, which corresponds to 3, the pump displacement 0 changes as the straight line 47 corresponding to the characteristic line 41 in Fig. 15, and the pump discharge flow Q changes As S increases, it becomes as shown by the straight line 48.
  • the same effects as those of the first embodiment can be obtained in the hydraulic drive device including the engine controlled to the reverse dollar characteristic.
  • the mode switch 17A to the first position and selecting the correction value Sa calculated by the first tilt angle correction value calculator 83A, even in the governor region 33, As the engine load becomes lighter, the pump discharge flow rate Q can be gradually increased, and the increase in the flow rate due to the droop characteristic in the mechanical governor can be almost as large as the increase in the pump discharge flow rate.
  • the mode switching switch 17A is switched to the second position, and the mode is calculated by the second tilt angle correction value calculator 83B.
  • the correction value Sb the discharge flow rate of the hydraulic pump 2 becomes constant irrespective of the engine load
  • the speed of the hydraulic actuator is set to the same speed regardless of the increase or decrease of the engine load, and good running operation and lifting , Leveling work can be performed.
  • the hydraulic pump 2 since the hydraulic pump 2 is driven by using the engine controlled to the reverse drive characteristic, it is lighter than the first embodiment using the engine controlled to the isochronous characteristic. The engine speed at the time of load can be further reduced, and further low fuel consumption and low noise can be realized.
  • the present invention is applied to the hydraulic drive device including the engine that controls the governor region with isochronous characteristics or reverse droop characteristics.
  • the characteristics of the governor region are not limited thereto.
  • the present invention is applied to an engine having an engine whose governor region is controlled to a combination of the isochronous characteristic and the reverse droop characteristic.
  • Figure 18 shows the relationship between the engine speed N and the output torque Te controlled to a combination of isochronous characteristics and reverse droop characteristics.
  • an iso-rotation speed N is maintained at a constant value of the rated speed NO regardless of a decrease in the engine output torque Te (engine load). It has a characteristic 90 that combines a chronous characteristic and a reverse droop characteristic in which the engine speed N decreases as the engine output torque Te decreases, as shown by a straight line 90 Ob.
  • the engine speed is kept constant by the isochronous characteristic at the time of medium load, a certain speed of the engine is secured, noise and fuel consumption are improved, and when the engine load is lighter, the reverse Droop characteristics can further improve noise and fuel efficiency.
  • FIG. 19 is a diagram showing the characteristics of the pump tilt correction value S in the tilt angle correction value calculation unit 83 (see FIG. 6) when the engine has such a characteristic 90.
  • the characteristics of the pump displacement correction value S are plotted according to the characteristics of the straight lines 90a and 90b shown in Fig. 18. Line is set.
  • FIG. 20 is a characteristic diagram similar to FIG. 9 in a case where the correction value S of the tilt angle correction value calculating section 83 has the characteristic shown in FIG.
  • the characteristic line 19 is corrected to the characteristic of the polygonal line similar to the polygonal line of the correction value S like the characteristic line 91.
  • the pump tilt in the range 23 between Pmin and P 1 corresponding to the governor region 33 is performed. 0 changes as indicated by the characteristic line 91, and accordingly, the discharge flow rate of the hydraulic pump changes as indicated by the straight line 36 in FIG. 11B, and the pump discharge flow rate increases as in the first embodiment. Control can be performed.
  • the characteristic of the correction value S for increasing the pump discharge flow rate at the time of the engine light load in which the pump discharge pressure P is equal to or less than P1 substantially matches the droop characteristic in the mechanical governor.
  • a pump capable of increasing the discharge flow rate of the pump is set, the present invention is not limited to the setting of such discharge flow characteristics.
  • the slope of the characteristic line of the pump displacement correction value S shown in FIG. 8 may be made larger so that the pump discharge flow rate increases more than the pump discharge flow rate increases due to the droop characteristic, and vice versa. You may.
  • the characteristic line of the pump displacement correction value S shown in FIG. 8 may be a broken line.
  • the characteristic line of the pump displacement correction value S may be a curve instead of a straight line.
  • the pump discharge pressure at which the correction value S is 0 is made equal to P1, which is the start pressure of the control based on the pump absorption torque curve 20, but may be a lower pressure.
  • one characteristic corresponding to the dollar-pull characteristic is set as the characteristic of the correction value S for increasing the pump discharge flow rate at the time of light load of the engine in which the pump discharge pressure P becomes P1 or less.
  • one or more other characteristics may be set so that the operator can select one of them by switching the mode selection switch.
  • the mode selection switch may be a dial type that continuously changes the output so that the characteristic of the correction value S can be continuously changed.
  • the electronic governor 12 is used as the electronic governor 12 in the fuel injection control device capable of controlling to the isochronous characteristic or the reverse dollar characteristic.
  • the present invention is not limited to this.
  • a common rail fuel injection control device or a unit fuel injector control device capable of controlling the injection amount regardless of the rotation speed may be provided.
  • the tilt angle control of the hydraulic pump 2 according to the required flow rate, the absorption torque control of the hydraulic pump 2 (absorption horsepower control), the tilting of the hydraulic pump at light load, which is a feature of the present invention, are described. All the calculation of the command value of the turning angle increase control was performed by the work equipment controller 18 and the control current signal was sent to the Regulayer 16 to control the tilt angle of the hydraulic pump.
  • the control of the tilt angle of the hydraulic pump 2 according to the required flow rate and the control of the absorbing torque of the hydraulic pump 2 (absorbing horsepower control) may be performed hydraulically by a regi- ration.
  • the tilt angle of the hydraulic pump 2 is detected by the tilt angle detector 15 and the tilt angle is controlled by the feedback buckle so that the tilt angle matches the target tilt angle.
  • the displacement angle of the hydraulic pump may be controlled in an open loop without providing the displacement angle detector 15.
  • a hydraulic drive apparatus including an engine capable of controlling at least a part of a governor region to one of an isochronous characteristic, a reverse droop characteristic, and a combination of the isochronous characteristic and the reverse droop characteristic.
  • a good operation feeling can be given to an operator who is accustomed to the operation of a working machine equipped with a mechanical governor type engine.
  • control for selectively keeping the discharge flow rate of the oil pump constant is realized.
  • the operating speed of the hydraulic actuator can be kept constant irrespective of the increase or decrease of the engine load, and the operation or operation desired by the operating can be performed well.

Abstract

A fuel injection control device (an electronic governor (12) and a controller (13)) for an engine (1) is capable of controlling the governor region to an isochronus characteristic. A working machine controller (18) receives a delivery pressure signal (P) and controls a regulator (16) in such a manner that when the delivery pressure of a hydraulic pump exceeds a predetermined pressure (P1), the displacement volume of the hydraulic pump does not exceed a value determined by a preset pump absorption torque curve (20) and when the delivery pressure of the hydraulic pump (2) is not more than a predetermined pressure (P1), the displacement volume of the hydraulic pump increases as the delivery pressure of the hydraulic pump lowers from the predetermined pressure (p1). This makes it possible, even if the governor region is controlled to an isochronus characteristic, to increase the delivery rate of the hydraulic pump as the engine load becomes lighter.

Description

明細書 作業機の油圧駆動装置及び油圧駆動方法 技術分野  TECHNICAL FIELD Hydraulic drive device and hydraulic drive method for working machine
本発明は、 油圧ショベルなどの作業機に設けられ、 ガバナ領域をアイソクロナ ス特性或いは逆ドル一プ特性に制御可能な燃料噴射制御装置を有するエンジンと、 このエンジンにより駆動される可変容量型の油圧ポンプとを備えた作業機の油圧 駆動装置及び油圧駆動方法に関する。 背景技術  The present invention relates to an engine provided in a working machine such as a hydraulic excavator and having a fuel injection control device capable of controlling a governor region to an isochronous characteristic or a reverse dollar characteristic, and a variable displacement hydraulic driven by the engine. The present invention relates to a hydraulic drive device and a hydraulic drive method for a working machine including a pump. Background art
従来、 例えば特開平 7— 8 3 0 8 4号公報に示されるようにメカニカルガバナ 式エンジンを備えた作業機の油圧駆動装置がある。  2. Description of the Related Art Conventionally, there is a hydraulic drive device of a working machine equipped with a mechanical governor type engine as disclosed in, for example, Japanese Patent Application Laid-Open No. 7-83804.
この種のメカニカルガバナ式エンジンを有する従来技術は、 一般に、 エンジン により駆動される可変容量型の油圧ポンプと、 この油圧ポンプの押し除け容積を 制御するレギユレ一夕と、 油圧ポンプから吐出される圧油によつて駆動する複数 の油圧ァクチユエ一夕と、 油圧ポンプの吐出圧力を検出し吐出圧力信号を出力す る圧力検出器と、 この圧力検出器から出力される吐出圧力信号を入力し、 レギュ レー夕に油圧ポンプの押し除け容積を制御する制御信号を出力するコントローラ とを備えている。  In the prior art having a mechanical governor type engine of this kind, generally, a variable displacement hydraulic pump driven by the engine, a regulator for controlling the displacement of the hydraulic pump, and a pressure discharged from the hydraulic pump are generally used. A plurality of hydraulic actuators driven by oil, a pressure detector that detects the discharge pressure of the hydraulic pump and outputs a discharge pressure signal, and a discharge pressure signal output from the pressure detector is input and regulated. And a controller for outputting a control signal for controlling the displacement of the hydraulic pump during the race.
このメカ二力ルガバナ式ェンジンを有する従来技術では、 エンジン出力特性は、 メカニカルガバナが制御される領域であるガバナ領域において、 エンジン出カル ク （エンジン負荷） が低下するに従って、 エンジン回転数が増加するドループ特 性を有している。 このようなドループ特性は、 メカニカルガバナに含まれるフラ ィホイールの慣性により生じる。  In the conventional technology having the mechanical two-wheel governor engine, the engine output characteristic is such that the engine speed increases as the engine output torque (engine load) decreases in the governor region where the mechanical governor is controlled. It has droop characteristics. Such droop characteristics are caused by the inertia of the flywheel included in the mechanical governor.
従って作業機が例えば油圧ショベルの場合、 バケツトに土砂等を積み込んで放 土した後の空荷動作に際しては、 油圧ポンプの吐出圧が低くなり、 エンジン負荷 が軽くなつてエンジン回転数が増加するため、 油圧ポンプの吐出流量が増大し、 油圧ァクチユエ一夕に供給される流量が多くなり、 比較的速い油圧ァクチユエ一 夕速度が得られるようになる。 その結果、 空荷動作での作業速度が速くなり、 作 業能率を向上できる。 - また、 従来、 例えば特開平 1 0— 8 9 1 1 1号公報ゃ特開平 1 0— 1 5 9 5 9 9号公報、 上述のようなメカニカルガバナ式エンジンとは異なり、 ガバナ領域を アイソクロナス特性或いは逆ドル一プ特性に制御可能な燃料噴射制御装置を有す Therefore, if the work equipment is, for example, a hydraulic excavator, the discharge pressure of the hydraulic pump will be low during empty load operation after loading earth and sand in a bucket, and the engine load will be reduced and the engine speed will increase. The discharge flow rate of the hydraulic pump increases, the flow rate supplied to the hydraulic actuator increases, and the hydraulic pump discharges relatively quickly. Evening speed can be obtained. As a result, the work speed in unloading operation is increased, and work efficiency can be improved. -Conventionally, for example, Japanese Patent Application Laid-Open No. H10-891111 and Japanese Patent Application Laid-Open No. H10-15959 / 99, unlike the governor type engine described above, the governor region has an isochronous characteristic. Or have a fuel injection control device that can be controlled to reverse dollar characteristics
(以下、 適宜、 ァイソクロナス制御或いは逆ドループ制御を実施する を備えた作業機の油圧駆動装置も提案されている。 エンジン制 御のアイソクロナス特性とはェンジン負荷の軽重に係わらず、 すなわちエンジン 出力トルクの低下に係わらず、 ガバナ領域においてエンジン回転数が一定に保た れる特性であり、 逆ドループ特性とはエンジン出力トルク (エンジン負荷) が低 下するに従って、 エンジン回転数が減少する特性である。  (Hereinafter, there has been proposed a hydraulic drive device for a working machine having a function of performing an isochronous control or a reverse droop control as appropriate. The isochronous characteristic of the engine control is independent of the lightness of the engine load, that is, the engine output torque. Regardless of the decrease, the engine speed is maintained constant in the governor region. The reverse droop characteristic is a characteristic in which the engine speed decreases as the engine output torque (engine load) decreases.
このような従来技術では、 メカニカルガバナのようなフライホイ一ルの慣性に よる影響を除くことができ、 メカ'二力ルガバナを有するエンジンを備えた作業機 に比べて、 低燃費及び低騒音を実現できる。 発明の開示  Such conventional technology eliminates the influence of the inertia of a flywheel, such as a mechanical governor, and achieves low fuel consumption and low noise compared to a work machine equipped with an engine that has a mechanical two-wheel governor. it can. Disclosure of the invention
上述のようにアイソクロナス制御或いは逆ドループ制御を実施するエンジンを 備えた作業機では、 低燃費化、 低騒音化を実現できる利点はあるものの、 ェンジ ンが軽負荷の場合でもエンジン回転数が増加しないため、 作業上問題を生じるこ とがある。 例えば、 前述したように作業機が油圧ショベルの場合であって、 空荷 動作が行われ、 エンジン負荷が軽負荷であるときでも、 エンジン回転数は増加し ないため油圧ポンプの吐出流量は増えず、 油圧ァクチユエ一夕に供給される流量 を増加させることができず、 作業能率の向上を見込めない。  As described above, a working machine equipped with an engine that performs isochronous control or reverse droop control has the advantage of realizing low fuel consumption and low noise, but the engine speed does not increase even when the engine is lightly loaded. Therefore, there may be a problem in work. For example, as described above, when the work equipment is a hydraulic excavator, and the empty load operation is performed and the engine load is light, the discharge flow rate of the hydraulic pump does not increase because the engine speed does not increase. However, it is not possible to increase the flow rate supplied to the hydraulic actuator overnight, and it is not possible to expect an improvement in work efficiency.
また、 アイソクロナス制御或いは逆ドループ制御を実施するエンジンを備えた 作業機で作業する場合、 メカニカルガバナ式エンジンを備えた作業機の操作に慣 れたオペレータにとっては、 エンジン負荷が軽負荷であるにも係わらずメカ二力 ルガバナ式エンジン付きの作業機のように油圧ァクチユエ一夕速度が増加しない ので、 操作フィーリングに違和感を感じることがある。  Also, when working with a work machine equipped with an engine that performs isochronous control or reverse droop control, if the operator is accustomed to operating a work machine equipped with a mechanical governor engine, the engine load may be light. Regardless, the operating speed of the hydraulic actuator does not increase as in the case of a work machine with a mechanical two-strength engine with a governor-type engine, so the operation feeling may be uncomfortable.
本発明の目的は、 ガバナ領域の少なくとも一部をァイソクロナス特性及び逆ド ループ特性のいずれかに制御可能な燃料噴射制御装置を有するエンジンを備えた 油圧駆動装置において、 ガバナ領域にあってもエンジン負荷が軽くなるに従つて 油圧ポンプの吐出流量を増加させることができる作業機の油圧駆動装置及び油圧 駆動方法を提供することにある。 An object of the present invention is to provide at least a part of the governor region with isochronous characteristics and reverse In a hydraulic drive system equipped with an engine having a fuel injection control device that can control any of the loop characteristics, work that can increase the discharge flow rate of the hydraulic pump as the engine load becomes lighter even in the governor region A hydraulic drive device and a hydraulic drive method for a machine.
( 1 ) 上記目的を達成するために、 本発明は、 ガバナ領域の少なくとも一部を ァイソクロナス特性、 逆ドループ特性、 ァイソクロナス特性と逆ドループ特性を 組み合わせた特性のいずれかに制御可能な燃料噴射制御装置を有するエンジンと、 このエンジンにより駆動される可変容量型の油圧ポンプと、 この油圧ポンプから 吐出される圧油によって駆動する複数の油圧ァクチユエ一夕とを備える作業機の 油圧駆動装置において、 上記油圧ポンプの吐出圧力が第 1所定圧力を越えると油 圧ポンプの押しのけ容積が予め設定されたボンプ吸収トルク曲線により定まる値 を越えないよう上記油圧ポンプの押しのけ容積を制御するポンプ吸収トルク制御 手段と、 上記油圧ポンプの吐出圧力が上記第 1所定圧力以下にあるとき、 油圧ポ ンプの吐出圧力が第 2所定圧力から低くなるに従って油圧ポンプの押しのけ容積 が増加するよう制御する流量補正制御手段とを備えるものとする。  (1) In order to achieve the above object, the present invention provides a fuel injection control device capable of controlling at least a part of a governor region to one of a isochronous characteristic, a reverse droop characteristic, and a combination of the isochronous characteristic and the reverse droop characteristic. A hydraulic pump for a working machine, comprising: an engine having a hydraulic pump driven by the engine; and a plurality of hydraulic factories driven by hydraulic oil discharged from the hydraulic pump. Pump absorption torque control means for controlling the displacement of the hydraulic pump so that the displacement of the hydraulic pump does not exceed a value determined by a preset pump absorption torque curve when the discharge pressure of the pump exceeds the first predetermined pressure; When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure, the discharge of the hydraulic pump Force is assumed and a flow rate correction control means for controlling so that the displacement volume of the hydraulic pump in accordance with decreases from the second predetermined pressure is increased.
このように構成した本発明では、 作業時のエンジン負荷が重負荷であり、 油圧 ポンプの吐出圧力が第 1所定圧力よりも高いときは、 ポンプ吸収トルク制御 （ポ ンプ吸収馬力制御） によるエンジンの出力馬力の有効利用が可能となる。 また、 作業時に例えばェンジン負荷が重負荷から軽負荷に移行し、 油圧ポンプの吐出圧 力が第 2所定圧力以下となると、 流量補正制御手段によってポンプ吐出圧力の低 下に応じて油圧ポンプの押し除け容積が増加するように制御され、 これによりガ バナ領域においてァイソクロナス特性或いは逆ドループ特性によりエンジン回転 数が上昇しなくても油圧ポンプの吐出流量を増加させることができ、 エンジン軽 負荷時に油圧ァクチユエ一夕速度を増速させることができる。  According to the present invention configured as described above, when the engine load during operation is heavy and the discharge pressure of the hydraulic pump is higher than the first predetermined pressure, the engine is controlled by the pump absorption torque control (pump absorption horsepower control). The output horsepower can be used effectively. Also, for example, when the engine load shifts from a heavy load to a light load during work and the discharge pressure of the hydraulic pump falls below the second predetermined pressure, the flow rate correction control means pushes the hydraulic pump in accordance with a decrease in the pump discharge pressure. The displacement is controlled so as to increase, so that the discharge flow rate of the hydraulic pump can be increased in the governor region even if the engine speed does not increase due to the isochronous characteristic or the reverse droop characteristic. Overnight speed can be increased.
( 2 ) また、 上記目的を達成するために、 本発明は、 ガバナ領域の少なくとも 一部をァイソクロナス特性、 逆ドループ特性、 アイソクロナス特性と逆ドループ 特性を組み合わせた特性のいずれかに制御可能な燃料噴射制御装置を有するェン ジンと、 このエンジンにより駆動される可変容量型の油圧ポンプと、 この油圧ポ ンプから吐出される圧油によって駆動する複数の油圧ァクチユエ一夕とを備える 作業機の油圧駆動装置において、 上記油圧ポンプの押し除け容積を制御するレギ ユレ一夕と、 上記油圧ポンプの吐出圧力を検出する圧力検出器と、 この圧力検出 器により検出された上記油圧ポンプの吐出圧力が第 1所定圧力を越えると油圧ポ ンプの押しのけ容積が予め設定されたボンプ吸収トルク曲線により定まる値を越 えないよう上記レギユレ一夕を制御するポンプ吸収トルク制御手段と、 上記油圧 ポンプの吐出圧力が上記第 1所定圧力以下にあるとき、 油圧ポンプの吐出圧力が 第 2所定圧力から低くなるに従って油圧ポンプの押し除け容積が増加するよう上 記レギユレ一タを制御する流量補正制御手段とを備えるものとする。 (2) In order to achieve the above object, the present invention provides a fuel injection system in which at least a part of the governor region can be controlled to any of isochronous characteristics, reverse droop characteristics, or a combination of isochronous characteristics and reverse droop characteristics. An engine having a control device; a variable displacement hydraulic pump driven by the engine; and a plurality of hydraulic actuators driven by pressure oil discharged from the hydraulic pump. In a hydraulic drive device for a work machine, the hydraulic pump for controlling the displacement of the hydraulic pump, a pressure detector for detecting a discharge pressure of the hydraulic pump, and a pressure detector for detecting the discharge pressure of the hydraulic pump. A pump absorption torque control means for controlling the regulation so that the displacement of the hydraulic pump does not exceed a value determined by a preset pump absorption torque curve when the discharge pressure exceeds a first predetermined pressure; and the hydraulic pump. When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure, the flow rate correction control means for controlling the regulator so that the displacement of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure. Are provided.
このように構成した本発明においても、 上記 （1 ) で述べたように、 ポンプ吸 収トルク制御 （ポンプ吸収馬力制御） によるエンジンの出力馬力の有効利用とェ ンジン軽負荷時のポンプ吐出流量の増加制御が可能となり、 エンジン軽負荷時に 油圧ァクチユエ一タ速度を増速させることができる。  Also in the present invention configured as described above, as described in the above (1), the pump output torque control (pump absorption horsepower control) effectively utilizes the output horsepower of the engine and the pump discharge flow rate when the engine is lightly loaded. Increase control becomes possible, and the hydraulic actuator speed can be increased when the engine is lightly loaded.
( 3 ) 上記 （1 ) 又は （2 ) において、 好ましくは、 上記第 2所定圧力は上記 第 1所定圧力に一致している。  (3) In the above (1) or (2), preferably, the second predetermined pressure is equal to the first predetermined pressure.
これにより油圧ポンプの吐出圧力が第 1所定圧力以下になると、 直ちに流量補 正制御手段が機能し、 油圧ポンプの押し除け容積を増加させることができる。  As a result, when the discharge pressure of the hydraulic pump becomes equal to or lower than the first predetermined pressure, the flow rate correction control means functions immediately, and the displacement of the hydraulic pump can be increased.
( 4 ) また、 上記 （1 ) 又は （2 ) において、 上記流量補正制御手段による上 記油圧ポンプの押し除け容積の増加制御を無効にする制御解除手段を更に備える。 これにより必要に応じ流量補正制御手段による制御を解除することができ、 作 業内容に応じた流量制御が可能となる。  (4) Further, in the above (1) or (2), further provided is a control canceling means for invalidating the increase control of the displacement of the hydraulic pump by the flow rate correction control means. As a result, the control by the flow rate correction control means can be released as necessary, and the flow rate can be controlled according to the work content.
( 5 ) 上記 （4 ) において、 好ましくは、 上記燃料噴射制御装置は、 ガバナ領 域の少なくとも一部をアイソクロナス特性に制御可能なものであり、 上記制御解 除手段は、 走行モ一.ドスイッチ、 吊荷モードスィッチ、 整地モ一ドスイッチの少 なくとも 1つを含む。  (5) In the above (4), preferably, the fuel injection control device is capable of controlling at least a part of the governor region to have an isochronous characteristic, and the control release means is preferably provided with a traveling mode switch. Includes at least one of a lifting mode switch and a leveling mode switch.
これにより走行操作、 吊荷作業、 整地作業のように油圧ポンプの吐出流量の増 加制御を望まない操作或いは作業では、 油圧ァクチユエ一夕速度をエンジン負荷 の増減に係わらず等速度にし、 良好な走行操作、 吊荷作業、 整地作業を実施させ ることができる。  As a result, in operations or operations where it is not desired to increase the discharge flow rate of the hydraulic pump, such as traveling operation, hanging load operation, and ground leveling operation, the hydraulic actuator speed is set to a constant speed regardless of the increase or decrease of the engine load. Traveling operation, lifting work, and ground preparation work can be performed.
( 6 ) また、 上記 （1 ) 又は （2 ) において、 好ましくは、 上記流量補正制御 手段は、 上記油圧ポンプの吐出圧力が上記第 2所定圧力から低くなるに従って上 記油圧ポンプの吐出流量が増加するよう上記油圧ポンプの押しのけ容積を制御す る。 (6) In the above (1) or (2), preferably, the flow rate correction control The means controls the displacement of the hydraulic pump so that the discharge flow rate of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure.
これにより上記 （1 ) で述べたように、 ガバナ領域においてァイソクロナス特 性或いは逆ドループ特性によりエンジン回転数が上昇しなくても油圧ボンプの吐 出流量を増加させることができる。  As a result, as described in (1) above, the discharge flow rate of the hydraulic pump can be increased in the governor region even if the engine speed does not increase due to the isochronous characteristic or the reverse droop characteristic.
( 7 ) 更に、 上記 （1 ) 又は （2 ) において、 上記燃料噴射制御装置は、 ガバ ナ領域の少なくとも一部を逆ドループ特性に制御可能なものであり、 上記流量補 正制御手段は、 上記油圧ポンプの吐出圧力が上記第 2所定圧力から低くなるに従 つて上記油圧ポンプの吐出流量が増加するよう上記油圧ポンプの押しのけ容積を 制御する第 1手段と、 上記油圧ポンプの吐出圧力が上記第 2所定圧力から低くな るときに上記油圧ポンプの吐出流量が一定に保たれるよう上記油圧ポンプの押し のけ容積を制御する第 2手段と、 上記第 1手段と第 2手段の一方を選択する選択 手段とを有する。  (7) Further, in the above (1) or (2), the fuel injection control device can control at least a part of the governor region to have a reverse droop characteristic, and the flow rate correction control means First means for controlling the displacement of the hydraulic pump so that the discharge flow rate of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure; and (2) Select the second means for controlling the displacement of the hydraulic pump so that the discharge flow rate of the hydraulic pump is kept constant when the pressure decreases from a predetermined pressure, and select one of the first means and the second means Selection means to perform the selection.
これによりガパナ領域の特性に係わらず、 第 1手段を選択したときは油圧ボン プの吐出流量が増加するよう制御され、 第 2手段を選択したときは油圧ポンプの 吐出流量が一定に保たれるよう制御され、 作業内容に応じた流量制御が可能とな る。  As a result, regardless of the characteristics of the governor region, the discharge flow rate of the hydraulic pump is controlled to increase when the first means is selected, and the discharge flow rate of the hydraulic pump is kept constant when the second means is selected. The flow is controlled in accordance with the work content.
( 8 ) 上記 （7 ) において、 好ましくは、 上記流量補正制御手段は、 更に、 上 記油圧ポンプの押し除け容積の増加制御を無効にする第 3手段を更に有し、 上記 選択手段は、 上記第 1手段と第 2手段と第 3手段のいずれか 1つを選択するもの である。  (8) In the above (7), preferably, the flow rate correction control means further includes a third means for invalidating the increase control of the displacement of the hydraulic pump, and the selection means includes: One of the first means, the second means, and the third means is selected.
これにより第 3手段を選択したときは油圧ポンプの押し除け容積の増加制御が 無効となり、 作業内容に応じた流量制御が可能となる。  As a result, when the third means is selected, the control for increasing the displacement of the hydraulic pump is invalidated, and the flow rate can be controlled according to the work content.
( 9 ) また、 上記 （1 ) 又は （2 ) において、 好ましくは、 上記ポンプ吸収ト ルク制御手段は、 上記油圧ポンプの吐出圧力とポンプ吸収トルク曲線とからボン プ吸収トルク制御のための目標押しのけ容積を演算するとともに、 上記油圧ボン プの吐出圧力が上記第 1所定圧力以下にあるときに前記目標押しのけ容積を一定 値に保持する手段を有し、 上記流量補正制御手段は、 上記油圧ポンプの吐出圧力 が上記第 2所定圧力から低くなるに従って増加する押しのけ容積補正値を演算す る手段と、 上記目標押しのけ容積に前記押しのけ容積補正値を加算し補正された 第 2押しのけ容積を演算する手段とを有し、 この補正された目標押しのけ容積に より上記油圧ポンプの押しのけ容積を制御する。 (9) In the above (1) or (2), preferably, the pump absorption torque control means includes a target displacement for controlling a pump absorption torque based on a discharge pressure of the hydraulic pump and a pump absorption torque curve. Means for calculating the volume, and for holding the target displacement volume at a constant value when the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure. Discharge pressure Means for calculating a displacement correction value that increases as the pressure decreases from the second predetermined pressure, and means for calculating a corrected second displacement by adding the displacement correction value to the target displacement. Then, the displacement of the hydraulic pump is controlled based on the corrected target displacement.
これによりポンプ吸収トルク制御手段及び流量補正制御手段をコンピュータ化 することができる。  Thus, the pump absorption torque control means and the flow rate correction control means can be computerized.
( 1 0 ) また、 上記 （1 ) 又は （2 ) において、 好ましくは、 上記ポンプ吸収 トルク制御手段は、 上記油圧ポンプの押しのけ容積の最大値を上記ポンプ吸収ト ルク曲線により定まる値以下に制限する手段であり、 上記流量補正制御手段は、 上記油圧ポンプの吐出圧力が第 2所定圧力から低くなるに従つて上記油圧ポンプ の押しのけ容積の最大値が増加するよう制御する手段である。  (10) In the above (1) or (2), preferably, the pump absorption torque control means limits the maximum value of the displacement of the hydraulic pump to a value not more than a value determined by the pump absorption torque curve. The flow rate correction control means is means for controlling so that the maximum value of the displacement of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from a second predetermined pressure.
これにより上記 （1 ) で述べたように、 ポンプ吸収トルク制御 （ポンプ吸収馬 力制御） によるエンジンの出力馬力の有効利用とエンジン軽負荷時のポンプ吐出 流量の増加制御が可能となるとともに、 複数のァクチユエ一夕の要求流量が少な い場合はそれに応じて油圧ポンプの押しのけ容積を制御し、 所望のァクチユエ一 夕速度を得ることができる。  As described in (1) above, this makes it possible to effectively use the output horsepower of the engine by pump absorption torque control (pump absorption horsepower control) and increase the pump discharge flow rate when the engine is lightly loaded. If the required flow rate in the factory is small, the displacement of the hydraulic pump is controlled accordingly to obtain a desired factory speed.
( 1 1 ) 更に、 上記 （1 ) 又は （2 ) において、 上記複数の油圧ァクチユエ一 夕の要求流量に応じた第 1目標押しのけ容積を演算する第 1演算手段を更に備え、 上記ポンプ吸収トルク制御手段は、 上記油圧ポンプの吐出圧力とボンプ吸収トル ク曲線とからポンプ吸収トルク制御のための第 2目標押しのけ容積を演算すると ともに、 上記油圧ポンプの吐出圧力が上記第 1所定圧力以下にあるときに前記目 標押しのけ容積を一定値に保持する第 2演算手段を有し、 上記流量補正制御手段 は、 上記油圧ポンプの吐出圧力が上記第 2所定圧力から低くなるに従つて増加す る押しのけ容積補正値を演算する手段と、 前記第 2目標押しのけ容積に前記押し のけ容積補正値を加算し補正された第 2目標押しのけ容積を演算する手段とを有 し、 前記第 1目標押しのけ容積と前記補正された第 2目標押しのけ容積の小さな 方を制御用の目標押しのけ容積として選択し、 上記油圧ポンプの押しのけ容積を 制御する。  (11) Further, in the above (1) or (2), further comprising a first calculating means for calculating a first target displacement according to a required flow rate of the plurality of hydraulic factories, wherein the pump absorption torque control The means calculates a second target displacement for pump absorption torque control from the discharge pressure of the hydraulic pump and the pump absorption torque curve, and when the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure. A second calculating means for maintaining the target displacement at a constant value, wherein the flow rate correction control means includes a displacement which increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure. Means for calculating a correction value, and means for adding the displacement correction value to the second target displacement to calculate a corrected second target displacement, wherein the first target The small name how the second target displacement volume selected as the target displacement volume for the control and Shinoke volume is the correction, controls the displacement volume of the hydraulic pump.
これにより複数の油圧ァクチユエ一夕の要求流量に応じた第 1目標押しのけ容 積が補正された第 2目標押しのけ容積より大きいときは、 補正された第 2目標押 しのけ容積が制御用の目標押しのけ容積となるので、 油圧ポンプの押しのけ容積 は補正された第 2目標押しのけ容積に制限され、 上記 （1 ) で述べたようにボン プ吸収トルク制御 （ポンプ吸収馬力制御） によるエンジンの出力馬力の有効利用 とエンジン軽負荷時のポンプ吐出流量の増加制御が可能となる。 一方、 第 1目標 押しのけ容積が補正された第 2目標押しのけ容積より小さいときは、 第 1目標押 しのけ容積が制御用の目標押しのけ容積となるので、 油圧ポンプの押しのけ容積 は第 1目標押しのけ容積に基づき要求流量に応じて制御され、 所望のァクチユエ 一夕速度を得ることができる。 As a result, the first target displacement according to the required flow rate of multiple hydraulic When the product is larger than the corrected second target displacement, the corrected second target displacement becomes the target displacement for control, and the displacement of the hydraulic pump becomes the corrected second target displacement. As described in (1) above, it is possible to effectively use the output horsepower of the engine by pump absorption torque control (pump absorption horsepower control) and increase the pump discharge flow rate when the engine is lightly loaded, as described in (1) above. On the other hand, when the first target displacement is smaller than the corrected second target displacement, the first target displacement becomes the target displacement for control, so that the displacement of the hydraulic pump is equal to the first target displacement. It is controlled according to the required flow rate based on the volume, and a desired factor overnight speed can be obtained.
( 1 2 ) また、 上記目的を達成するために、 本発明は、 ガバナ領域の少なくと も一部をァイソクロナス特性、 逆ドループ特性、 ァイソクロナス特性と逆ドルー プ特性を組み合わせた特性のいずれかに制御可能な燃料噴射制御装置を有するェ ンジンと、 このエンジンにより駆動される可変容量型の油圧ポンプと、 この油圧 ポンプから吐出される圧油によって駆動する複数の油圧ァクチユエ一夕とを備え る作業機の油圧駆動方法において、 上記油圧ポンプの吐出圧力が第 1所定圧力を 越えるときは、 油圧ポンプの押しのけ容積が予め設定されたポンプ吸収トルク曲 線により定まる値を越えないよう上記油圧ポンプの押しのけ容積を制御し、 上記 油圧ポンプの吐出圧力が上記第 1所定圧力以下にあるときは、 油圧ポンプの吐出 圧力が第 2所定圧力から低くなるに従って油圧ポンプの押しのけ容積が増加する よう制御することものとする。  (12) Further, in order to achieve the above object, the present invention controls at least a part of the governor region to any one of the isochronous characteristic, the reverse droop characteristic, and the characteristic combining the isochronous characteristic and the reverse droop characteristic. A work machine comprising: an engine having a possible fuel injection control device; a variable displacement hydraulic pump driven by the engine; and a plurality of hydraulic factories driven by hydraulic oil discharged from the hydraulic pump. In the hydraulic driving method of (1), when the discharge pressure of the hydraulic pump exceeds the first predetermined pressure, the displacement of the hydraulic pump is adjusted so that the displacement of the hydraulic pump does not exceed a value determined by a predetermined pump absorption torque curve. When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure, the discharge pressure of the hydraulic pump Control shall be performed so that the displacement of the hydraulic pump increases as the pressure decreases from the constant pressure.
これにより上記 （1 ) で述べたように、 ポンプ吸収トルク制御 （ポンプ吸収馬 力制御） によるエンジンの出力馬力の有効利用とエンジン軽負荷時のポンプ吐出 流量の増加制御が可能となり、 エンジン軽負荷時に油圧ァクチユエ一夕速度を増 速させることができる。  As described in (1) above, this enables effective use of the engine output horsepower by pump absorption torque control (pump absorption horsepower control) and control to increase the pump discharge flow rate when the engine is lightly loaded. Occasionally, the speed of the hydraulic actuator can be increased.
( 1 3 ) 上記 （1 2 ) において、 好ましくは、 上記油圧ポンプの吐出圧力が上 記第 1所定圧力以下にあるときは、 油圧ポンプの吐出圧力が第 2所定圧力から低 くなるに従つて油圧ポンプの押しのけ容積が増加させる制御と、 油圧ポンプの押 しのけ容積を一定に保つ制御のいずれか一方を選択可能である。  (13) In the above (12), preferably, when the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure, as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure, Either control to increase the displacement of the hydraulic pump or control to keep the displacement of the hydraulic pump constant can be selected.
これにより必要に応じ押しのけ容積の増加制御を解除することができ、 作業内 容に応じた流量制御が可能となる。 As a result, the control for increasing the displacement can be released as necessary, Flow control according to the volume becomes possible.
( 1 4 ) また、 上記 （1 2 ) において、 好ましくは、 上記油圧ポンプの吐出圧 力が上記第 1所定圧力以下にあるときは、 油圧ポンプの吐出圧力が第 2所定圧力 から低くなるに従つて上記油圧ポンプの吐出流量が増加するよう油圧ポンプの押 しのけ容積を制御する。  (14) In the above (12), preferably, when the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure, the discharge pressure of the hydraulic pump becomes lower than the second predetermined pressure. Then, the displacement of the hydraulic pump is controlled so that the discharge flow rate of the hydraulic pump increases.
これにより上記 （1 ) で述べたように、 ガパナ領域においてァイソクロナス特 性或レゝは逆ドループ特性によりエンジン回転数が上昇しなくても油圧ポンプの吐 出流量を増加させることができる。  Thus, as described in (1) above, the discharge flow rate of the hydraulic pump can be increased in the governor region even if the engine speed does not increase due to the isochronous characteristic or the reverse droop characteristic.
( 1 5 ) また、 上記 （1 2 ) において、 好ましくは、 上記燃料噴射制御装置は、 ガバナ領域の少なくとも一部を逆ドル一プ特性に制御可能なものであり、 上記油 圧ポンプの吐出圧力が上記第 1所定圧力以下にあるときは、 上記油圧ポンプの吐 出圧力が上記第 2所定圧力から低くなるに従つて上記油圧ポンプの吐出流量が増 加するよう上記油圧ポンプの押しのけ容積を増加させる制御と、 上記油圧ポンプ の吐出圧力が上記第 2所定圧力から低くなるに従つて上記油圧ポンプの吐出流量 が一定に保たれるよう上記油圧ポンプの押しのけ容積を増加させる制御のいずれ か一方を選択可能である。  (15) Further, in the above (12), preferably, the fuel injection control device is capable of controlling at least a part of the governor region to a reverse-drip characteristic, and the discharge pressure of the hydraulic pump When the pressure is below the first predetermined pressure, the displacement of the hydraulic pump is increased so that the discharge flow rate of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure. Control to increase the displacement of the hydraulic pump so that the discharge flow rate of the hydraulic pump is kept constant as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure. Can be selected.
これによりガバナ領域の特性に係わらず、 作業内容に応じた流量制御が可能と なる。 図面の簡単な説明  As a result, regardless of the characteristics of the governor area, it is possible to control the flow rate according to the work content. BRIEF DESCRIPTION OF THE FIGURES
図 1は、 本発明の第 1の実施の形態に係わる作業機の油圧駆動装置の油圧回路 を含むシステム全体を示す図である。  FIG. 1 is a diagram showing an entire system including a hydraulic circuit of a hydraulic drive device of a working machine according to a first embodiment of the present invention.
図 2は、 本実施の形態に係わる油圧駆動装置が搭載される油圧ショベルの外観 を示す図である。  FIG. 2 is a diagram showing an external appearance of a hydraulic shovel on which the hydraulic drive device according to the present embodiment is mounted.
図 3は、 アイソクロナス制御を実施する電子ガバナを有するエンジンの回転数 と出力トルクとの関係を示す特性図である。  FIG. 3 is a characteristic diagram showing a relationship between a rotation speed and an output torque of an engine having an electronic governor that performs isochronous control.
図 4は、 レギユレ一夕の構造の詳細を示す図である。  FIG. 4 is a diagram showing the details of the structure of the Reggiore.
図 5は、 レギュレー夕の電磁比例減圧弁に与えられる制御電流信号と油圧ボン プの傾転角との関係を示す図である。 図 6は、 作業機コントローラの演算機能を示す機能プロック図である。 FIG. 5 is a diagram showing a relationship between a control current signal supplied to the electromagnetic proportional pressure reducing valve in the regulation and a tilt angle of the hydraulic pump. FIG. 6 is a functional block diagram showing the arithmetic functions of the work implement controller.
図 7は、 作業機コントローラの第 2目標傾転角演算部で用いるポンプ吐出圧力 と第 2目標傾転との関係を示す図である。  FIG. 7 is a diagram showing the relationship between the pump discharge pressure used in the second target tilt angle calculation unit of the work machine controller and the second target tilt.
図 8は、 作業機コントローラの傾転角補正値演算部で用いるポンプ吐出圧力と ポンプ傾転角補正値との関係を示す図である。  FIG. 8 is a diagram illustrating a relationship between a pump discharge pressure and a pump tilt angle correction value used in the tilt angle correction value calculation unit of the work machine controller.
図 9は、 加算部で補正されたポンプ吐出圧力と第 2目標ポンプ傾転との関係を 示す図である。  FIG. 9 is a diagram illustrating a relationship between the pump discharge pressure corrected by the adding unit and the second target pump displacement.
図 1 O Aは、 ガバナ領域をドル一プ特性に制御するメカニカルガバナ式ェンジ ンを有する従来技術によるポンプ吐出圧力 Pとポンプ傾転 0との関係を示す図で あり、 図 1 0 Bは、 同従来技術によるポンプ吐出圧力とポンプ吐出流量との関係 を示す図である。  FIG. 1 OA is a diagram showing the relationship between the pump discharge pressure P and the pump displacement 0 according to the related art having a mechanical governor engine for controlling the governor region to a dollar-gap characteristic, and FIG. FIG. 6 is a diagram showing a relationship between a pump discharge pressure and a pump discharge flow rate according to a conventional technique.
図 1 1 Aは、 ガバナ領域をァイソクロナス特性に制御するエンジンを有する従 来技術と本実施の形態によるポンプ吐出圧力 Pとポンプ傾転 0との関係を示す図 であり、 図 1 1 Bは、 同従来技術と本実施の形態によるポンプ吐出圧力とポンプ 吐出流量との関係を示す図である。  FIG. 11A is a diagram showing the relationship between the pump discharge pressure P and the pump displacement 0 according to the present embodiment and a conventional technology having an engine that controls the governor region to the isochronous characteristic, and FIG. It is a figure which shows the relationship between the pump discharge pressure and the pump discharge flow rate by the same prior art and this embodiment.
図 1 2は、 本発明の第 2の実施の形態に係わる逆ドループ特性の制御を実施す る電子ガバナを有するエンジンの回転数と出力トルクとの関係を示す特性図であ る。  FIG. 12 is a characteristic diagram showing the relationship between the rotation speed and the output torque of an engine having an electronic governor that controls the reverse droop characteristic according to the second embodiment of the present invention.
図 1 3に、 第 2の実施の形態に係わる作業機コントローラの演算機能を示す機 能ブロック図である。  FIG. 13 is a functional block diagram showing an arithmetic function of a work implement controller according to the second embodiment.
図 1 4は、 作業機コントローラの傾転角補正値演算部で用いるポンプ吐出圧力 とボンプ傾転角補正値との関係を示す図である。  FIG. 14 is a diagram illustrating the relationship between the pump discharge pressure and the pump tilt angle correction value used in the tilt angle correction value calculation unit of the work machine controller.
図 1 5は、 加算部で補正された吐出圧力信号と第 2目標傾転との関係を示す図 である。  FIG. 15 is a diagram illustrating a relationship between the discharge pressure signal corrected by the adding unit and the second target displacement.
図 1 6 Aは、 ガバナ領域を逆ドループ特性に制御するエンジンを有する従来技 術によるポンプ吐出圧力 Pとポンプ傾転 0との関係を示す図であり、 図 1 6 Bは、 同従来技術によるポンプ吐出圧力とポンプ吐出流量との関係を示す図である。 図 1 7 Aは、 第 2の実施の形態によるポンプ吐出圧力 Pとポンプ傾転 Θとの関 係を示す図であり、 図 1 7 Bは、 第 2の実施の形態によるポンプ吐出圧力とボン プ吐出流量との関係を示す図である。 FIG. 16A is a diagram showing the relationship between the pump discharge pressure P and the pump tilt 0 according to the conventional technology having an engine that controls the governor region to have the reverse droop characteristic, and FIG. FIG. 4 is a diagram illustrating a relationship between a pump discharge pressure and a pump discharge flow rate. FIG. 17A is a diagram showing the relationship between the pump discharge pressure P and the pump tilt に よ る according to the second embodiment, and FIG. 17B is a diagram showing the pump discharge pressure and the pump displacement according to the second embodiment. FIG. 4 is a diagram showing a relationship with a discharge flow rate.
図 1 8は、 本発明の第 3の実施の形態に係わるァイソクロナス特性と逆ドルー プ特性を組み合わせた制御を実施する電子ガバナを有するエンジンの回転数と出 力トルクとの関係を示す特性図である。  FIG. 18 is a characteristic diagram showing the relationship between the rotation speed and output torque of an engine having an electronic governor that performs control combining the isochronous characteristic and the reverse droop characteristic according to the third embodiment of the present invention. is there.
図 1 9は、 作業機コントローラの傾転角補正値演算部で用いるポンプ吐出圧力 とポンプ傾転角補正値との関係を示す図である。  FIG. 19 is a diagram showing the relationship between the pump discharge pressure and the pump tilt angle correction value used in the tilt angle correction value calculation unit of the work machine controller.
図 2 0は、 加算部で補正された吐出圧力信号と第 2目標傾転との関係を示す図 である。 発明を実施するための最良の形態  FIG. 20 is a diagram illustrating a relationship between the discharge pressure signal corrected by the adding unit and the second target displacement. BEST MODE FOR CARRYING OUT THE INVENTION
以下、 本発明の実施の形態を図面を用いて説明する。  Hereinafter, embodiments of the present invention will be described with reference to the drawings.
図 1は、 本発明の一実施の形態に係わる作業機の油圧駆動装置の油圧回路を含 むシステム全体を示す図である。  FIG. 1 is a diagram showing an entire system including a hydraulic circuit of a hydraulic drive device for a working machine according to an embodiment of the present invention.
本実施の形態に係わる油圧駆動装置は、 作業機、 例えば油圧ショベルに備えら れるもので、 図 1に示すように、 エンジン 1と、 このエンジン 1の燃料噴射制御 装置を構成する電子ガバナ 1 2とエンジンコントローラ 1 3を備えている。 電子 ガバヂ 1 2とエンジンコントローラ 1 3は、 ガパナ領域をァイソクロナス特性に 制御可能なもの、 つまりガバナ領域においてェンジン負荷の増減に係わらずェン ジン 1の回転数を定格回転数に維持するァイソクロナス制御を実施するものであ り、 電子ガバナ 1 2はエンジンコントローラ 1 3により制御され、 エンジン 1に 燃料を噴射する。 この種の燃料噴射制御装置は、 例えば、 特開平 1 0— 1 5 9 5 9 9号公報より公知である。  The hydraulic drive device according to the present embodiment is provided in a working machine, for example, a hydraulic excavator. As shown in FIG. 1, an engine 1 and an electronic governor 1 2 constituting a fuel injection control device of the engine 1 are provided. And an engine controller 13. The electronic governor 12 and the engine controller 13 are capable of controlling the governor region to have an isochronous characteristic. The electronic governor 12 is controlled by the engine controller 13 and injects fuel into the engine 1. This type of fuel injection control device is known, for example, from Japanese Patent Application Laid-Open No. H10-159599.
また、 本実施の形態に係わる油圧駆動装置は、 図 1に示すように、 エンジン 1 により駆動される例えば斜板式の可変容量型の油圧ポンプ 2と、 この油圧ポンプ 2の押し除け容積 （斜板の傾転角） を制御するレギユレ一夕 1 6と、 油圧ポンプ 2から吐出される圧油によって駆動する油圧シリンダ 3、 油圧モータ 4、 油圧シ リンダ 5， 6等の複数の油圧ァクチユエ一夕と、 これらの油圧ァクチユエ一夕に 供給される圧油の流れを制御する方向制御弁 7〜1 0と、 メインリリーフ弁 1 1 と、 方向制御弁 7〜1 0を切り換え操作するためのパイロット圧力を発生する操 作レバー装置 5 0 , … （1つのみ図示） と、 油圧ポンプ 2の吐出圧力を検出し吐 出圧力信号 Pを出力する圧力検出器 1 4と、 油圧ポンプ 2の斜板の傾転角 （押し のけ容積） を検出し傾転角信号 Θを出力する傾転角検出器 1 5と、 制御解錠信号 Fを出力可能なモード選択スィッチ 1 7と、 操作レバー装置 5 0， …からのパイ ロット圧力を入力しそのうちの 1つのパイロット圧力を選択し出力するシャトル 弁の組み合わせを有する信号制御弁 5 3と、 信号制御弁 5 3から出力されたパイ ロット圧力を検出しパイロット圧信号 Dを出力する圧力検出器 5 5と、 圧力検出 器 1 4から出力される吐出圧力信号 P、 及び傾転角検出器 1 5から出力される傾 転角信号 、 モード選択スィッチ 1 7から出力される制御解除信号 F、 圧力検出 器 5 5から出力されるパイロット圧力信号 Dを入力し、 レギユレ一タ 1 6に押し 除け容積を制御する制御電流信号 Rを出力する作業機コントローラ 1 8とを備え ている。 As shown in FIG. 1, the hydraulic drive device according to the present embodiment includes, for example, a swash plate type variable displacement hydraulic pump 2 driven by an engine 1 and a displacement volume (swash plate) of the hydraulic pump 2. And a plurality of hydraulic actuators, such as a hydraulic cylinder 3, a hydraulic motor 4, a hydraulic cylinder 5, 6, etc., driven by hydraulic oil discharged from the hydraulic pump 2. The pilot pressure for switching the directional control valves 7 to 10, the main relief valve 11 and the directional control valves 7 to 10 for controlling the flow of the pressure oil supplied to these hydraulic actuators is controlled. Operations that occur (Only one is shown), a pressure detector 14 that detects the discharge pressure of the hydraulic pump 2 and outputs a discharge pressure signal P, and a tilt angle of the swash plate of the hydraulic pump 2 ( The displacement angle detector 15 that detects the displacement and outputs the displacement angle signal Θ, the mode selection switch 17 that can output the control unlock signal F, and the operating lever device 50,. A signal control valve 53 that has a combination of a shuttle valve that inputs pilot pressure and selects and outputs one of the pilot pressures, and a pilot pressure signal D that is detected by detecting the pilot pressure output from the signal control valve 53 The output pressure detector 55, the discharge pressure signal P output from the pressure detector 14 and the tilt angle signal output from the tilt angle detector 15 and the control output from the mode selection switch 17 Release signal F, Pilot output from pressure detector 5 5 And a work machine controller 18 that inputs a pressure signal D and outputs a control current signal R for controlling the displacement volume to a regulator 16.
図 2に本実施の形態に係わる油圧駆動装置が搭載される油圧ショベルの外観を 示す。  FIG. 2 shows an external view of a hydraulic shovel on which the hydraulic drive device according to the present embodiment is mounted.
油圧ショベルは、 下部走行体 1 0 2、 上部旋回体 1 0 3、 フロント作業機 1 0 4を有し、 上部旋回体 1 0 3は下部走行体 1 0 2の上部に旋回可能に搭載され、 フロント作業機 1 0 4は上部旋回体 1 0 3の前部に上下動可能に取り付けられて いる。 上部旋回体 1 0 3にはエンジンルーム 1 0 5、 運転室 1 0 6が備えられて いる。 フロント作業機 1 0 4はブーム 1 0 8、 アーム 1 0 9、 バケツト 1 1 0を 有する多関節構造である。 下部走行体 1 0 2、 上部旋回体 1 0 3、 フロント作業 機 1 0 4は、 それぞれァクチユエ一夕として左右の走行モータ 1 1 1 (一方のみ 図示）、 旋回モータ 1 1 2、 ブームシリンダ 1 1 3、 アームシリンダ 1 1 4、 バ ケットシリンダ 1 1 5を有し、 下部走行体 1 0 2は左右の走行モータ 1 1 1の回 転より走行し、 上部旋回体 1 0 3は旋回モータ 1 1 2の回転により旋回し、 フロ ント作業機 1 0 4のブーム 1 0 8はブームシリンダ 1 1 3の伸縮により上下方向 に回動し、 アームシリンダ 1 0 9はァ一ムシリンダ 1 1 4の伸縮により上下、 前 後方向に回動し、 バケツト 1 1 0はバケツトシリンダ 1 1 5の伸縮により上下、 前後方向に回動する。  The hydraulic excavator has a lower traveling body 102, an upper revolving body 103, and a front work machine 104, and the upper revolving body 103 is mounted on the lower traveling body 102 so as to be pivotable, The front work machine 104 is attached to the front of the upper swing body 103 so as to be vertically movable. The upper revolving structure 103 is provided with an engine room 105 and a driver's cab 106. The front work machine 104 is a multi-joint structure having a boom 108, an arm 109, and a bucket 110. The lower traveling body 102, the upper revolving body 103, and the front work machine 104 are respectively left and right traveling motors 111 (only one is shown), rotating motors 112, and boom cylinders 111 as an actuator. 3, arm cylinder 1 1 4, bucket cylinder 1 1 5, the lower traveling body 102 travels by rotation of the left and right traveling motors 1 1 1, and the upper revolving body 1 0 3 is the rotating motor 1 1 The boom 1108 of the front work machine 104 rotates vertically by the expansion and contraction of the boom cylinder 113, and the arm cylinder 109 rotates by the expansion and contraction of the arm cylinder 114. The bucket 110 rotates up and down and forward and backward by expansion and contraction of the bucket cylinder 115.
図 1に示した油圧シリンダ 3 , 5， 6及び油圧モータ 4は上記ァクチユエ一夕 を代表するものであり、 例えば油圧シリンダ 3， 5 , 6はブームシリンダ 1 1 3、 アームシリンダ 1 1 4、 バケツトシリンダ 1 1 5であり、 油圧モータ 4は旋回モ —夕 1 1 2である。 The hydraulic cylinders 3, 5, 6 and the hydraulic motor 4 shown in FIG. For example, the hydraulic cylinders 3, 5, and 6 are a boom cylinder 113, an arm cylinder 111, and a bucket cylinder 115, and the hydraulic motor 4 is a rotating motor—112. .
また、 操作レバー装置 5 0 , …及びモード選択スィツチ 1 7は運転室 1 0 6内 に配置され、 エンジン 1及び油圧ポンプ 2はエンジンルーム 1 0 5内に設置され ている。 方向制御弁 7〜 1 0、 エンジンコントローラ 1 3、 作業機コント口一ラ 1 8等の油圧機器及び電子機器は上部旋回体 1 0 3の適所に設置されている。 図 3にァイソクロナス制御を実施する燃料噴射制御装置 （電子ガバナ 1 2とェ ンジンコントローラ 1 3に） よるエンジン 1の回転数 Nと出力トルク T eとの関 係を示す。  The operation lever devices 50,... And the mode selection switch 17 are arranged in the operator cab 106, and the engine 1 and the hydraulic pump 2 are arranged in the engine room 105. Hydraulic equipment and electronic equipment, such as the directional control valves 7 to 10, the engine controller 13, and the work equipment controller 18, are installed at appropriate places on the upper swing body 103. Fig. 3 shows the relationship between the rotational speed N of the engine 1 and the output torque Te by the fuel injection control device (using the electronic governor 12 and the engine controller 13) that implements isochronous control.
エンジン 1の出力トルク特性は、 図 3に示す如く、 直線 3 2で表されるガバナ 領域 3 3の特性 (ァイソクロナス特性) と曲線 3 0で表される全負荷領域の特性 に分けられる。 ガバナ領域 3 3はガバナの開度が 1 0 0 %以下での出力領域であ り、 全負荷領域はガバナ開度が 1 0 0 %の出力領域である。 図中、 破線 3 1は、 比較のため、 従来のメカニカルガバナ式エンジンのガバナ領域における特性 （ド ループ特性) を示している。 メ力二力ルガバナはフライホイールとバネのつり合 いによって燃料噴射量を調整する構造であるため、 メカニカルガバナ式エンジン のガバナ領域は、 破線 3 1のように、 エンジント出カルク （エンジン負荷） T e が低下するに従って、 エンジン回転数 Nが増加するドループ特性を有している。 これに対し、 本実施の形態におけるエンジン 1では、 直線 3 2のように、 ガバナ 領域では電子ガバナ 1 2によりエンジン出力トルク T eの低下に係わらずェンジ ン回転数 Nを定格回転数 NOに一定に保つアイソクロナス制御を実施するァイソ クロナス特性を有している。 このアイソクロナス制御により、 メカニカルガバナ 式エンジンを備えた作業機に比べて、 低燃費及び低騒音を実現できる。  As shown in FIG. 3, the output torque characteristic of the engine 1 is divided into a characteristic of a governor region 33 (isochronous characteristic) represented by a straight line 32 and a characteristic of a full load region represented by a curve 30. The governor region 33 is an output region where the governor opening is 100% or less, and the full load region is an output region where the governor opening is 100%. In the figure, the broken line 31 shows the characteristic (droop characteristic) in the governor region of the conventional mechanical governor engine for comparison. Since the two-force lug governor has a structure in which the amount of fuel injection is adjusted by the balance between the flywheel and the spring, the governor region of the mechanical governor engine is as shown by the broken line 31 1. The engine has a droop characteristic in which the engine speed N increases as Te decreases. On the other hand, in the engine 1 according to the present embodiment, as shown by the straight line 32, in the governor region, the engine speed N is kept at the rated speed NO by the electronic governor 12 irrespective of the decrease in the engine output torque Te. It has isochronous characteristics for performing isochronous control to maintain the above. By this isochronous control, lower fuel consumption and lower noise can be realized as compared to a working machine equipped with a mechanical governor type engine.
図 4にレギユレ一夕 1 6の詳細を示す。 レギユレ一タ 1 6は、 作業機コント口 ーラ 1 8から出力された制御電流信号 Rにより油圧ポンプ 2の傾転角を制御電流 信号 Rが示す目標ポンプ傾転角に一致するよう制御するものであり、 電磁比例減 圧弁 6 0と、 サ一ボ弁 6 1と、 サ一ボピストン 6 2とを有している。 電磁比例減 圧弁 6 0は作業機コントローラ 1 8から制御電流信号 Rを入力し、 その制御電流 信号 Rに比例した制御圧力を出力し、 サーポ弁 6 1はその制御圧力により作動し てサ一ボピストン 6 2の位置を制御し、 サーボピストン 6 2は油圧ポンプ 2の斜 板 2 aを駆動し、 その傾転角を制御する。 Fig. 4 shows the details of the regiyure overnight. The regulator 16 controls the tilt angle of the hydraulic pump 2 according to the control current signal R output from the work machine controller 18 so as to match the target pump tilt angle indicated by the control current signal R. It has an electromagnetic proportional pressure reducing valve 60, a servo valve 61, and a servo piston 62. The proportional solenoid pressure reducing valve 60 receives the control current signal R from the work implement controller 18 and receives the control current The control pressure is output in proportion to the signal R, the servo valve 61 operates by the control pressure to control the position of the servo piston 62, and the servo piston 62 drives the swash plate 2a of the hydraulic pump 2. Control its tilt angle.
油圧ポンプ 2の吐出圧力は、 チェックバルブ 6 3を介してサ一ボ弁 6 1の入力 ポートに導かれるとともに、 通路 5 4を介してサーポピストン 6 2の小径室 6 2 aに常時作用している。 パイロットポンプ 6 6の吐出圧力が電磁比例減圧弁 6 0 の入力ポートに導かれ、 電磁比例減圧弁 6 0が作動することにより減圧されて制 御圧力となる。 この制御圧力は通路 6 7を通ってサ一ボ弁 6 1のパイロットピス トン 6 1 aに作用する。 また、 油圧ポンプ 2の吐出圧力がパイロットポンプ 6 6 の吐出圧力より低いとき、 パイロットポンプ 6 6の吐出圧力がサーボアシスト圧 としてチェックバルブ 6 9を介してサ一ボ弁 6 1の入力ポートに導かれる。 図 5に電磁比例減圧弁 6 0に与えられる制御電流信号 Rと油圧ポンプ 2の斜板 2 aの傾転角 （以下、 適宜、 単に油圧ポンプ 2の傾転角或いはポンプ傾転とい う） との関係を示す。  The discharge pressure of the hydraulic pump 2 is guided to the input port of the servo valve 61 via the check valve 63, and is always acting on the small diameter chamber 62a of the servo piston 62 via the passage 54. . The discharge pressure of the pilot pump 66 is led to the input port of the electromagnetic proportional pressure reducing valve 60, and the pressure is reduced to the control pressure by operating the electromagnetic proportional pressure reducing valve 60. This control pressure acts on the pilot piston 61 a of the servo valve 61 through the passage 67. When the discharge pressure of the hydraulic pump 2 is lower than the discharge pressure of the pilot pump 66, the discharge pressure of the pilot pump 66 is guided to the input port of the servo valve 61 via the check valve 69 as servo assist pressure. I will FIG. 5 shows the control current signal R given to the electromagnetic proportional pressure reducing valve 60 and the tilt angle of the swash plate 2a of the hydraulic pump 2 (hereinafter, simply referred to as the tilt angle of the hydraulic pump 2 or the pump tilt). Shows the relationship.
制御電流信号 Rが R 1以下のとき電磁比例減圧弁 6 0は作動せず、 電磁比例減 圧弁 6 0からの制御圧力は 0である。 このためサーボ弁 6 1のスプール 6 1 bは スプリング 6 1 cによって図示左方向に押され、 油圧ポンプ 2の吐出圧力 （或い はパイロットポンプ 6 6の吐出圧） がチェックバルブ 6 3、 スリーブ 6 1 d、 ス プール 6 1 bを通ってサーボピストン 6 2の大径室 6 2 bに作用する。 サ一ボピ ストン 6 2の小径室 6 2 aにも、 通路 5 4を通って自己ポンプ 2の吐出圧力が作 用しているが、 面積差によってサ一ボピストン 6 2は図示右方に移動する。 サーボピストン 6 2が図示右方に移動すると、 フィードバックレバー 7 1はピ ン 7 2を支点として図示反時計方向に回転する。 フィードバックレバー 7 1の先 端は、 ピン 7 3でスリーブ 6 1 dと連結しているため、 スリーブ 6 1 dは図示左 方向に移動する。 サーボピストン 6 2の移動は、 スリーブ 6 1 dとスプール 6 1 bの開口部の切り欠きが閉じるまで行われ、 それが完全に閉じるとサ一ボピスト ン 6 1は停止する。  When the control current signal R is less than R1, the electromagnetic proportional pressure reducing valve 60 does not operate, and the control pressure from the electromagnetic proportional pressure reducing valve 60 is zero. For this reason, the spool 6 1b of the servo valve 6 1 is pushed leftward by a spring 6 1c, and the discharge pressure of the hydraulic pump 2 (or the discharge pressure of the pilot pump 66) is increased by the check valve 63 and the sleeve 6 1d, acting on the large-diameter chamber 6 2b of the servo piston 62 through the spool 61b. The discharge pressure of the self-pump 2 also acts on the small diameter chamber 62 a of the servo piston 62 through the passage 54, but the piston 62 moves to the right in the figure due to the area difference. I do. When the servo piston 62 moves rightward in the figure, the feedback lever 71 rotates counterclockwise in the figure about the pin 72 as a fulcrum. Since the end of the feedback lever 71 is connected to the sleeve 61d by the pin 73, the sleeve 61d moves leftward in the figure. The movement of the servo piston 62 is performed until the notch in the opening of the sleeve 61d and the spool 61b is closed, and when it is completely closed, the servo piston 61 stops.
これらの作動により油圧ポンプ 2の傾転角は最小位置になり、 油圧ポンプ 2の 吐出流量が最少になる。 制御電流信号 Rが R 1よりも大きくなり電磁比例減圧弁 6 0が作動すると、 電 磁比例減圧弁 6 0の作動量に応じた制御圧力が通路 6 7を通ってサーポ弁 6 1の パイロットピストン 6 1 aに作用し、 スプール 6 1 bをスプリング 6 1 cの力と つりあう位置まで図示右方に移動させる。 スプール 6 l bが移動するとサーボピ ストン 6 2の大径室 6 2 bは、 スプール 6 1 b内部の通路を経由してタンク 7 5 につながる。 サ一ボピストン 6 2の小径室 6 2 aには、 通路 5 4を通じて常時油 圧ポンプ 2の吐出圧力 （或いはパイロットポンプ 6 6の吐出圧） が作用している ためサーボピストン 6 2は図示左方に移動し、 大径室 6 2 bの作動油はタンク 7 5に戻される。 By these operations, the tilt angle of the hydraulic pump 2 becomes the minimum position, and the discharge flow rate of the hydraulic pump 2 becomes the minimum. When the control current signal R becomes larger than R1 and the proportional solenoid pressure reducing valve 60 operates, the control pressure corresponding to the operation amount of the solenoid proportional pressure reducing valve 60 passes through the passage 67 and the pilot piston of the servo valve 61. Acts on 6 1a and moves the spool 6 1b rightward in the figure to a position where it balances the force of the spring 6 1c. When the spool 6 lb moves, the large-diameter chamber 6 2 b of the servo piston 62 is connected to the tank 75 via a passage inside the spool 61 b. Since the discharge pressure of the hydraulic pump 2 (or the discharge pressure of the pilot pump 66) is constantly applied to the small diameter chamber 62a of the servo piston 62 through the passage 54, the servo piston 62 is on the left side of the drawing. The hydraulic oil in the large-diameter chamber 62b is returned to the tank 75.
サ一ボピストン 6 2が図示左方に移動すると、 フィードバックレバー 7 1はピ ン 7 2を支点として図示時計方向に回転し、 サーポ弁 6 1のスリーブ 6 1 dは図 示右方向に移動する。 サーボピストン 6 2の移動は、 スリーブ 6 I dとスプール 6 1 bの開口部の切り欠きが閉じるまで行われ、 それが完全に閉じるとサ一ボピ ストン 6 1は停止する。  When the servo piston 62 moves leftward in the figure, the feedback lever 71 rotates clockwise about the pin 72, and the sleeve 61d of the servo valve 61 moves rightward in the figure. The movement of the servo piston 62 is performed until the notch in the opening of the sleeve 6Id and the spool 61b closes, and when it is completely closed, the servo piston 61 stops.
これらの作動により油圧ポンプ 2の傾転角が大きくなり、 油圧ポンプ 2の吐出 流量が増加する。 また、 油圧ポンプ 2の吐出流量の増加量は制御圧力の上昇量、 つまり制御電流信号 Rの増加量に比例する。  With these operations, the tilt angle of the hydraulic pump 2 increases, and the discharge flow rate of the hydraulic pump 2 increases. The increase in the discharge flow rate of the hydraulic pump 2 is proportional to the increase in the control pressure, that is, the increase in the control current signal R.
制御電流信号 Rが低下し電磁比例減圧弁 6 0からの制御圧力が低下すると、 サ ーボ弁 6 1のスプール 6 1 bはスプリング 6 1 cの力とつりあう位置まで図示左 方に戻され、 油圧ポンプ 2の吐出圧力 （或いはパイロットポンプ 6 6の吐出圧） がサーボ弁 6 2のスリーブ 6 1 d、 スプール 6 1 bを通ってサーボピストン 6 2 の大径室 6 2 bに作用し、 小径室 6 2 aとの面積差によってサ一ボピストン 5 2 は図示右方に移動する。  When the control current signal R decreases and the control pressure from the electromagnetic proportional pressure reducing valve 60 decreases, the spool 61b of the servo valve 61 returns to the left in the figure to a position where the spool 61b balances the force of the spring 61c. The discharge pressure of the hydraulic pump 2 (or the discharge pressure of the pilot pump 66) passes through the sleeve 61d of the servo valve 62, the spool 61b, and acts on the large-diameter chamber 62b of the servo piston 62 to reduce the diameter. The servo piston 52 moves to the right in the figure due to the area difference with the chamber 62a.
サーポピストン 6 2が図示右方に移動すると、 フィードバックレバー 7 1はピ ン 7 2を支点として図示反時計方向に回転し、 サーボ弁 6 1のスリーブ 6 1 dは 図示左方向に移動する。 サーボピストン 6 2の移動は、 スリーブ 6 I dとスプ一 ル 6 1 bの開口部の切り欠きが閉じるまで行われ、 それが完全に閉じるとサーボ ピストン 6 1は停止する。  When the servo piston 62 moves rightward in the figure, the feedback lever 71 rotates counterclockwise in the figure with the pin 72 as a fulcrum, and the sleeve 61d of the servo valve 61 moves leftward in the figure. The movement of the servo piston 62 is performed until the notch in the opening of the sleeve 6Id and the spool 61b is closed, and when it is completely closed, the servo piston 61 stops.
これらの作動によりポンプ 2の傾転角が小さくなり、 油圧ポンプ 2の吐出流量 が減少する。 油圧ポンプ 2の吐出流量の減少量は制御圧力の低下量、 つまり制御 電流信号 Rの低下量に比例する。 By these operations, the tilt angle of the pump 2 becomes smaller, and the discharge flow rate of the hydraulic pump 2 Decreases. The decrease in the discharge flow rate of the hydraulic pump 2 is proportional to the decrease in the control pressure, that is, the decrease in the control current signal R.
図 6に、 モード選択スィッチ 1 7の詳細及び作業機コントローラ 1 8の演算機 能を機能ブロック図で示す。  FIG. 6 is a functional block diagram showing details of the mode selection switch 17 and an arithmetic function of the work implement controller 18.
モ一ド選択スィツチ 1 7は、 例えば走行モードスィツチ 1 7 a、 吊荷モードス イッチ 1 7 b、 整地モードスィッチ 1 7 cを備え、 これらのスィッチ 1 7 a〜l 7 cのいずれかがオペレータにより操作されると制御解除信号 Fを出力する。 作業機コントローラ 1 8は、 第 1目標ポンプ傾転角演算部 8 1と、 第 2目標ポ ンプ傾転角演算部 8 2と、 傾転角補正値演算部 8 3と、 スイッチング部 8 4と、 加算部 8 5と、 最小値選択部 8 6と、 減算部 8 7と、 制御電流演算部 8 8の各機 能を有している。  The mode selection switch 17 includes, for example, a traveling mode switch 17a, a load mode switch 17b, and a terrain mode switch 17c, and one of these switches 17a to l7c is operated by an operator. Outputs control release signal F when operated. The work implement controller 18 includes a first target pump tilt angle calculating section 81, a second target pump tilt angle calculating section 82, a tilt angle correction value calculating section 83, and a switching section 84. , An addition unit 85, a minimum value selection unit 86, a subtraction unit 87, and a control current calculation unit 88.
第 1目標ポンプ傾転角演算部 8 1は、 圧力検出器 5 5からのパイロット圧力信 号 Dを入力し、 これをメモリに記憶してあるテーブルに参照させ、 そのときの信 号 Dが示すパイロット圧力に対応する油圧ポンプ 2の第 1目標傾転 を演算す る。 この第 1目標傾転 は操作レバー装置 5 0 , … （図 1参照） のレバー操作 量 （要求流量） に応じたポジティブ制御の目標傾転であり、 メモリのテーブルに は、 パイロット圧力が増大するに従って第 1目標傾転 も増大するように両者 の関係が設定されている。  The first target pump tilt angle calculation unit 81 receives the pilot pressure signal D from the pressure detector 55, refers to this table in a table stored in the memory, and indicates the signal D at that time. Calculate the first target displacement of the hydraulic pump 2 corresponding to the pilot pressure. This first target tilt is a target tilt of positive control according to the lever operation amount (required flow rate) of the operating lever device 50,... (See FIG. 1), and the pilot pressure increases in the memory table. Therefore, the relationship between the two is set so that the first target displacement also increases.
第 2目標ポンプ傾転角演算部 8 2は、 圧力検出器 1 4からの油圧ポンプ 2の吐 出圧力信号 Pを入力し、 これをメモリに記憶してあるテーブルに参照させ、 その ときの信号 Pが示すポンプ吐出圧力 （以下、 便宜上、 信号と同じ符号 Pを付す） に対応する油圧ポンプ 2の第 2目標傾転 θ Tを演算する。 この第 2目標傾転 θ Tは 油圧ポンプ 2のトルク制御を行うための制限値となるものであり、 メモリのテー ブルには、 図 7に示すように、 ポンプ吸収トルク曲線に基づくポンプ吐出圧力 P と油圧ポンプ 2の第 2目標傾転 (制限値） との関係が設定されている。  The second target pump displacement angle calculation unit 82 receives the discharge pressure signal P of the hydraulic pump 2 from the pressure detector 14 and refers to this to a table stored in the memory. The second target displacement θT of the hydraulic pump 2 corresponding to the pump discharge pressure indicated by P (hereinafter, for convenience, the same sign P as the signal) is calculated. The second target displacement θ T is a limit value for controlling the torque of the hydraulic pump 2, and the table of the memory stores the pump discharge pressure based on the pump absorption torque curve as shown in FIG. The relationship between P and the second target tilt (limit value) of the hydraulic pump 2 is set.
図 7において、 2 0がポンプ吸収トルク曲線であり、 エンジン 1の所定回転数 (例えば、 定格回転数 NO) における出力トルク T e (図 3参照） の曲線 2 1に 一致するよう設定されている。 ポンプ吐出圧力 Pが P 1以上の範囲では、 第 2目 標ポンプ傾転 はそのポンプ吸収トルク曲線 2 0に沿って変化し、 ポンプ吐出 圧力 Pが増大するに従い第 2目標ポンプ傾転 θ Tは減少する。 In FIG. 7, reference numeral 20 denotes a pump absorption torque curve, which is set to match the curve 21 of the output torque Te (see FIG. 3) at a predetermined rotation speed (for example, the rated rotation speed NO) of the engine 1. . When the pump discharge pressure P is in the range of P1 or more, the second target pump displacement changes along its pump absorption torque curve 20 and the pump discharge As the pressure P increases, the second target pump displacement θ T decreases.
ポンプ吐出圧力 Pが P 1のとき第 2目標ポンプ傾転 θ Tは第 1最大傾転 Θ maxl であり、 吐出圧力 Pが P 1より低い範囲では、 特性線 1 9のように第 2目標ボン プ傾転 は第 1最大傾転 0 maxlに保たれる。 この第 1最大傾転 0 maxlは、 油圧 ショベルの設計仕様、 例えば前述した旋回モータ 1 1 2、 ブームシリンダ 1 1 3、 アームシリンダ 1 1 4、 バケツトシリンダ 1 1 5 (油圧シリンダ 3， 4， 6及び 油圧モータ 4 ) の動作速度等の設計仕様により定まる値である。 つまり、 第 1最 大傾転 0 maxlは、 それにより得られるポンプ吐出流量がそれらァクチユエ一夕の 所望の動作速度を与えるように設定されている。  When the pump discharge pressure P is P1, the second target pump displacement θT is the first maximum displacement Θ maxl.If the discharge pressure P is lower than P1, the second target pump displacement θ The tilt is kept at the first maximum tilt 0 maxl. The first maximum tilt 0 maxl is based on the design specifications of the hydraulic excavator, for example, the above-described swing motor 1 1 2, boom cylinder 1 1 3, arm cylinder 1 1 4, bucket cylinder 1 1 5 (hydraulic cylinders 3, 4, This value is determined by design specifications such as the operating speed of 6 and the hydraulic motor 4). In other words, the first maximum displacement 0 maxl is set so that the pump discharge flow rate obtained thereby gives a desired operating speed for those factories.
P minは油圧ポンプ 2の最低吐出圧力、 Pmaxは油圧ポンプ 2の最大吐出圧力で ある。 最大吐出圧力 P maxはメインリリーフ弁 1 1 (図 1参照） の設定圧力に対 応する。  P min is the minimum discharge pressure of the hydraulic pump 2, and Pmax is the maximum discharge pressure of the hydraulic pump 2. The maximum discharge pressure P max corresponds to the set pressure of the main relief valve 11 (see Fig. 1).
また、 最低吐出圧力 Pminと圧力 P 1の間の範囲 2 3は前述したガバナ領域 3 3に相当する領域である。  A range 23 between the minimum discharge pressure Pmin and the pressure P1 is a region corresponding to the governor region 33 described above.
油圧ポンプ 2の吸収トルクは油圧ポンプ 2の吐出圧力と油圧ポンプ 2の押しの け容積 （傾転角） との積で表される。 よって、 ポンプ吸収トルク曲線 2 0からポ ンプ吐出圧力 Pに対応する第 2目標傾転 0 Tを演算し、 この第 2目標ポンプ傾転 0 Tとなるよう油圧ポンプ 2の傾転角を制御することは、 ボンプ吐出圧力 Pと第 2目標ポンプ傾転 の積 （油圧ポンプ 2の吸収トルク） が曲線 2 0で表される ポンプ吸収トルク （一定値） に維持されるよう油圧ポンプ 2の傾転を制御するこ とを意味する。  The absorption torque of the hydraulic pump 2 is represented by the product of the discharge pressure of the hydraulic pump 2 and the displacement (tilt angle) of the hydraulic pump 2. Therefore, the second target tilt 0 T corresponding to the pump discharge pressure P is calculated from the pump absorption torque curve 20, and the tilt angle of the hydraulic pump 2 is controlled to be the second target pump tilt 0 T. This means that the hydraulic pump 2 tilts so that the product of the pump discharge pressure P and the second target pump tilt (absorbing torque of the hydraulic pump 2) is maintained at the pump absorbing torque (constant value) represented by the curve 20. Means to control
傾転角補正値演算部 8 3は、 圧力検出器 1 4からの油圧ポンプ 2の吐出圧力信 号 Pを入力し、 これをメモリに記憶してあるテーブルに参照させ、 そのときの信 号 Pが示すポンプ吐出圧力 （以下、 同様に、 信号と同じ符号 Pを付す） に対応す る油圧ポンプ 2の第 2目標傾転 の補正値 Sを演算する。 この補正値 Sは、 ァ イソクロナス制御によりガパナ領域 3 3 (図 3 ) でのエンジン回転数が一定であ つても、 エンジン負荷が軽くなるに従い油圧ポンプ 2の傾転角を増加させ吐出流 量が増加するよう油圧ポンプ 2の傾転角を補正するためのものであり、 メモリの テーブルには、 図 8に示すように、 ポンプ吐出圧力 Pが P 1以上のときは補正値 S = 0であり、 吐出圧力 Pが P 1より小さくなると、 吐出圧力 Pが小さくなるに 従つて直線比例的に補正値 Sが大きくなるように吐出圧力 Pと補正値 Sとの関係 が設定されている。 The tilt angle correction value calculation unit 83 receives the discharge pressure signal P of the hydraulic pump 2 from the pressure detector 14 and refers to the table to a table stored in the memory. Calculates the correction value S of the second target displacement of the hydraulic pump 2 corresponding to the pump discharge pressure indicated by (hereinafter similarly denoted by the same reference symbol P as the signal). The correction value S is controlled by the isochronous control to increase the displacement angle of the hydraulic pump 2 as the engine load becomes lighter, even when the engine speed in the governor region 33 (FIG. 3) is constant, and the discharge flow rate is reduced. This is for correcting the tilt angle of the hydraulic pump 2 so that it increases, and the table in the memory stores the correction value when the pump discharge pressure P is equal to or higher than P1, as shown in FIG. When S = 0 and the discharge pressure P becomes smaller than P1, the relationship between the discharge pressure P and the correction value S is set so that the correction value S increases linearly proportionally as the discharge pressure P decreases. ing.
スイッチング部 8 4は、 モード選択スィッチ 1 7から制御解除信号 Fが出力さ れると開き、 目標ポンプ傾転の補正値 Sを無効にする。  The switching unit 84 opens when the control release signal F is output from the mode selection switch 17 to invalidate the correction value S of the target pump displacement.
加算部 8 5は、 第 2目標ポンプ傾転角演算部 8 2で演算された油圧ポンプ 2の 第 2目標傾転 に傾転角補正値演算部 8 3で演算された目標ポンプ傾転の補正 値 Sを加算し、 補正された第 2目標傾転 を演算する。  The addition unit 85 corrects the target pump displacement calculated by the tilt angle correction value calculation unit 83 to the second target displacement of the hydraulic pump 2 calculated by the second target pump displacement angle calculation unit 82. The value S is added, and the corrected second target tilt is calculated.
図 9に、 加算部 8 5で補正された吐出圧力 Pと第 2目標傾転 θ Tとの関係を示 す。  FIG. 9 shows the relationship between the discharge pressure P corrected by the adding unit 85 and the second target displacement θT.
第 2目標傾転 に補正値 Sを加算することにより、 図 7に示した特性線 1 9 は特性線 2 2のように補正され、 ポンプ吐出圧力 Pが P 1から Pminに低下する に従い、 補正された第 2目標傾転 θ Tは第 1最大傾転 Θ maxlから第 2最大傾転 Θ ma 2 (二第 1最大傾転 0 maxl + S max) まで直線的に増大する。 この第 2最大傾 転 0 max2は、 例えば油圧ポンプ 2の構造上の最大傾転 （ポンプ性能限界） に対応 して設定されている。  By adding the correction value S to the second target tilt, the characteristic line 19 shown in FIG. 7 is corrected as shown by the characteristic line 22, and the correction is made as the pump discharge pressure P decreases from P1 to Pmin. The obtained second target tilt θ T linearly increases from the first maximum tilt Θ maxl to the second maximum tilt Θ ma 2 (second first maximum tilt 0 maxl + S max). The second maximum tilt 0 max2 is set, for example, corresponding to the structural maximum tilt of the hydraulic pump 2 (pump performance limit).
最小値選択部 8 6は、 第 1目標ポンプ傾転角演算部 8 1で演算された油圧ボン プ 2の第 1目標傾転 > Dと加算部 8 5で補正された第 2目標傾転 θ Tの小さい方を 選択し、 油圧ポンプ 2の制御用の目標傾転 S cとする。 これにより第 1目標ボン プ傾転角演算部 8 1で演算された油圧ポンプ 2の第 1目標傾転 が補正された 第 2目標傾転 0 Tにより大きいときは補正された第 2目標傾転 6 Tが制御用の目標 ポンプ傾転 0 cとして出力され、 制御用の目標ポンプ傾転 0 cは補正された第 2 目標傾転 Θ T以下に制限される。  The minimum value selector 86 is configured to calculate the first target tilt> D of the hydraulic pump 2 calculated by the first target pump tilt angle calculator 81 and the second target tilt θ corrected by the adder 85. The smaller of T is selected and set as the target tilt Sc for controlling the hydraulic pump 2. As a result, when the first target tilt of the hydraulic pump 2 calculated by the first target pump tilt angle calculator 81 is larger than the corrected second target tilt 0 T, the corrected second target tilt is calculated. 6 T is output as the target pump displacement 0 c for control, and the target pump displacement 0 c for control is limited to the corrected second target displacement ΘT or less.
減算部 8 7は、 制御用の目標ポンプ傾転 0 cと傾転角検出器 1 5から出力され る傾転角信号 0の偏差 を演算し、 制御電流演算部 8 8は、 例えば積分制御演 算によりその偏差 Δ 0から制御電流信号 Rを演算する。 これにより傾転角信号 0 が制御用の目標ポンプ傾転 0 cに一致するように制御される。  The subtraction unit 87 calculates the deviation between the control target pump displacement 0 c and the displacement angle signal 0 output from the displacement angle detector 15, and the control current calculation unit 88 performs, for example, an integral control operation. The control current signal R is calculated from the difference Δ 0 by calculation. Thus, the tilt angle signal 0 is controlled so as to coincide with the control target pump tilt 0c.
以上のように構成した本実施の形態における動作は以下の通りである。  The operation in the present embodiment configured as described above is as follows.
まず、 モード選択スィッチ 1 7の何れのスィッチ 1 7 a ~ l 7 cも操作されて おらず、 御解除信号 Fが出力されていない場合、 つまり作業機コントローラ 1 8 のスイッチング部 8 4が閉成している場合について説明する。 First, all the switches 17 a to l 7 c of the mode selection switch 17 are operated. The case where the release signal F is not output, that is, the case where the switching unit 84 of the work implement controller 18 is closed will be described.
エンジン 1を起動させて油圧ポンプ 2を駆動し、 操作レバー装置 5 0， …のい ずれかを操作すると、 油圧ポンプ 2から吐出された圧油が方向制御弁 7〜1 0の 該当するものを介して油圧シリンダ 3， 5 , 6、 或いは油圧モータ 4等に供給さ れ、 例えば図 2に示した油圧ショベルのフロント作業機 1 0 4が駆動し、 土砂の 掘削作業等が実施される。  When the engine 1 is started, the hydraulic pump 2 is driven, and one of the operation lever devices 50,... Is operated, the hydraulic oil discharged from the hydraulic pump 2 is applied to the corresponding one of the directional control valves 7 to 10. The hydraulic excavator is supplied to the hydraulic cylinders 3, 5, 6, or the hydraulic motor 4 via the hydraulic excavator. For example, the front work machine 104 of the hydraulic excavator shown in FIG. 2 is driven to perform excavation work of earth and sand.
作業機コントローラ 1 8では、 第 1目標ポンプ傾転角演算部 8 1において、 圧 力検出器 5 5から出力されるパイ口ット圧力信号 Dに対応する油圧ポンプ 2の第 1目標傾転 が演算され、 第 2目標ポンプ傾転角演算部 8 2において、 圧力検 出器 1 4から出力される油圧ポンプ 2の吐出圧力信号 Pに対応する油圧ポンプ 2 の第 2目標傾転 が演算され、 傾転角補正値演算部 8 3において、 圧力検出器 1 4から出力される油圧ポンプ 2の吐出圧力信号 Pに対応する油圧ポンプ 2の目 標傾転の補正値 Sが演算される。  In the work implement controller 18, the first target displacement of the hydraulic pump 2 corresponding to the pilot pressure signal D output from the pressure detector 55 is calculated by the first target pump displacement angle calculation unit 81. In the second target pump tilt angle calculating section 82, the second target tilt of the hydraulic pump 2 corresponding to the discharge pressure signal P of the hydraulic pump 2 output from the pressure detector 14 is calculated, The tilt angle correction value calculating section 83 calculates a target tilt correction value S of the hydraulic pump 2 corresponding to the discharge pressure signal P of the hydraulic pump 2 output from the pressure detector 14.
このとき、 操作レバ一装置のレバー操作量が小さく、 0 D< 0 c ( = Θ Ί) であ ると、 最小値選択部 8 6では第 1目標ポンプ傾転角演算部 8 1で演算された油圧 ポンプ 2の第 1目標傾転 が制御用の目標傾転 Θ cとして選択され、 減算部 8 7及び制御電流演算部 8 8により傾転角信号 0を目標傾転 0 cに一致させるため の制御電流信号 Rが演算され、 この制御電流信号 Rがレギユレ一夕 1 6の電磁比 例減圧弁 6 0に出力される。 これにより油圧ポンプ 2の傾転角は制御用の目標傾 転 0 c (= 0 D) に一致するよう制御され、 油圧ポンプ 2は目標傾転 0 cとその ときのエンジン 1の回転数 Nとの積に比例した流量を吐出する。 この吐出流量は 操作レバー装置のレバー操作量に応じた流量であり、 この吐出流量が油圧シリン ダ 3 , 5 , 6、 或いは油圧モータ 4の該当するものに供給され、 当該ァクチユエ —夕が操作レバー装置の操作量に応じた速度で駆動される。  At this time, if the lever operation amount of the operation lever device is small and 0 D <0 c (= Θ Ί), the minimum value selection unit 86 calculates in the first target pump tilt angle calculation unit 81. The first target tilt of the hydraulic pump 2 is selected as the target tilt Θc for control, and the subtraction unit 87 and the control current calculation unit 88 adjust the tilt angle signal 0 to the target tilt 0c. The control current signal R is calculated, and the control current signal R is output to the electromagnetic proportional pressure reducing valve 60 of the regulator 16. As a result, the tilt angle of the hydraulic pump 2 is controlled so as to match the target tilt 0 c (= 0 D) for control, and the hydraulic pump 2 adjusts the target tilt 0 c and the rotation speed N of the engine 1 at that time. Is discharged in proportion to the product of The discharge flow rate is a flow rate corresponding to the lever operation amount of the operation lever device, and the discharge flow rate is supplied to a corresponding one of the hydraulic cylinders 3, 5, 6, or the hydraulic motor 4, and the actuator is operated by the operation lever. It is driven at a speed corresponding to the operation amount of the device.
一方、 例えば操作レバ一装置の操作レバーをフル操作し、 0 D> 0 c ( = Θ Ό であると、 最小値選択部 8 6では第 2目標ポンプ傾転角演算部 8 2で演算された 油圧ポンプ 2の第 2目標傾転 が制御用の目標傾転 0 cとして選択され、 この 目標傾転 0 cと傾転角信号 0とから演算された制御電流信号 Rがレギユレ一夕 1 6の電磁比例減圧弁 6 0に出力される。 On the other hand, for example, when the operation lever of the operation lever device is fully operated and 0 D> 0 c (= Θ Ό, the minimum value selection unit 86 calculates the second target pump tilt angle calculation unit 82. The second target tilt of the hydraulic pump 2 is selected as the target tilt 0 c for control, and the control current signal R calculated from the target tilt 0 c and the tilt angle signal 0 is used as a control signal. Output to the electromagnetic proportional pressure reducing valve 60.
このとき例えば、 重掘削等が実施され、 圧力検出器 1 4から出力される信号 P が示すポンプ吐出圧力が図 9に示す P 1よりも高い P 2であると、 傾転角補正値 演算部 8 3では補正値 S = 0が演算され、 第 2目標ポンプ傾転角演算部 8 2では 第 2目標傾転 0 T= 0 2が演算され、 その 0 2がそのまま補正された第 2目標傾 転 となる。 このため油圧ポンプ 2の傾転角は 0 2に制限され、 油圧ポンプ 2 の吐出流量も下記の流量 Q 1に制限される。  At this time, for example, if heavy excavation is performed and the pump discharge pressure indicated by the signal P output from the pressure detector 14 is P2 higher than P1 shown in FIG. 9, the tilt angle correction value calculation unit In 8 3, the correction value S = 0 is calculated, and in the second target pump tilt angle calculation unit 8 2, the second target tilt 0 T = 02 is calculated, and the second target tilt 0 T = 02 is corrected as it is. Inverted. For this reason, the tilt angle of the hydraulic pump 2 is limited to 02, and the discharge flow rate of the hydraulic pump 2 is also limited to the following flow rate Q1.
Q 1 = a · 0 2 · Ν  Q 1 = a
( aは定数）  (a is a constant)
このように油圧ポンプ 2の吐出流量が制限される結果、 油圧ポンプ 2の吐出流 量と吐出圧力との積で表される油圧ポンプ 2の消費馬力も制限される。 これによ りエンジン 1の過負荷を防止し、 エンジンス 1 ^一ルを生じない範囲でエンジン 1 の出力馬力の有効活用を実施できる。  As a result of the restriction on the discharge flow rate of the hydraulic pump 2, the horsepower consumption of the hydraulic pump 2 represented by the product of the discharge flow rate of the hydraulic pump 2 and the discharge pressure is also restricted. This prevents the engine 1 from overloading and makes it possible to effectively use the output horsepower of the engine 1 within a range that does not cause engine engine failure.
このボンプ吸収トルク曲線 2 0に基づく油圧ポンプ 2の傾転角の制御はポンプ 吸収トルク制御と呼ばれ、 油圧ポンプ 2の吐出流量の制御はポンプ吸収馬力制御 と呼ばれる。  Control of the tilt angle of the hydraulic pump 2 based on the pump absorption torque curve 20 is called pump absorption torque control, and control of the discharge flow rate of the hydraulic pump 2 is called pump absorption horsepower control.
上述のような状態から、 例えばバケツト 1 1 0から土砂が捨てられ、 空荷動作 となったような場合には、 油圧ポンプ 2の吐出圧力 Pが P 2から低下する。 この ポンプ吐出圧力 Pが例えば P 1より小さい P 3に低下すると、 傾転角補正値演算 部 8 3では補正値 S = S 1が演算され、 第 2目標ポンプ傾転角演算部 8 2では第 2目標傾転 0 T= 0 maxlが演算され、 補正値 S 1を 0 maxlに加算した値が補正さ れた第 2目標傾転 Θ Τとなる。 このため油圧ポンプ 2の傾転角は S maxl + S 1と なるよう制御され、 油圧ポンプ ·2の吐出流量も下記の流量 Q 3となるよう制御さ れる。  In the state described above, for example, when the earth and sand are discarded from the bucket 110 and the empty operation is performed, the discharge pressure P of the hydraulic pump 2 decreases from P2. When the pump discharge pressure P decreases to, for example, P3 which is smaller than P1, the correction value S = S1 is calculated in the tilt angle correction value calculation section 83, and the correction value S = S1 is calculated in the second target pump tilt angle calculation section 82. 2 Target tilt 0 T = 0 maxl is calculated, and the value obtained by adding the correction value S1 to 0 maxl is the corrected second target tilt Θ Τ. Therefore, the tilt angle of the hydraulic pump 2 is controlled to be Smaxl + S1, and the discharge flow rate of the hydraulic pump 2 is also controlled to be the following flow rate Q3.
Q 3 = a · ( ^ maxl + S 1 ) · N  Q 3 = a · (^ maxl + S 1) · N
つまり、 油圧ポンプ 2の傾転角は、 油圧ポンプ 2の吐出圧力が P 1にあるとき の傾転角である第 1最大傾転 0 maxlに比べ補正値 S 1の分だけ増加し、 これに伴 つて油圧ポンプ 2の吐出流量も増加する。  In other words, the tilt angle of the hydraulic pump 2 is increased by the correction value S 1 compared to the first maximum tilt 0 maxl which is the tilt angle when the discharge pressure of the hydraulic pump 2 is at P 1, Accordingly, the discharge flow rate of the hydraulic pump 2 also increases.
ここで、 補正値 Sは、 吐出圧力 Pが P 1より低くなるに従い直線比例的に大き くなるように設定されており、 補正された第 2目標傾転 θ Tは、 特性線 2 2のよ うにポンプ吐出圧力 Pが P 1から低下するに従い直線比例的に第 1最大傾転 Θ m axlから第 2最大傾転 S max2 (=第 1最大傾転 0 maxl + S max) まで増大する。 こ のため、 ァイソクロナス制御によりガバナ領域 3 3 (図 3 ) に相当する範囲 2 3 でエンジン 1の回転数が一定であっても、 ェンジン負荷が軽くなるに従い油圧ポ ンプ 2の吐出流量が増加するよう制御され、 それに応じて油圧シリンダ 3， 5 , 6、 油圧モ一夕 4等の油圧ァクチユエ一夕の動作速度を速くすることができる。 この特性線 2 2が示す特性は、 図 3に示したメカニカルガバナにおけるドル一プ 特性線 3 1と見かけ上ほぼ一致する。 Here, the correction value S increases linearly proportionally as the discharge pressure P becomes lower than P1. And the corrected second target displacement θ T is linearly proportional to the first maximum displacement Θ m as the pump discharge pressure P decreases from P 1 as shown by the characteristic line 22. axl to the second maximum tilt S max2 (= first maximum tilt 0 maxl + S max). Therefore, the discharge flow rate of the hydraulic pump 2 increases as the engine load decreases, even if the engine 1 speed is constant in the range 23 corresponding to the governor region 33 (FIG. 3) by the isochronous control. Thus, the operating speed of the hydraulic cylinders 3, 5, 6, and hydraulic motor 4 can be increased accordingly. The characteristic indicated by the characteristic line 22 substantially matches the dollar-pull characteristic line 31 of the mechanical governor shown in FIG.
図 1 0 A及び図 1 0 Bに、 ガバナ領域をドル一プ特性に制御するメカニカルガ バナ式エンジンを有する従来技術によるポンプ吐出圧力 Pとポンプ傾転 0との関 係及びポンプ吐出圧力とポンプ吐出流量との関係を示す。  FIGS. 10A and 10B show the relationship between the pump discharge pressure P and the pump displacement 0 and the pump discharge pressure and the pump according to the related art having a mechanical governor type engine that controls the governor region to a dollar-gap characteristic. This shows the relationship with the discharge flow rate.
作業機コントローラの演算機能に図 6に示した傾転角補正値演算部 8 3、 スィ ツチング部 8 4及び加算部 8 5を備えていない従来技術では、 ガバナ領域 3 3 (図 3 ) に相当する Pminと P 1の間の範囲 2 3では直線 2 5で示すようにボン プ傾転 Sは一定である。 一方、 メカニカルガバナ式エンジンのガバナ領域 3 3で は、 図 3の破線 3 1のように、 エンジント出力ルク （エンジン負荷） T eが低下 するに従ってエンジン回転数 Nが増加するドループ特性が得られる。 このため P minと P 1の間の範囲 2 3では、 ポンプ吐出圧力 Pが P 1から低下するに従って エンジン回転数 Nが増加するため、 ボンプ傾転 0が一定であつてもエンジン回転 数 Nの増加によりポンプ吐出流量 Qは破線 2 6で示すように増加する。 これによ り油圧ァクチユエ一夕に供給される流量が多くなり、 空荷動作での作業速度が速 くなり、 作業能率を向上できる。  In the conventional technology, which does not include the tilt angle correction value calculation unit 83, switching unit 84, and addition unit 85 shown in Fig. 6 in the calculation function of the work equipment controller, it corresponds to the governor region 33 (Fig. 3). In the range 23 between Pmin and P1, the pump displacement S is constant as shown by the straight line 25. On the other hand, in the governor region 33 of the mechanical governor engine, a droop characteristic in which the engine speed N increases as the engine output torque (engine load) Te decreases as shown by a broken line 31 in FIG. . Therefore, in the range 23 between P min and P 1, the engine speed N increases as the pump discharge pressure P decreases from P 1, so that even if the pump displacement 0 is constant, the engine speed N With the increase, the pump discharge flow rate Q increases as shown by the broken line 26. As a result, the flow rate supplied to the hydraulic actuator is increased, and the work speed in unloading operation is increased, and work efficiency can be improved.
図 1 1 A及び図 1 1 Bに、 ガバナ領域をアイソクロナス特性に制御するェンジ ンを有する従来技術と本実施の形態によるポンプ吐出圧力 Pとポンプ傾転 0との 関係及びポンプ吐出圧力とポンプ吐出流量との関係を示す。  FIGS. 11A and 11B show the relationship between the pump discharge pressure P and the pump tilt 0 according to this embodiment and the related art having an engine for controlling the governor region to have isochronous characteristics, and the relationship between the pump discharge pressure and the pump discharge. This shows the relationship with the flow rate.
ガバナ領域をアイソクロナス特性に制御するエンジンのガパナ領域 3 3では、 図 3の直線 3 2のようにエンジン出力トルク T eの低下に係わらずエンジン回転 数 Nは定格回転数 NOで一定である。 このため、 ガバナ領域 3 3に相当する Pmi nと P 1の間の範囲 2 3において、 一点鎖線 2 7で示すようにポンプ傾転 0がー 定である場合は、 ポンプ吐出流量 Qも、 図 1 1 Bに一点鎖線 2 8で示すように一 定である。 これに対し、 本実施の形態では、 ガバナ領域 3 3に相当する Pminと P 1の間の範囲 2 3において、 ポンプ傾転 Θは図 9の特性線 2 2に対応して直線 3 5のように変化し、 ポンプ吐出流量 Qはポンプ傾転 0の増加により直線 3 6で 示すように変化する。 つまり、 エンジン回転数 Nが一定であっても、 ポンプ吐出 圧力 Pが P 1から低下するに従ってポンプ吐出流量 Qは直線比例的に増加する。 これにより図 1 O A及び図 1 0 Bに示した従来技術と同様、 油圧ァクチユエ一夕 に供給される流量が多くなり、 空荷動作での作業速度が速くなり、 作業能率を向 上できる。 In the governor region 33 of the engine in which the governor region is controlled to have the isochronous characteristic, the engine speed N is constant at the rated speed NO irrespective of the decrease in the engine output torque Te as indicated by the straight line 32 in FIG. Therefore, Pmi equivalent to governor region 3 3 In the range 23 between n and P1, if the pump tilt 0 is constant as shown by the chain line 27, the pump discharge flow rate Q will also be as shown by the chain line 28 in FIG. It is constant. On the other hand, in the present embodiment, in the range 23 between Pmin and P1, which corresponds to the governor region 33, the pump tilt Θ is represented by a straight line 35 corresponding to the characteristic line 22 in FIG. , And the pump discharge flow rate Q changes as indicated by the straight line 36 as the pump displacement 0 increases. That is, even if the engine speed N is constant, the pump discharge flow rate Q increases linearly as the pump discharge pressure P decreases from P1. As a result, as in the prior art shown in FIGS. 1OA and 10B, the flow rate supplied to the hydraulic actuator is increased, the working speed in the unloading operation is increased, and the working efficiency can be improved.
また、 上述のようにェンジン軽負荷時における油圧ポンプ 2の吐出流量の増加 制御を望まない操作或いは作業として、 走行操作、 吊荷作業、 整地作業がある。 このような操作或いは作業をする場合は、 オペレータはモード選択スィッチ 1 7 のスィッチ 1 7 a〜l 7 cの該当するものを操作する。 これによりモード選択ス イッチ 1 7から制御解除信号 Fが作業機コントローラ 1 8に出力され、 スィッチ ング部 8 4が開かれ、 目標ポンプ傾転の補正値 Sが無効にされる。 その結果、 転 角補正値演算部 8 3の補正値 Sによる油圧ポンプ 2の吐出流量の増加制御が実施 されなくなる。  In addition, as described above, the operation or the operation that does not require the increase control of the discharge flow rate of the hydraulic pump 2 when the engine is lightly loaded includes a traveling operation, a hanging load operation, and a leveling operation. When performing such an operation or work, the operator operates a corresponding one of the switches 17 a to 17 c of the mode selection switch 17. As a result, the control release signal F is output from the mode selection switch 17 to the work implement controller 18, the switching section 84 is opened, and the correction value S of the target pump displacement is invalidated. As a result, the control for increasing the discharge flow rate of the hydraulic pump 2 based on the correction value S of the rotation angle correction value calculation unit 83 is not performed.
なお、 上述したモ一ド選択スィツチ 1 7の例えば走行モードスィツチ 1 7 aは、 走行操作レバーの操作を検出する検出手段からの信号が作業機コントローラ 1 8 に入力されたときに作動する構成となっていてもよい。 他のモードスィッチ 1 7 b , 1 7 cについても同様である。  In addition, for example, the traveling mode switch 17a of the mode selection switch 17 described above is configured to operate when a signal from the detecting means for detecting the operation of the traveling operation lever is input to the work implement controller 18. It may be. The same applies to the other mode switches 17b and 17c.
このように構成した本実施形態によれば、 ァイソクロナス制御を適用したェン ジン 1を備えたものにおいて、 ガバナ領域 3 3にあってもェンジン負荷が軽くな るに従ってポンプ吐出流量 Qを次第に増加させることができる。 すなわち、 メカ 二カルガバナにおけるドループ特性による流量の増加とほぼ同等のポンプ吐出流 量の増加を実施させることができる。 これによつてエンジン軽負荷時の油圧ァク チユエ一夕速度を増速させることができ、 空荷作業等の軽負荷時の作業能率を向 上させることができる。 また、 メカニカルガバナ式エンジンを備えた作業機の操 作に慣れたオペレータに対しても良好な操作フィーリングを与えることができる また、 走行操作、 吊荷作業、 整地作業が実施されるときには、 転角補正値演算 部 8 3による補正値 Sを無効化し、 図 3に示すアイソクロナス特性線 3 2に従つ たアイソクロナス制御を実施させることにより、 エンジン負荷に係わらず油圧ポ ンプ 2の吐出流量は一定となり、 油圧ァクチユエ一夕速度をエンジン負荷の増減 に係わらず等速度にし、 良好な走行操作、 吊荷作業、 整地作業を実施させること ができる。 According to the present embodiment configured as described above, in the engine equipped with the engine 1 to which the isochronous control is applied, even in the governor region 33, the pump discharge flow rate Q is gradually increased as the engine load becomes lighter. be able to. That is, it is possible to increase the pump discharge flow rate substantially equal to the flow rate increase due to the droop characteristic in the mechanical governor. As a result, it is possible to increase the speed of the hydraulic actuator at a light load of the engine, and to improve the work efficiency at a light load such as unloading work. In addition, the operation of a working machine equipped with a mechanical governor engine Good operation feeling can be given to the operator who is used to the operation.In addition, when traveling operation, hanging load work, and terrain work are performed, the correction value S by the turning angle correction value calculation unit 83 is invalidated. By performing isochronous control in accordance with the isochronous characteristic line 32 shown in Fig. 3, the discharge flow rate of the hydraulic pump 2 becomes constant irrespective of the engine load, and the hydraulic actuating speed can be adjusted to increase or decrease the engine load. Regardless, it is possible to carry out good running operation, hanging load work, and leveling work at the same speed.
本発明の第 2の実施の形態を図 1 2〜図 1 7 Bにより説明する。 本実施の形態 は、 ガバナ領域を逆ドル一プ特性に制御可能な燃料噴射制御装置を有するェンジ ンを備えた油圧駆動装置に本発明を適用したものである。  A second embodiment of the present invention will be described with reference to FIGS. 12 to 17B. In the present embodiment, the present invention is applied to a hydraulic drive device including an engine having a fuel injection control device capable of controlling a governor region to a reverse droop characteristic.
本実施の形態に係わる油圧駆動装置の全体構成は図 1に示した第 1の実施の形 態と下記の点を除いて実質的に同じである。  The overall configuration of the hydraulic drive device according to the present embodiment is substantially the same as that of the first embodiment shown in FIG. 1 except for the following points.
本実施の形態において、 図 1に示した電子ガバナ 1 2及びエンジンコント口一 ラ 1 3からなる燃料噴射制御装置は、 ガバナ領域を逆ドループ特性に制御可能な ものであり、 これによりエンジン 1は、 ガバナ領域においてエンジン出力トルク T e (エンジン負荷） が軽くなるに従いエンジン 1の回転数が低下するよう制御 される。  In the present embodiment, the fuel injection control device including the electronic governor 12 and the engine controller 13 shown in FIG. 1 can control the governor region to have a reverse droop characteristic. In the governor region, the engine 1 is controlled so that the rotational speed of the engine 1 decreases as the engine output torque Te (engine load) decreases.
図 1 2に逆ドループ特性に制御されるエンジン 1の回転数 Nと出力トルク T e との関係を示す。 図 1 2において、 ガバナ領域 3 3では、 直線 3 4のように、 ェ ンジン出力トルク T e (エンジン負荷） が低下するに従ってエンジン回転数 Nが 減少する逆ドル一プ特性を有している。 この逆ドループ特性により、 ドル一プ特 性ゃァイソクロナス特性に比べ、 軽負荷時のエンジン回転数を更に低下させ、 更 なる低燃費と低騒音を実現することができる。  FIG. 12 shows the relationship between the rotational speed N of the engine 1 controlled by the reverse droop characteristic and the output torque Te. In FIG. 12, in the governor region 33, as indicated by a straight line 34, the engine speed N decreases as the engine output torque Te (engine load) decreases. Due to the reverse droop characteristic, the engine speed at light load can be further reduced, and further lower fuel consumption and lower noise can be realized as compared with the dollar characteristic characteristic isochronous characteristic.
図 1 3に、 本実施の形態に係わる作業機コントローラ 1 8の演算機能を機能ブ ロック図で示す。  FIG. 13 is a functional block diagram illustrating the arithmetic functions of the work implement controller 18 according to the present embodiment.
作業機コントローラ 1 8は、 第 1目標ポンプ傾転角演算部 8 1と、 第 2目標ポ ンプ傾転角演算部 8 2と、 第 1傾転角補正値演算部 8 3 Aと、 第 2傾転角補正値 演算部 8 3 Bと、 0設定部 8 3 Cと、 スイッチング部 8 4 Aと、 加算部 8 5と、 最小値選択部 8 6と、 減算部 8 7と、 制御電流演算部 8 8の各機能を有している。 第 1及び第 2傾転角補正値演算部 8 3 A, 8 3 Bは、 それぞれ、 圧力検出器 1 4からの油圧ポンプ 2の吐出圧力信号 Pを入力し、 これをメモリに記憶してある テーブルに参照させ、 油圧ポンプ 2の第 2目標傾転 の補正値 Sを演算する。 第 1傾転角補正値演算部 8 3 Aは、 逆ドループ特性によりガパナ領域 3 3での エンジン回転数が低下しても、 エンジン負荷が軽くなるに従い油圧ポンプ 2の吐 出流量が増加するよう油圧ポンプ 2の傾転角を補正するためのものであり、 メモ リのテーブルには、 図 1 4に示すように、 ポンプ吐出圧力 Pが P 1以上のときは 補正値 S a = 0であり、 吐出圧力 Pが P 1より小さくなると、 吐出圧力 Pが小さ くなるに従って直線比例的に捕正値 S aが大きくなるように吐出圧力 Pと補正値 S aとの関係が設定されている。 The work implement controller 18 includes a first target pump tilt angle calculating section 81, a second target pump tilt angle calculating section 82, a first tilt angle correction value calculating section 83A, and a second Tilting angle correction value calculation section 83 B, 0 setting section 83 C, switching section 84 A, addition section 85, minimum value selection section 86, subtraction section 87, control current calculation Unit 8 has the functions of 8. The first and second tilt angle correction value calculators 8 3 A and 8 3 B receive the discharge pressure signal P of the hydraulic pump 2 from the pressure detector 14 and store this in the memory, respectively. Referring to the table, a correction value S for the second target tilt of the hydraulic pump 2 is calculated. The first tilt angle correction value calculator 8 3 A is configured to increase the discharge flow rate of the hydraulic pump 2 as the engine load becomes lighter, even if the engine speed decreases in the governor region 33 due to the reverse droop characteristic. This is for correcting the tilt angle of the hydraulic pump 2, and as shown in Fig. 14, when the pump discharge pressure P is P1 or more, the correction value S a = 0 in the memory table. The relationship between the discharge pressure P and the correction value Sa is set so that when the discharge pressure P becomes smaller than P1, the capture value Sa increases linearly in proportion to the discharge pressure P.
また、 第 2傾転角補正値演算部 8 3 Bは、 逆ドループ特性によりガバナ領域 3 3でのエンジン回転数が低下しても、 エンジン負荷に係わらず油圧ポンプ 2の吐 出流量が一定となるよう油圧ポンプ 2の傾転角を補正するためのものであり、 メ モリのテ一ブルには、 図 1 4に示すように、 ポンプ吐出圧力 Pが P 1以上のとき は補正値 S b = 0であり、 吐出圧力 Pが P 1より小さくなると、 第 1傾転角補正 値演算部 8 3 Aで演算される補正値 S aよりも小さい割合で、 吐出圧力 Pが小さ くなるに従つて直線比例的に補正値 S bが大きくなるように吐出圧力 Pと補正値 S bとの関係が設定されている。  In addition, the second tilt angle correction value calculation unit 83B determines that the discharge flow rate of the hydraulic pump 2 is constant irrespective of the engine load even when the engine speed in the governor region 33 decreases due to the reverse droop characteristic. This is for correcting the tilt angle of the hydraulic pump 2 so that the memory table contains a correction value S b when the pump discharge pressure P is equal to or higher than P1, as shown in Fig. 14. = 0, and when the discharge pressure P becomes smaller than P1, the discharge pressure P becomes smaller at a rate smaller than the correction value Sa calculated by the first tilt angle correction value calculator 83 A. Therefore, the relationship between the discharge pressure P and the correction value Sb is set so that the correction value Sb increases linearly.
0設定部 8 3 Cは、 補正値 Sとして 0を出力する。  The 0 setting unit 83C outputs 0 as the correction value S.
モード選択スィッチ 1 7 Aはダイヤル式であり、 第 1、 第 2、 第 3の 3つの切 換位置を有している。  The mode selection switch 17A is a dial type and has first, second and third switching positions.
スイッチング部 8 4 Aは、 モード選択スィッチ 1 7 Aが図示の第 1位置にある ときは図示の如く第 1傾転角補正値演算部 8 3 Aで演算された補正値 S aを選択 し、 モード選択スィッチ 1 7 Aが第 2位置に切り換えられると第 2傾転角補正値 演算部 8 3 Bで演算された補正値 S bを選択し、 モード選択スィッチ 1 7 Aが第 3位置に切り換えられると 0設定部 8 3 Cから出力された補正値 S (= 0 ) を選 択し、 それぞれ補正値 Sとして出力する。  When the mode selection switch 17A is at the first position shown in the figure, the switching unit 84A selects the correction value Sa calculated by the first tilt angle correction value calculation unit 83A as shown in the figure. When the mode selection switch 17A is switched to the second position, the correction value Sb calculated by the second tilt angle correction value calculator 83B is selected, and the mode selection switch 17A is switched to the third position. Then, the correction value S (= 0) output from the 0 setting unit 83 C is selected and output as the correction value S for each.
加算部 8 5では、 第 1の実施の形態と同様、 第 2目標ポンプ傾転角演算部 8 2 で演算された油圧ポンプ 2の第 2目標傾転 0 Tとスイッチング部 8 4 Aで選択し た補正値 Sを加算し、 補正された第 2目標傾転 を演算する。 In the adder 85, as in the first embodiment, the second target displacement 0 T of the hydraulic pump 2 calculated by the second target pump displacement angle calculator 82 is selected by the switching unit 84A. Then, the corrected second target tilt is calculated by adding the corrected value S.
図 1 5に、 加算部 85で補正されたポンプ吐出圧力 Pと第 2目標傾転 0 との 関係を示す。  FIG. 15 shows the relationship between the pump discharge pressure P corrected by the adding unit 85 and the second target displacement 0.
スイッチング部 84Aにおいて第 1傾転角補正値演算部 83 Aで演算された補 正値 S aを選択したとき、 ガバナ領域 33に相当する範囲 34における特性線 1 9は特性線 40のように補正され、 ポンプ吐出圧力 Pが P 1から Pminに低下す るに従い、 補正された第 2目標傾転 は第 1最大傾転 0maxlから第 4最大傾転 0max4 (=第 1最大傾転 0maxl+Samax) まで直線的に増大する。 この第 4最大 傾転 emax4は、 例えば油圧ポンプ 2の構造上の最大傾転 （ポンプ性能限界） に対 応して設定されている。  When the switching unit 84A selects the correction value Sa calculated by the first tilt angle correction value calculation unit 83A, the characteristic line 19 in the range 34 corresponding to the governor region 33 is corrected as shown by the characteristic line 40. As the pump discharge pressure P decreases from P1 to Pmin, the corrected second target tilt changes from the first maximum tilt 0maxl to the fourth maximum tilt 0max4 (= first maximum tilt 0maxl + Samax) Increase linearly to The fourth maximum displacement emax4 is set according to, for example, the maximum displacement (pump performance limit) in the structure of the hydraulic pump 2.
スイッチング部 84 Aにおいて第 2傾転角補正値演算部 83 Bで演算された補 正値 S bを選択したとき、 ガバナ領域 33に相当する範囲 34における特性線 1 9は特性線 41のように補正され、 ポンプ吐出圧力 Pが P 1から Pminに低下す るに従い、 補正された第 2目標傾転 θ Tは第 1最大傾転 Θ maxlから第 3最大傾転 0max3 (=第 1最大傾転 0maxl+Sbmax) まで直線的に増大する。  When the switching unit 84A selects the correction value Sb calculated by the second tilt angle correction value calculation unit 83B, the characteristic line 19 in the range 34 corresponding to the governor region 33 becomes like the characteristic line 41. As the pump discharge pressure P decreases from P1 to Pmin, the corrected second target tilt θ T becomes the first maximum tilt Θ maxl to the third maximum tilt 0max3 (= the first maximum tilt) 0maxl + Sbmax).
スイッチング部 84 Aにおいて 0設定部 83 Cの補正値 S==0が選択されたと き、 ガバナ領域 33に相当する範囲 34における特性線 1 9は補正されず、 第 2 目標ボンプ傾転角演算部 82で演算された第 2目標傾転 0 Tがそのまま出力され る。  When the correction value S == 0 of the 0 setting section 83 C is selected in the switching section 84 A, the characteristic line 19 in the range 34 corresponding to the governor area 33 is not corrected, and the second target pump tilt angle calculating section The second target tilt 0 T calculated in 82 is output as it is.
特性線 40が示す特性は、 図 1 2に示したメカニカルガバナにおけるドループ 特性線 3 1と見かけ上ほぼ一致し、 特性線 41が示す特性は、 図 3に示したアイ ソクロナス制御の特性線 32と見かけ上ほぼ一致する。  The characteristic indicated by the characteristic line 40 is apparently almost identical to the droop characteristic line 31 of the mechanical governor shown in FIG. 12, and the characteristic indicated by the characteristic line 41 is the same as the characteristic line 32 of the isochronous control shown in FIG. They almost match in appearance.
以上のように構成した本実施の形態における動作は、 エンジン 1が逆ドル一プ 特性に制御され、 油圧ポンプ 2の吐出流量増加制御が補正値 S aと補正値 S bの いずれかによりなされる点を除いて、 第 1の実施の形態と実質的に同じである。 つまり、 例えば重掘削等の作業時に操作レバー装置の操作レバーをフル操作し、 θϋ>θ c (=ΘΊ) で、 Ρ>Ρ 1であるとき、 モード切換スィッチ 17 Αを第 1 位置に切り換え、 第 1傾転角補正値演算部 83 Aで演算された補正値 S aが選択 されたときは、 図 15に示す特性線 40による油圧ポンプ 2の傾転角の増加制御 (吐出流量の増加制御） がなされ、 モード切換スィッチ 1 7 Aを第 2位置に切り 換え、 第 2傾転角補正値演算部 8 3 bで演算された補正値 S bが選択されたとき は、 図 1 5に示す特性線 4 1による油圧ポンプ 2の傾転角の増加制御 （吐出流量 の保持制御） がなされる。 In the operation according to the present embodiment configured as described above, the engine 1 is controlled to the reverse-drag characteristic, and the discharge flow rate increase control of the hydraulic pump 2 is performed by either the correction value Sa or the correction value Sb. Except for this point, it is substantially the same as the first embodiment. In other words, for example, during heavy excavation work, the operating lever of the operating lever device is fully operated, and when θϋ> θ c (= ΘΊ) and Ρ> Ρ1, the mode switching switch 17Α is switched to the first position, When the correction value Sa calculated by the first tilt angle correction value calculation unit 83A is selected, the tilt angle of the hydraulic pump 2 is increased by the characteristic line 40 shown in FIG. (Discharge flow rate increase control) is performed, the mode switch 17A is switched to the second position, and the correction value Sb calculated by the second tilt angle correction value calculator 83b is selected. The control for increasing the tilt angle of the hydraulic pump 2 (control for maintaining the discharge flow rate) is performed according to the characteristic line 41 shown in FIG.
図 1 6 A及び図 1 6 Bに、 ガバナ領域を逆ドループ特性に制御するエンジンを 有する従来技術によるポンプ吐出圧力 Pとポンプ傾転 0との関係及びポンプ吐出 圧力とボンプ吐出流量との関係を示す。  Figures 16A and 16B show the relationship between the pump discharge pressure P and the pump displacement 0 and the relationship between the pump discharge pressure and the pump discharge flow rate according to the conventional technology having an engine that controls the governor region to the reverse droop characteristic. Show.
前述したように、 作業機コントローラの演算機能に図 6に示した傾転角補正値 演算部 8 3、 スイッチング部 8 4及び加算部 8 5を備えていない場合は、 ガバナ 領域 3 3に相当する P minと P 1の間の範囲 2 3では直線 2 5で示すようにポン プ傾転 0は一定である。 一方、 逆ドループ特性では、 図 1 2の直線 3 4のように、 エンジント出力ルク （エンジン負荷） T eが低下するに従ってエンジン回転数 N が減少する。 このため P minと P 1の間の範囲 2 3では、 ポンプ吐出圧力 Pが P 1から低下するに従ってエンジン回転数 Nが減少するため、 ボンプ傾転 0が一定 であってもエンジン回転数 Nの減少によりポンプ吐出流量 Qは破線 4 4で示すよ うに減少する。 これにより油圧ァクチユエ一夕に供給される流量が少なくなり、 空荷動作での作業速度がアイソクロナス制御の場合よりも更に遅くなるという問 題がある。  As described above, when the calculation function of the work implement controller does not include the tilt angle correction value calculation unit 83, the switching unit 84, and the addition unit 85 shown in FIG. 6, it corresponds to the governor region 33. In the range 23 between P min and P 1, the pump displacement 0 is constant as shown by the straight line 25. On the other hand, in the reverse droop characteristic, the engine speed N decreases as the engine output torque (engine load) Te decreases, as indicated by the straight line 34 in FIG. Therefore, in the range 23 between P min and P 1, the engine speed N decreases as the pump discharge pressure P decreases from P 1, so that even if the pump displacement 0 is constant, the engine speed N Due to the decrease, the pump discharge flow rate Q decreases as shown by the broken line 44. As a result, there is a problem that the flow rate supplied to the hydraulic actuator is reduced, and the working speed in the unloading operation is lower than that in the case of the isochronous control.
図 1 7 A及び図 1 7 Bに、 本実施の形態によるポンプ吐出圧力 Pとポンプ傾転 Θとの関係及びポンプ吐出圧力とポンプ吐出流量との関係を示す。  FIG. 17A and FIG. 17B show the relationship between the pump discharge pressure P and the pump tilt angle and the relationship between the pump discharge pressure and the pump discharge flow rate according to the present embodiment.
本実施の形態では、 第 1傾転角補正値演算部 8 3 Aで演算された補正値 S aが 選択され、 図 1 5に示す特性線 1 9が特性線 4 0に補正される場合は、 ガバナ領 域 3 3に相当する P minと P 1の間の範囲 2 3において、 ポンプ傾転 0は図 1 5 の特性線 4 0に対応して直線 4 5のように変化し、 ボンプ吐出流量 Qはポンプ傾 転 Sの増加により直線 4 6で示すように変化する。 つまり、 逆ドループ特性によ りエンジン回転数 Nが低下しても、 ポンプ吐出圧力 Pが P 1から低下するに従つ てポンプ吐出流量 Qは直線比例的に増加する。 これにより図 1 0 A及び図 1 0 B に示した従来技術と同様、 油圧ァクチユエ一夕に供給される流量が多くなり、 空 荷動作での作業速度が速くなり、 作業能率を向上できる。 また、 第 2傾転角補正値演算部 8 3 Bで演算された補正値 S bが選択され、 図 1 5に示す特性線 1 9が特性線 4 1に補正される場合は、 ガバナ領域 3 3に相当 する Pminと P 1の間の範囲 2 3において、 ポンプ傾転 0は図 1 5の特性線 4 1 に対応して直線 4 7のように変化し、 ポンプ吐出流量 Qはポンプ傾転 Sの増加に より直線 4 8で示すようになる。 つまり、 逆ドル一プ特性によりエンジン回転数 Nが低下しても、 それによるポンプ吐出流量 Qの減少がポンプ傾転の増大により 相殺され、 ポンプ吐出流量 Qは一定に保たれるよう制御される。 これにより走行 操作、 吊荷作業、 整地作業のように油圧ポンプ 2の吐出流量の増加制御を望まな い操作或いは作業では、 油圧ァクチユエ一夕速度をエンジン負荷の増減に係わら ず等速度にし、 良好な走行操作、 吊荷作業、 整地作業を実施させることができる。 In the present embodiment, when the correction value Sa calculated by the first tilt angle correction value calculation unit 83 A is selected and the characteristic line 19 shown in FIG. 15 is corrected to the characteristic line 40, In the range 23 between Pmin and P1, which corresponds to the governor region 33, the pump displacement 0 changes as a straight line 45 corresponding to the characteristic line 40 in FIG. The flow rate Q changes as shown by the straight line 46 as the pump tilt S increases. That is, even if the engine speed N decreases due to the reverse droop characteristic, the pump discharge flow rate Q increases linearly as the pump discharge pressure P decreases from P1. As a result, similarly to the prior art shown in FIGS. 10A and 10B, the flow rate supplied to the hydraulic actuator is increased, the working speed in the unloading operation is increased, and the working efficiency can be improved. When the correction value Sb calculated by the second tilt angle correction value calculation unit 83B is selected and the characteristic line 19 shown in Fig. 15 is corrected to the characteristic line 41, the governor region 3 In the range 23 between Pmin and P1, which corresponds to 3, the pump displacement 0 changes as the straight line 47 corresponding to the characteristic line 41 in Fig. 15, and the pump discharge flow Q changes As S increases, it becomes as shown by the straight line 48. In other words, even if the engine speed N decreases due to the reverse droop characteristic, the decrease in the pump discharge flow Q due to the decrease is offset by the increase in the pump displacement, and the pump discharge flow Q is controlled to be kept constant. . As a result, in the operation or work in which it is not desired to increase the discharge flow rate of the hydraulic pump 2 such as traveling operation, hanging load operation, and ground leveling operation, the hydraulic actuator speed is kept constant regardless of the increase or decrease of the engine load. It is possible to carry out simple traveling operation, hanging load work and terrain work.
0設定部 8 3 Cの補正値 S = 0が選択され、 図 1 5に示す特性線 1 9が補正さ れない場合は、 ガバナ領域 3 3に相当する Pminと P 1の間の範囲 2 3において、 ポンプ傾転 0は図 1 5の特性線 1 9に対応して直線 4 9のように一定となり、 ポ ンプ吐出流量 Qは、 図 1 6 Bと同様、 逆ドループ特性によるエンジン回転数 Nの 低下によりポンプ吐出流量 Qは直線 5 0のように減少する。 これにより燃費を更 に向上させることができる。  0 Setting section 8 3 C If the correction value S = 0 is selected and the characteristic line 19 shown in Fig. 15 is not corrected, the range between Pmin and P 1 corresponding to the governor area 3 3 2 3 , The pump displacement 0 becomes constant as shown by the straight line 49 corresponding to the characteristic line 19 in FIG. 15, and the pump discharge flow rate Q becomes the engine speed N due to the reverse droop characteristic, as in FIG. 16B. , The pump discharge flow rate Q decreases as indicated by the straight line 50. This can further improve fuel efficiency.
以上のように構成した本実施の形態によっても、 逆ドル一プ特性に制御される エンジンを備えた油圧駆動装置において、 第 1の実施の形態と同様の効果が得ら れる。 つまり、 モード切換スィッチ 1 7 Aを第 1位置に切り換え、 第 1傾転角補 正値演算部 8 3 Aで演算された補正値 S aを選択することにより、 ガバナ領域 3 3にあってもエンジン負荷が軽くなるに従ってポンプ吐出流量 Qを次第に増加さ せることができ、 メカニカルガバナにおけるドループ特性による流量の増加とほ ぼ同等のポンプ吐出流量の増加を実施させることができる。 これによつてェンジ ン軽負荷時の油圧ァクチユエ一夕速度を増速させることができ、 空荷作業等の軽 負荷時の作業能率を向上させることができる。 また、 メカニカルガバナ式ェンジ ン 1を備えた作業機の操作に慣れたオペレータに対しても良好な操作フィーリン グを与えることができる。  According to the present embodiment configured as described above, the same effects as those of the first embodiment can be obtained in the hydraulic drive device including the engine controlled to the reverse dollar characteristic. In other words, by switching the mode switch 17A to the first position and selecting the correction value Sa calculated by the first tilt angle correction value calculator 83A, even in the governor region 33, As the engine load becomes lighter, the pump discharge flow rate Q can be gradually increased, and the increase in the flow rate due to the droop characteristic in the mechanical governor can be almost as large as the increase in the pump discharge flow rate. As a result, it is possible to increase the speed of the hydraulic actuator when the engine is lightly loaded, and to improve the work efficiency at the time of light load such as unloading work. In addition, it is possible to provide a good operation feeling to an operator who is accustomed to operating a working machine equipped with the mechanical governor engine 1.
また、 走行操作、 吊荷作業、 整地作業が実施されるときには、 モード切換スィ ツチ 1 7 Aを第 2位置に切り換え、 第 2傾転角補正値演算部 8 3 Bで演算された 補正値 S bを選択することにより、 エンジン負荷に係わらず油圧ポンプ 2の吐出 流量は一定となり、 油圧ァクチユエ一夕速度をエンジン負荷の増減に係わらず等 速度にし、 良好な走行操作、 吊荷作業、 整地作業を実施させることができる。 また、 本実施の形態によれば、 逆ドル一プ特性に制御されるエンジンを用い、 油圧ポンプ 2を駆動するので、 ァイソクロナス特性に制御されるエンジンを用い た第 1の実施の形態に比べ軽負荷時のエンジン回転数を更に低下させ、 更なる低 燃費と低騒音を実現することができる。 When the traveling operation, the lifting work, and the terrain work are performed, the mode switching switch 17A is switched to the second position, and the mode is calculated by the second tilt angle correction value calculator 83B. By selecting the correction value Sb, the discharge flow rate of the hydraulic pump 2 becomes constant irrespective of the engine load, the speed of the hydraulic actuator is set to the same speed regardless of the increase or decrease of the engine load, and good running operation and lifting , Leveling work can be performed. Further, according to the present embodiment, since the hydraulic pump 2 is driven by using the engine controlled to the reverse drive characteristic, it is lighter than the first embodiment using the engine controlled to the isochronous characteristic. The engine speed at the time of load can be further reduced, and further low fuel consumption and low noise can be realized.
また、 軽掘削時に燃費を最優先させたい場合は、 モード切換スィッチ 1 7 Aを 第 3位置に切り換え、 0設定部 8 3 Cの設定値 S = 0を選択することにより、 油 圧ポンプ 2の吐出流量は減少し、 更に燃費を向上させることができる。  In addition, if the highest priority is given to fuel efficiency during light excavation, switch the mode switch 17A to the third position and select the set value S = 0 in the 0 setting section 83C to set the hydraulic pump 2 The discharge flow rate is reduced, and the fuel efficiency can be further improved.
本発明の第 3の実施の形態を図 1 8〜図 2 0により説明する。  A third embodiment of the present invention will be described with reference to FIGS.
以上の実施の形態では、 ガバナ領域をアイソクロナス特性或いは逆ドループ特 性に制御するエンジンを備えた油圧駆動装置に本発明を適用したが、 ガバナ領域 の特性はそれに限定されるものではない。 本実施の形態は、 その一例として、 ガ バナ領域をァイソクロナス特性と逆ドループ特性を組み合わせた特性に制御され るエンジンを備えたものに本発明を適用したものである。  In the above embodiment, the present invention is applied to the hydraulic drive device including the engine that controls the governor region with isochronous characteristics or reverse droop characteristics. However, the characteristics of the governor region are not limited thereto. In the present embodiment, as an example, the present invention is applied to an engine having an engine whose governor region is controlled to a combination of the isochronous characteristic and the reverse droop characteristic.
図 1 8にァイソクロナス特性と逆ドループ特性を組み合わせた特性に制御され るエンジンの回転数 Nと出力トルク T eとの関係を示す。 図 1 8において、 ガバ ナ領域 3 3では、 直線 9 0 aのように、 エンジン出力トルク T e (エンジン負 荷） の低下に係わらずエンジン回転数 Nを定格回転数 NOの一定値に保つアイソ クロナス特性と、 直線 9 O bのように、 エンジン出力トルク T eが低下するに従 つてエンジン回転数 Nが減少する逆ドループ特性とを組み合わせた特性 9 0を有 している。 この特性 9 0により、 中負荷時にはァイソクロナス特性によりェンジ ン回転数を一定に保ち、 ある程度のァクチユエ一夕速度を確保しかつ騒音及び燃 費を向上させ、 エンジン負荷がそれより小さい軽負荷時には、 逆ドループ特性に より更なる騒音及び燃費の向上が可能となる。  Figure 18 shows the relationship between the engine speed N and the output torque Te controlled to a combination of isochronous characteristics and reverse droop characteristics. In Fig. 18, in the governor region 33, as shown by a straight line 90a, an iso-rotation speed N is maintained at a constant value of the rated speed NO regardless of a decrease in the engine output torque Te (engine load). It has a characteristic 90 that combines a chronous characteristic and a reverse droop characteristic in which the engine speed N decreases as the engine output torque Te decreases, as shown by a straight line 90 Ob. Due to this characteristic 90, the engine speed is kept constant by the isochronous characteristic at the time of medium load, a certain speed of the engine is secured, noise and fuel consumption are improved, and when the engine load is lighter, the reverse Droop characteristics can further improve noise and fuel efficiency.
図 1 9は、 エンジンがそのような特性 9 0を有する場合の傾転角補正値演算部 8 3 (図 6参照） におけるポンプ傾転補正値 Sの特性を示す図である。 ポンプ傾 転補正値 Sの特性は、 図 1 8に示す直線 9 0 aと直線 9 0 bの特性に対応して折 れ線に設定されている。 FIG. 19 is a diagram showing the characteristics of the pump tilt correction value S in the tilt angle correction value calculation unit 83 (see FIG. 6) when the engine has such a characteristic 90. The characteristics of the pump displacement correction value S are plotted according to the characteristics of the straight lines 90a and 90b shown in Fig. 18. Line is set.
図 2 0は、 傾転角補正値演算部 8 3の補正値 Sが図 1 9に示すような特性を有 する場合の図 9と同様な特性図である。 第 2目標傾転 に補正値 Sを加算する ことにより、 特性線 1 9は特性線 9 1のように補正値 Sの折れ線と同様の折れ線 の特性に補正される。 これにより重掘削等、 油圧ポンプ 2の傾転角が第 2目標傾 転 に制限されるような作業では、 ガパナ領域 3 3に相当する Pminと P 1の間 の範囲 2 3において、 ポンプ傾転 0は特性線 9 1のように変化し、 これに伴い油 圧ポンプの吐出流量は図 1 1 Bの直線 3 6のように変化し、 第 1の実施の形態と 同様のポンプ吐出流量の増加制御が行うことができる。  FIG. 20 is a characteristic diagram similar to FIG. 9 in a case where the correction value S of the tilt angle correction value calculating section 83 has the characteristic shown in FIG. By adding the correction value S to the second target displacement, the characteristic line 19 is corrected to the characteristic of the polygonal line similar to the polygonal line of the correction value S like the characteristic line 91. In the work where the tilt angle of the hydraulic pump 2 is limited to the second target tilt such as heavy excavation, the pump tilt in the range 23 between Pmin and P 1 corresponding to the governor region 33 is performed. 0 changes as indicated by the characteristic line 91, and accordingly, the discharge flow rate of the hydraulic pump changes as indicated by the straight line 36 in FIG. 11B, and the pump discharge flow rate increases as in the first embodiment. Control can be performed.
なお、 上記の実施の形態においては、 ポンプ吐出圧力 Pが P 1以下となったェ ンジン軽負荷時にポンプ吐出流量を増加させる補正値 Sの特性として、 メカ二力 ルガバナにおけるドループ特性にほぼ一致するポンプ吐出流量の増加制御が行え るものを設定したが、 本発明は、 このような吐出流量特性に設定することには限 られない。 例えば、 図 8に示すポンプ傾転補正値 Sの特性線の傾きをより大きく し、 ドループ特性によるボンプ吐出流量の増加より多くボンプ吐出流量が増加す るようにしても良いし、 その逆であってもよい。 また、 ガバナ領域の特性がアイ ソクロナス特性と逆ドループ特性を組み合わせたものでない場合でも、 図 8に示 すポンプ傾転補正値 Sの特性線を折れ線にしてもよい。 更に、 ポンプ傾転補正値 Sの特性線は直線でなく、 曲線であってもよい。  In the above embodiment, the characteristic of the correction value S for increasing the pump discharge flow rate at the time of the engine light load in which the pump discharge pressure P is equal to or less than P1 substantially matches the droop characteristic in the mechanical governor. Although a pump capable of increasing the discharge flow rate of the pump is set, the present invention is not limited to the setting of such discharge flow characteristics. For example, the slope of the characteristic line of the pump displacement correction value S shown in FIG. 8 may be made larger so that the pump discharge flow rate increases more than the pump discharge flow rate increases due to the droop characteristic, and vice versa. You may. Further, even when the governor region characteristic is not a combination of the isochronous characteristic and the reverse droop characteristic, the characteristic line of the pump displacement correction value S shown in FIG. 8 may be a broken line. Further, the characteristic line of the pump displacement correction value S may be a curve instead of a straight line.
更に、 上記実施の形態では、 補正値 Sを 0とするポンプ吐出圧力をポンプ吸収 トルク曲線 2 0による制御の開始圧力である P 1に一致させたがそれよりも低い 圧力であってもよい。  Further, in the above-described embodiment, the pump discharge pressure at which the correction value S is 0 is made equal to P1, which is the start pressure of the control based on the pump absorption torque curve 20, but may be a lower pressure.
また、 上記の実施の形態では、 ポンプ吐出圧力 Pが P 1以下となったエンジン 軽負荷時にポンプ吐出流量を増加させる補正値 Sの特性としてドル一プ特性に対 応する 1つの特性を設定したが、 それ以外に 1つ或いは複数の特性を設定し、 ォ ペレ一夕がモード選択スィッチの切り換えによりそのうちの 1つを選択できるよ うにしてもよい。 また、 モード選択スィッチを出力を連続的に変化させるダイヤ ル式とし、 補正値 Sの特性を連続的変えれるようにしてもよい。 これによりアイ ソクロナス特性或いは逆ドループ特性のメリットである低燃費と低騒音の効果を 維持したまま、 作業機に複数の動作性能を持たせ、 オペレータの望む動作スピー ドをオペレータ自身で選択できるようになる。 Further, in the above embodiment, one characteristic corresponding to the dollar-pull characteristic is set as the characteristic of the correction value S for increasing the pump discharge flow rate at the time of light load of the engine in which the pump discharge pressure P becomes P1 or less. However, one or more other characteristics may be set so that the operator can select one of them by switching the mode selection switch. Further, the mode selection switch may be a dial type that continuously changes the output so that the characteristic of the correction value S can be continuously changed. As a result, the advantages of low fuel consumption and low noise, which are the advantages of the isochronous characteristic or reverse droop characteristic, can be achieved. While maintaining this, the work equipment can be provided with multiple operation performances, and the operator can select the operation speed desired by the operator himself.
また、 上記実施の形態では、 ァイソクロナス特性或いは逆ドル一プ特性に制御 可能な燃料噴射制御装置のァクチユエ一夕部分を電子ガバナ 1 2としたが、 本発 明はこれに限るものではなく、 エンジン回転数に関係なく噴射量の制御が可能な コモンレール式燃料噴射制御装置やュニットインジ： ϋクタ式燃料噴射制御装置を 設けてもよい。  Further, in the above-described embodiment, the electronic governor 12 is used as the electronic governor 12 in the fuel injection control device capable of controlling to the isochronous characteristic or the reverse dollar characteristic. However, the present invention is not limited to this. A common rail fuel injection control device or a unit fuel injector control device capable of controlling the injection amount regardless of the rotation speed may be provided.
また、 上記実施の形態では、 要求流量に応じた油圧ポンプ 2の傾転角の制御、 油圧ポンプ 2の吸収トルク制御 （吸収馬力制御）、 本発明の特徴である軽負荷時 の油圧ポンプの傾転角増加制御の指令値の演算を全て作業機コントローラ 1 8に より行い、 制御電流信号をレギユレ一夕 1 6に送ることで油圧ポンプの傾転角を 制御したが、 それらの一部 （例えば要求流量に応じた油圧ポンプ 2の傾転角の制 御や油圧ポンプ 2の吸収トルク制御 （吸収馬力制御)） をレギユレ一夕により油 圧的に行ってもよい。 更に、 上記実施の形態では、 傾転角検出器 1 5により油圧 ポンプ 2の傾転角を検出し、 その傾転角が目標傾転角に一致するようフィードバ ックル一プにより制御したが、 傾転角検出器 1 5を設けず、 オープンループにて 油圧ポンプの傾転角を制御してもよい。 産業上の利用可能性  Further, in the above embodiment, the tilt angle control of the hydraulic pump 2 according to the required flow rate, the absorption torque control of the hydraulic pump 2 (absorption horsepower control), the tilting of the hydraulic pump at light load, which is a feature of the present invention, are described. All the calculation of the command value of the turning angle increase control was performed by the work equipment controller 18 and the control current signal was sent to the Regulayer 16 to control the tilt angle of the hydraulic pump. The control of the tilt angle of the hydraulic pump 2 according to the required flow rate and the control of the absorbing torque of the hydraulic pump 2 (absorbing horsepower control) may be performed hydraulically by a regi- ration. Further, in the above-described embodiment, the tilt angle of the hydraulic pump 2 is detected by the tilt angle detector 15 and the tilt angle is controlled by the feedback buckle so that the tilt angle matches the target tilt angle. The displacement angle of the hydraulic pump may be controlled in an open loop without providing the displacement angle detector 15. Industrial applicability
本発明によれば、 ガバナ領域の少なくとも一部をァイソクロナス特性、 逆ドル ープ特性、 ァイソクロナス特性と逆ドループ特性を組み合わせた特性のいずれか に制御可能なエンジンを備えた油圧駆動装置において、 ガバナ領域にあってもェ ンジン負荷が軽くなるに従つて油圧ポンプの吐出流量を増加させることができ、 エンジン軽負荷時の油圧ァクチユエ一夕速度をメカニカルガバナ式エンジンを備 えたものと同様に増速させることができ、 軽負荷時の作業能率を向上させること ができる。  According to the present invention, there is provided a hydraulic drive apparatus including an engine capable of controlling at least a part of a governor region to one of an isochronous characteristic, a reverse droop characteristic, and a combination of the isochronous characteristic and the reverse droop characteristic. Even when the engine load becomes lighter, the discharge flow rate of the hydraulic pump can be increased as the engine load becomes lighter, and the speed of the hydraulic actuator at light engine load is increased in the same way as with a mechanical governor-type engine. It is possible to improve work efficiency at light load.
また、 メカニカルガバナ式エンジンを備えた作業機の操作に慣れたオペレータ に対しても良好な操作フィーリングを与えることができる。  Also, a good operation feeling can be given to an operator who is accustomed to the operation of a working machine equipped with a mechanical governor type engine.
また、 本発明によれば、 選択的に油庄ポンプの吐出流量が一定となる制御を実 施し、 油圧ァクチユエ一夕速度をエンジン負荷の増減に係わらず等速度にし、 ォ ペレ一夕の望む操作或いは作業を良好に実施させることができる。 Further, according to the present invention, control for selectively keeping the discharge flow rate of the oil pump constant is realized. In this way, the operating speed of the hydraulic actuator can be kept constant irrespective of the increase or decrease of the engine load, and the operation or operation desired by the operating can be performed well.

Claims

請求の範囲 The scope of the claims
1 . ガバナ領域の少なくとも一部をァイソクロナス特性、 逆ドループ特性、 アイ ソクロナス特性と逆ドループ特性を組み合わせた特性のいずれかに制御可能な燃 料噴射制御装置（12, 13)を有するエンジン（1)と、 1. An engine (1) that has a fuel injection control device (12, 13) that can control at least a part of the governor region to one of isochronous characteristics, reverse droop characteristics, and a combination of isochronous characteristics and reverse droop characteristics. When,
このエンジンにより駆動される可変容量型の油圧ポンプ (2)と、  A variable displacement hydraulic pump (2) driven by this engine,
この油圧ポンプから吐出される圧油によって駆動する複数の油圧ァクチユエ一 夕（3-6)とを備える作業機の油圧駆動装置において、  In a hydraulic drive device for a working machine, comprising a plurality of hydraulic actuators (3-6) driven by hydraulic oil discharged from the hydraulic pump,
上記油圧ポンプ（2)の吐出圧力が第 1所定圧力（P1)を越えると油圧ポンプの押 しのけ容積が予め設定されたポンプ吸収トルク曲線 (20)により定まる値を越えな いよう上記油圧ポンプの押しのけ容積を制御するポンプ吸収トルク制御手段（8 When the discharge pressure of the hydraulic pump (2) exceeds the first predetermined pressure (P1), the hydraulic pressure is adjusted so that the displacement of the hydraulic pump does not exceed a value determined by a preset pump absorption torque curve (20). Pump absorption torque control means for controlling the displacement of the pump (8
2)と、 2) and
上記油圧ポンプの吐出圧力が上記第 1所定圧力（Π)以下にあるとき、 油圧ボン プの吐出圧力が第 2所定圧力（P 1 )から低くなるに従つて油圧ポンプの押しのけ容 積が増加するよう制御する流量補正制御手段（83， 85； 17A, 83A， 83B, 84A， 85)とを備 えることを特徴とする作業機の油圧駆動装置。  When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure (Π), the displacement of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (P 1). A hydraulic drive device for a working machine, comprising flow rate correction control means (83, 85; 17A, 83A, 83B, 84A, 85) for controlling the flow rate.
2 . ガバナ領域の少なくとも一部をァイソクロナス特性、 逆ドル一プ特性、 アイ ソクロナス特性と逆ドループ特性を組み合わせた特性のいずれかに制御可能な燃 料噴射制御装置（12， 13)を有するエンジン（1)と、 2. An engine (12, 13) that has a fuel injection control device (12, 13) that can control at least a part of the governor region to one of isochronous characteristics, reverse dollar characteristics, and a combination of isochronous characteristics and reverse droop characteristics. 1) and
このエンジンにより駆動される可変容量型の油圧ポンプ (2)と、  A variable displacement hydraulic pump (2) driven by this engine,
この油圧ポンプから吐出される圧油によって駆動する複数の油圧ァクチユエ一 夕（3- 6)とを備える作業機の油圧駆動装置において、  In a hydraulic drive device for a working machine, comprising a plurality of hydraulic actuators (3-6) driven by hydraulic oil discharged from the hydraulic pump,
上記油圧ポンプ（2)の押し除け容積を制御するレギユレ一夕（16)と、 上記油圧ポンプの吐出圧力を検出する圧力検出器（14)と、  A regulator (16) for controlling the displacement of the hydraulic pump (2), a pressure detector (14) for detecting a discharge pressure of the hydraulic pump,
この圧力検出器により検出された上記油圧ポンプの吐出圧力が第 1所定圧力 The discharge pressure of the hydraulic pump detected by the pressure detector is equal to the first predetermined pressure.
(P1)を越えると油圧ポンプの押しのけ容積が予め設定されたポンプ吸収トルク曲 線 (20)により定まる値を越えないよう上記レギユレ一夕（16)を制御するポンプ吸 収トルク制御手段（82)と、 上記油圧ポンプの吐出圧力が上記第 1所定圧力（P1)以下にあるとき、 油圧ボン プの吐出圧力が第 2所定圧力（P1)から低くなるに従つて油圧ポンプの押し除け容 積が増加するよう上記レギユレ一夕（16)を制御する流量補正制御手段 (83, 85； Π A, 83A, 83B, 84A, 85)とを備えることを特徴とする作業機の油圧駆動装置。 (P1) The pump absorption torque control means (82) that controls the above-mentioned regulation (16) so that the displacement of the hydraulic pump does not exceed the value determined by the preset pump absorption torque curve (20). When, When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure (P1), the displacement capacity of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (P1). A hydraulic drive device for a working machine, comprising: flow rate correction control means (83, 85; 83A, 83A, 83B, 84A, 85) for controlling the above-mentioned regulation (16).
3 . 請求項 1又は 2記載の作業機の油圧駆動装置において、 3. The hydraulic drive device for a working machine according to claim 1 or 2,
上記第 2所定圧力（P1)は上記第 1所定圧力（P1)に一致していることを特徴とす る作業機の油圧駆動装置。  The hydraulic drive device for a working machine, wherein the second predetermined pressure (P1) is equal to the first predetermined pressure (P1).
4 . 請求項 1又は 2記載の作業機の油圧駆動装置において、 4. The hydraulic drive for a working machine according to claim 1 or 2,
上記流量補正制御手段（83, 85 ; 17A, 83A, 83B, 84A, 85)による上記油圧ポンプの押 し除け容積の増加制御を無効にする制御解除手段（17， 84 ; 17A, 830を更に備える ことを特徴とする作業機の油圧駆動装置。  The system further includes control release means (17, 84; 17A, 830) for invalidating the increase control of the displacement of the hydraulic pump by the flow rate correction control means (83, 85; 17A, 83A, 83B, 84A, 85). A hydraulic drive device for a working machine, characterized in that:
5 . 請求項 4記載の作業機の油圧駆動装置において、 5. The hydraulic drive for a working machine according to claim 4,
上記燃料噴射制御装置（12, 13)は、 ガバナ領域の少なくとも一部をァイソクロ ナス特性に制御可能なものであり、  The fuel injection control device (12, 13) can control at least a part of the governor region to have an isochronous characteristic.
上記制御解除手段（17, 84)は、 走行モードスィッチ（17a)、 吊荷モードスィッチ (17b) , 整地モードスィッチ（17c)の少なくとも 1つを含むことを特徴とする作業 機の油圧駆動装置。  The hydraulic drive device for a working machine, wherein the control release means (17, 84) includes at least one of a traveling mode switch (17a), a suspended load mode switch (17b), and a terrain mode switch (17c).
6 . 請求項 1又は 2記載の作業機の油圧駆動装置において、 6. The hydraulic drive for a working machine according to claim 1 or 2,
上記流量補正制御手段（83, 85； 17A, 83A, 84A, 85)は、 上記油圧ポンプ (2)の吐出 圧力が上記第 2所定圧力（P1)から低くなるに従つて上記油圧ポンプの吐出流量が 増加するよう上記油圧ポンプの押しのけ容積を制御することを特徴とする作業機 の油圧駆動装置。  The flow rate correction control means (83, 85; 17A, 83A, 84A, 85) adjusts the discharge flow rate of the hydraulic pump as the discharge pressure of the hydraulic pump (2) decreases from the second predetermined pressure (P1). A hydraulic drive device for a working machine, wherein the displacement of the hydraulic pump is controlled so as to increase the hydraulic pressure.
7 . 請求項 1又は 2記載の作業機の油圧駆動装置において、 7. The hydraulic drive for a working machine according to claim 1 or 2,
上記燃料噴射制御装置 α 2, 13)は、 ガバナ領域の少なくとも一部を逆ドル一プ 特性に制御可能なものであり、 The above-mentioned fuel injection control device α 2, 13) It can be controlled by characteristics,
上記流量補正制御手段（17A， 83A, 83B, 84A, 85)は、 上記油圧ポンプ（2)の吐出圧 力が上記第 2所定圧力 (Π)から低くなるに従つて上記油圧ポンプの吐出流量が増 加するよう上記油圧ポンプの押しのけ容積を制御する第 1手段 (83A, 85)と、 上記 油圧ポンプの吐出圧力が上記第 2所定圧力（Π)から低くなるときに上記油圧ボン プの吐出流量が一定に保たれるよう上記油圧ポンプの押しのけ容積を制御する第 2手段 (83B, 85)と、 上記第 1手段と第 2手段の一方を選択する選択手段（17A, 84 A)とを有することを特徴とする作業機の油圧駆動装置。  The flow rate correction control means (17A, 83A, 83B, 84A, 85) adjusts the discharge flow rate of the hydraulic pump as the discharge pressure of the hydraulic pump (2) decreases from the second predetermined pressure (Π). First means (83A, 85) for controlling the displacement of the hydraulic pump so as to increase, and the discharge flow rate of the hydraulic pump when the discharge pressure of the hydraulic pump becomes lower than the second predetermined pressure (Π) (83B, 85) for controlling the displacement of the hydraulic pump so that the pressure is kept constant, and selecting means (17A, 84A) for selecting one of the first means and the second means. A hydraulic drive device for a working machine, characterized in that:
8 . 請求項 7記載の作業機の油圧駆動装置において、 8. The hydraulic drive for a working machine according to claim 7,
上記流量補正制御手段（17A, 83A, 83B, 84A, 85)は、 更に、 上記油圧ポンプ（2)の 押し除け容積の増加制御を無効にする第 3手段 (83C)を更に有し、 上記選択手段 (17A, 84A)は、 上記第 1手段と第 2手段と第 3手段のいずれか 1つを選択するも のであることを特徴とする作業機の油圧駆動装置。  The flow rate correction control means (17A, 83A, 83B, 84A, 85) further includes a third means (83C) for disabling the control for increasing the displacement volume of the hydraulic pump (2). The means (17A, 84A) selects one of the first means, the second means, and the third means, and is a hydraulic drive device for a working machine.
9 . 請求項 1又は 2記載の作業機の油圧駆動装置において、 9. The hydraulic drive for a working machine according to claim 1 or 2,
上記ボンプ吸収トルク制御手段（82)は、 上記油圧ポンプ (2)の吐出圧力とボン プ吸収トルク曲線とからポンプ吸収トルク制御のための目標押しのけ容積（ T) を演算するとともに、 上記油圧ポンプの吐出圧力が上記第 1所定圧力 (P1)以下に あるときに前記目標押しのけ容積を一定値（ 0 maxl)に保持する手段 (82)を有し、 上記流量補正制御手段（83, 85 ; 17A, 83A, 83B, 84A, 85)は、 上記油圧ポンプの吐出 圧力が上記第 2所定圧力 (P1)から低くなるに従つて増加する押しのけ容積補正値 (S)を演算する手段 (83 ; 83A，83B)と、 上記目標押しのけ容積に前記押しのけ容積 補正値を加算し補正された第 2押しのけ容積（ θ T)を演算する手段とを有し、 こ の補正された目標押しのけ容積により上記油圧ポンプの押しのけ容積を制御する ことを特徴とする作業機の油圧駆動装置。  The pump absorption torque control means (82) calculates a target displacement (T) for pump absorption torque control from the discharge pressure of the hydraulic pump (2) and the pump absorption torque curve, Means for holding the target displacement at a constant value (0 maxl) when the discharge pressure is equal to or lower than the first predetermined pressure (P1); and the flow rate correction control means (83, 85; 17A, 83A, 83B, 84A, 85) are means (83; 83A, 83B) for calculating a displacement correction value (S) that increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (P1). ), And means for calculating the corrected second displacement (θ T) by adding the displacement correction value to the target displacement, and displacing the hydraulic pump by the corrected target displacement. Working machine oil characterized by controlling the volume Drive.
1 0 . 請求項 1又は 2記載の作業機の油圧駆動装置において、 10. The hydraulic drive device for a working machine according to claim 1 or 2,
上記ポンプ吸収トルク制御手段（82)は、 上記油圧ポンプ (2)の押しのけ容積の 最大値を上記ポンプ吸収トルク曲線 (20)により定まる値以下に制限する手段であ Ό、 The pump absorption torque control means (82) is used to control the displacement of the hydraulic pump (2). A means for limiting the maximum value to a value determined by the pump absorption torque curve (20) or less,
上記流量補正制御手段（83, 85 ; 17Α, 83Α, 83Β, 84Α, 85)は、 上記油圧ポンプの吐出 圧力が第 2所定圧力から低くなるに従って上記油圧ポンプの押しのけ容積の最大 値が増加するよう制御する手段であることを特徴とする作業機の油圧駆動装置。  The flow rate correction control means (83, 85; 17 °, 83 °, 83 °, 84 °, 85) controls the maximum value of the displacement of the hydraulic pump to increase as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure. A hydraulic drive device for a working machine, which is a control unit.
1 1 . 請求項 1又は 2記載の作業機の油圧駆動装置において、 11. The hydraulic drive device for a working machine according to claim 1 or 2,
上記複数の油圧ァクチユエ一夕（3-6)の要求流量に応じた第 1目標押しのけ容 積（ 0 D)を演算する第 1演算手段 (81)を更に備え、  A first calculating means (81) for calculating a first target displacement volume (0D) according to a required flow rate of the plurality of hydraulic factories (3-6);
上記ポンプ吸収トルク制御手段 (82)は、 上記油圧ポンプ (2)の吐出圧力とボン プ吸収トルク曲線（20)とからポンプ吸収トルク制御のための第 2目標押しのけ容 積 (.0 Τ)を演算するとともに、 上記油圧ポンプの吐出圧力が上記第 1所定圧力（Ρ 1)以下にあるときに前記目標押しのけ容積を一定値（ 0 maxl)に保持する第 2演算 手段 (82)を有し、  The pump absorption torque control means (82) determines the second target displacement (.0 °) for pump absorption torque control from the discharge pressure of the hydraulic pump (2) and the pump absorption torque curve (20). A second calculating means (82) for calculating and maintaining the target displacement at a constant value (0 maxl) when the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure (Ρ1);
上記流量補正制御手段（83, 85 ; 17A, 83A, 83B, 84A, 85)は、 上記油圧ポンプの吐出 圧力が上記第 2所定圧力（P1)から低くなるに従つて増加する押しのけ容積補正値 (S)を演算する手段（82； 83A, 83B)と、 前記第 2目標押しのけ容積に前記押しのけ 容積補正値を加算し補正された第 2目標押しのけ容積（ θ T)を演算する手段（85) とを有し、  The flow rate correction control means (83, 85; 17A, 83A, 83B, 84A, 85) includes a displacement correction value (A) that increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (P1). Means (82; 83A, 83B) for calculating S), and means (85) for calculating the corrected second target displacement (θ T) by adding the displacement correction value to the second target displacement. Has,
前記第 1目標押しのけ容積と前記補正された第 2目標押しのけ容積の小さな方 を制御用の目標押しのけ容積として選択し、 上記油圧ポンプの押しのけ容積を制 御することを特徴とする作業機の油圧駆動装置。  Selecting the smaller of the first target displacement and the corrected second target displacement as a target displacement for control, and controlling the displacement of the hydraulic pump. apparatus.
1 2 . ガバナ領域の少なくとも一部をァイソクロナス特性、 逆ドループ特性、 ァ ィソクロナス特性と逆ドループ特性を組み合わせた特性のいずれかに制御可能な 燃料噴射制御装置（12， 13)を有するエンジン（1)と、 1 2. An engine (1) having a fuel injection control device (12, 13) capable of controlling at least a part of the governor region to one of an isochronous characteristic, a reverse droop characteristic, and a combination of the isochronous characteristic and the reverse droop characteristic. When,
このエンジンにより駆動される可変容量型の油圧ポンプ (2)と、  A variable displacement hydraulic pump (2) driven by this engine,
この油圧ポンプから吐出される圧油によって駆動する複数の油圧ァクチユエ一 タ（3-6)とを備える作業機の油圧駆動方法において、 上記油圧ポンプ (2)の吐出圧力が第 1所定圧力（Π)を越えるときは、 油圧ボン プの押しのけ容積が予め設定されたポンプ吸収トルク曲線（20)により定まる値を 越えないよう上記油圧ポンプの押しのけ容積を制御し、 A hydraulic drive method for a working machine including a plurality of hydraulic actuators (3-6) driven by hydraulic oil discharged from the hydraulic pump, When the discharge pressure of the hydraulic pump (2) exceeds the first predetermined pressure (Π), the hydraulic pump should be moved so that the displacement of the hydraulic pump does not exceed the value determined by the preset pump absorption torque curve (20). Control the displacement of
上記油圧ポンプの吐出圧力が上記第 1所定圧力（Π)以下にあるときは、 油圧ポ ンプの吐出圧力が第 2所定圧力（Π)から低くなるに従って油圧ポンプの押しのけ 容積が増加するよう制御することを特徴とする作業機の油圧駆動方法。  When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure (Π), the displacement of the hydraulic pump is controlled to increase as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (Π). A hydraulic drive method for a working machine, characterized in that:
1 3 . 請求項 1 2記載の作業機の油圧駆動方法において、 13. The hydraulic drive method for a working machine according to claim 12,
上記油圧ポンプの吐出圧力が上記第 1所定圧力 (P1)以下にあるときは、 油圧ポ ンプの吐出圧力が第 2所定圧力（P1)から低くなるに従って油圧ポンプの押しのけ 容積が増加させる制御と、 油圧ポンプの押しのけ容積を一定に保つ制御のいずれ か一方を選択可能であるであることを特徴とする作業機の油圧駆動方法。  When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure (P1), control is performed to increase the displacement of the hydraulic pump as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (P1); A hydraulic drive method for a working machine, wherein either one of control for keeping a displacement of a hydraulic pump constant can be selected.
1 . 請求項 1 2記載の作業機の油圧駆動方法において、 1. The hydraulic drive method for a working machine according to claim 12,
上記油圧ポンプの吐出圧力が上記第 1所定圧力 (P1)以下にあるときは、 油圧ポ ンプの吐出圧力が第 2所定圧力（Π)から低くなるに従って上記油圧ポンプの吐出 流量が増加するよう油圧ポンプの押しのけ容積を制御することを特徴とする作業 機の油圧駆動方法。  When the discharge pressure of the hydraulic pump is equal to or lower than the first predetermined pressure (P1), the hydraulic pressure is increased so that the discharge flow rate of the hydraulic pump increases as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (Π). A hydraulic drive method for a working machine, comprising controlling a displacement of a pump.
1 5 . 請求項 1 2記載の作業機の油圧駆動方法において、 15. The hydraulic drive method for a working machine according to claim 12,
上記燃料噴射制御装置（12， 13)は、 ガバナ領域の少なくとも一部を逆ドループ 特性に制御可能なものであり、  The fuel injection control device (12, 13) can control at least a part of the governor region to have a reverse droop characteristic,
上記油圧ポンプの吐出圧力が上記第 1所定圧力（P 1)以下にあるときは、 上記油 圧ポンプ (2)の吐出圧力が上記第 2所定圧力 (P1)から低くなるに従つて上記油圧 ポンプの吐出流量が増加するよう上記油圧ポンプの押しのけ容積を増加させる制 御と、 上記油圧ポンプの吐出圧力が上記第 2所定圧力（Π)から低くなるに従って 上記油圧ポンプの吐出流量が一定に保たれるよう上記油圧ポンプの押しのけ容積 を増加させる制御のいずれか一方を選択可能であることを特徴とする作業機の油 圧駆動方法。  When the discharge pressure of the hydraulic pump is lower than or equal to the first predetermined pressure (P1), as the discharge pressure of the hydraulic pump (2) decreases from the second predetermined pressure (P1), the hydraulic pump Control to increase the displacement of the hydraulic pump so as to increase the discharge flow rate of the hydraulic pump, and maintain the discharge flow rate of the hydraulic pump constant as the discharge pressure of the hydraulic pump decreases from the second predetermined pressure (Π). A hydraulic pump driving method for a working machine, wherein one of the controls for increasing the displacement of the hydraulic pump can be selected.
PCT/JP2002/006138 2001-06-21 2002-06-20 Hydraulic driving unit for working machine, and method of hydraulic drive WO2003001067A1 (en)

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AU2002313244A AU2002313244B2 (en) 2001-06-21 2002-06-20 Hydraulic driving unit for working machine, and method of hydraulic drive
US10/344,120 US7048515B2 (en) 2001-06-21 2002-06-20 Hydraulic drive system and method using a fuel injection control unit
DE60238983T DE60238983D1 (en) 2001-06-21 2002-06-20 HYDRAULIC DRIVE UNIT OF A WORKING MACHINE, AND HYDRAULIC DRIVE PROCESS
JP2003507430A JP4077789B2 (en) 2001-06-21 2002-06-20 Hydraulic drive device and hydraulic drive method for work machine
KR1020037002428A KR100540772B1 (en) 2001-06-21 2002-06-20 Hydraulic driving unit for working machine, and method of hydraulic drive
EP02738772A EP1398512B1 (en) 2001-06-21 2002-06-20 Hydraulic driving unit for working machine, and method of hydraulic drive

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Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US7588118B2 (en) 2003-05-07 2009-09-15 Komatsu Ltd. Work machine with engine control device
JP2017061795A (en) * 2015-09-24 2017-03-30 日立建機株式会社 Hydraulic drive unit of construction machine
KR20210105423A (en) * 2018-12-26 2021-08-26 웨이차이 파워 컴퍼니 리미티드 Hydraulic Variable Pump Set and Excavator

Families Citing this family (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US7797092B2 (en) * 2006-11-06 2010-09-14 Caterpillar Inc Method and system for controlling machine power
KR100919436B1 (en) * 2008-06-03 2009-09-29 볼보 컨스트럭션 이키프먼트 홀딩 스웨덴 에이비 Torque control system of plural variable displacement hydraulic pump and method thereof
JP5188444B2 (en) * 2009-04-23 2013-04-24 カヤバ工業株式会社 Hydraulic drive device for work equipment
US8365544B2 (en) * 2009-08-20 2013-02-05 Trane International Inc. Screw compressor drive control
JP2015161181A (en) * 2014-02-26 2015-09-07 コベルコ建機株式会社 Engine control device for construction machine
WO2015140622A1 (en) * 2014-03-20 2015-09-24 Danfoss Power Solutions Inc. Electronic torque and pressure control for load sensing pumps
JP6966830B2 (en) * 2018-04-27 2021-11-17 キャタピラー エス エー アール エル Calibration system for variable displacement hydraulic pumps
WO2020151821A1 (en) 2019-01-24 2020-07-30 Trumpf Lasersystems For Semiconductor Manufacturing Gmbh Arrangement for monitoring an optical element, laser source and euv radiation generation apparatus

Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4077744A (en) * 1974-11-18 1978-03-07 Massey-Ferguson Services N.V. Control systems for variable capacity hydraulic machines
US5267441A (en) 1992-01-13 1993-12-07 Caterpillar Inc. Method and apparatus for limiting the power output of a hydraulic system
JPH0783084A (en) 1993-09-16 1995-03-28 Hitachi Constr Mach Co Ltd Hydraulic construction machine
JPH1150868A (en) * 1997-08-04 1999-02-23 Yanmar Agricult Equip Co Ltd Engine speed adjustment structure of traveling vehicle

Family Cites Families (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3818883A (en) * 1969-07-28 1974-06-25 Caterpillar Tractor Co Isochronous governor
US4600364A (en) * 1983-06-20 1986-07-15 Kabushiki Kaisha Komatsu Seisakusho Fluid operated pump displacement control system
JPH1089111A (en) * 1996-09-17 1998-04-07 Yanmar Diesel Engine Co Ltd Control mechanism of engine loaded with working machine
JPH10159599A (en) * 1996-11-29 1998-06-16 Yanmar Diesel Engine Co Ltd Electronic governor control mechanism

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4077744A (en) * 1974-11-18 1978-03-07 Massey-Ferguson Services N.V. Control systems for variable capacity hydraulic machines
US5267441A (en) 1992-01-13 1993-12-07 Caterpillar Inc. Method and apparatus for limiting the power output of a hydraulic system
JPH0783084A (en) 1993-09-16 1995-03-28 Hitachi Constr Mach Co Ltd Hydraulic construction machine
JPH1150868A (en) * 1997-08-04 1999-02-23 Yanmar Agricult Equip Co Ltd Engine speed adjustment structure of traveling vehicle

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
See also references of EP1398512A4 *

Cited By (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US7588118B2 (en) 2003-05-07 2009-09-15 Komatsu Ltd. Work machine with engine control device
DE112004000751B4 (en) * 2003-05-07 2012-11-15 Komatsu Ltd. Work machine with engine control device
JP2017061795A (en) * 2015-09-24 2017-03-30 日立建機株式会社 Hydraulic drive unit of construction machine
KR20210105423A (en) * 2018-12-26 2021-08-26 웨이차이 파워 컴퍼니 리미티드 Hydraulic Variable Pump Set and Excavator
JP2022516145A (en) * 2018-12-26 2022-02-24 ▲い▼柴動力股▲ふん▼有限公司 Variable capacity hydraulic pump set and excavator
JP7200385B2 (en) 2018-12-26 2023-01-06 ▲い▼柴動力股▲ふん▼有限公司 Variable displacement hydraulic pump set and excavator
KR102577950B1 (en) * 2018-12-26 2023-09-14 웨이차이 파워 컴퍼니 리미티드 Hydraulic variable pump sets and excavators

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KR100540772B1 (en) 2006-01-10

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