JP2018084196A - Hydraulic drive system - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a hydraulic drive system capable of both the capacity controlling and horsepower controlling while simplifying the structure of a regulator.SOLUTION: A hydraulic drive system comprises: a pump for discharging hydraulic fluid of the flow rate in proportion to the tilt angle; at least one control valve disposed on a center by-pass line extending from the pump to a tank; a regulator that is for regulating the tilt angle of the pump and includes a servo piston, a spool, and a common piston; an electromagnetic proportional valve for outputting a secondary pressure in proportion to a command current; a high-pressure selection valve by selecting the higher pressure in a negative control pressure and the secondary pressure outputted from the electromagnetic proportional valve as a signal pressure to supply to the common piston; a pressure sensor for detecting the discharge pressure of the pump; and a control device that stores a horsepower control characteristic line in advance and that is for supplying the command current to the electromagnetic proportional valve so that the pump discharges the hydraulic fluid of the discharge flow rate determined by the horsepower control characteristic line in accordance with the pump discharge pressure detected by the pressure sensor.SELECTED DRAWING: Figure 1

Description

  The present invention relates to a hydraulic drive system mounted on, for example, a construction machine.

  In a hydraulic drive system used for industrial machines and construction machines, hydraulic oil is supplied from a pump to a plurality of hydraulic actuators. Specifically, a center bypass line extends from the pump to the tank, and a plurality of control valves are disposed on the center bypass line. Each control valve controls the supply and discharge of hydraulic fluid to the corresponding actuator.

  In general, the above pump is a variable displacement pump that discharges hydraulic oil at a flow rate corresponding to the tilt angle, and the tilt angle of the pump is adjusted by a regulator. Some regulators perform flow rate control that changes the pump discharge flow rate according to the operating speed of the actuator, and horsepower control that limits the pump discharge flow rate so that the pump load does not exceed the output of the engine that drives the pump. is there.

  For example, Patent Document 1 describes a regulator that controls the discharge flow rate of a pump by a negative control method, and includes a flow rate control piston and a horsepower control piston. The flow rate control piston and the horsepower control piston are for driving a spool for adjusting the control pressure for the servo piston, and are configured such that the smaller the discharge flow rate of the pump, the higher the function.

  More specifically, the flow control piston moves the spool in the flow reduction direction as the negative control pressure, which is the pressure between the control valve located on the most downstream side in the center bypass line and the throttle, increases. The horsepower control piston moves the spool in the flow rate reduction direction as the discharge pressure of the pump increases.

Japanese Patent No. 5965502

  However, in the regulator as described above, two pistons, that is, a flow rate control piston and a horsepower control piston are required as the spool driving means, and the structure is complicated.

  Therefore, an object of the present invention is to provide a hydraulic drive system capable of performing both flow rate control and horsepower control while simplifying the structure of the regulator.

  In order to solve the above-described problems, the hydraulic drive system of the present invention is disposed on a variable displacement pump that discharges hydraulic oil at a flow rate corresponding to a tilt angle, and a center bypass line that extends from the pump to the tank. An operation device including an operation lever for operating the at least one control valve, and a regulator for adjusting a tilt angle of the pump. A servo piston having a first end exposed in the first pressure receiving chamber into which the discharge pressure of the pump is introduced and a second end larger in diameter than the first end exposed in the second pressure receiving chamber; The spool moving in the flow rate decreasing direction for increasing the control pressure introduced into the second pressure receiving chamber and the flow rate increasing direction for decreasing the control pressure, and the higher the signal pressure, the earlier A proportional solenoid that outputs a secondary pressure corresponding to a command current, and a regulator that includes a common piston that moves the spool in the flow rate reduction direction, and a secondary pressure that increases as the command current increases. The signal pressure is selected by selecting a higher one of a proportional valve, a negative control pressure that is the pressure on the most downstream side of the at least one control valve in the center bypass line, and a secondary pressure output from the electromagnetic proportional valve. A pressure sensor for detecting the discharge pressure of the pump, a horsepower control characteristic line that is a relation line between the discharge pressure and the discharge flow rate of the pump, and the pressure sensor The command current is set so that the pump discharges hydraulic oil at a discharge flow rate determined by the horsepower control characteristic line according to the pump discharge pressure detected at And a control device for feeding to the electromagnetic proportional valve, characterized in that.

  According to the above configuration, when the negative control pressure is higher than the secondary pressure output by the electromagnetic proportional valve, the negative control pressure becomes the signal pressure of the common piston so that flow control can be performed, and the electromagnetic proportional valve outputs When the secondary pressure to be performed is higher than the negative control pressure, the secondary pressure output from the electromagnetic proportional valve becomes the signal pressure of the common piston, so that the horsepower control can be performed. In addition, since the regulator includes one common piston as the spool driving means, the structure of the regulator can be simplified. Furthermore, if a plurality of horsepower control characteristic lines are stored in the control device and one of them is selected according to the situation, the horsepower control can be made variable.

  The regulator is connected to the high pressure selection valve by a signal pressure line, the first working chamber for applying the signal pressure to the common piston, and the common piston connected to the center bypass line by a fail safe line. The hydraulic drive system includes a second working chamber for applying a discharge pressure of the pump, and is a switching valve provided in the fail-safe line, wherein a secondary pressure output from the electromagnetic proportional valve is a set pressure. The fail-safe line is opened when the pressure is lower than that, and the discharge pressure of the pump is guided to the second working chamber, and the fail-safe when the secondary pressure output from the electromagnetic proportional valve is higher than the set pressure. A switching valve for blocking the line and communicating the second working chamber with the tank may be further provided. According to this configuration, when the electromagnetic proportional valve operates normally, the fail-safe line is blocked and the second working chamber communicates with the tank, so that both flow control and horsepower control can be performed as described above. it can. On the other hand, when a malfunction occurs in which the secondary pressure of the solenoid proportional valve is maintained lower than the set pressure due to electrical system breakage or electromagnetic proportional valve failure, the pump discharge pressure is superimposed on the negative control pressure as the common piston signal pressure. Can be made. As a result, when a malfunction occurs, the discharge flow rate of the pump is limited when the relationship between the discharge pressure and the discharge flow rate of the pump exceeds a predetermined horsepower control characteristic line for each operation amount of the operating device while performing flow rate control. be able to. Therefore, it is possible to reliably prevent the engine from being stopped due to overload.

  Alternatively, the regulator includes an operation chamber that is connected to the high pressure selection valve by a signal pressure line and that causes the signal pressure to act on the common piston, and the signal pressure line includes the operation chamber from the high pressure selection valve to the operation chamber. A check valve is provided that allows the flow to the opposite direction but prohibits the reverse flow. The hydraulic drive system is branched from the center bypass line to the tank on the most upstream side of the at least one control valve. A fail-safe line provided with a pair of throttles, and a switching valve provided in the fail-safe line upstream of the pair of throttles, wherein the secondary pressure output from the electromagnetic proportional valve is a set pressure The fail-safe line is opened when the pressure is lower than the value, and the fail-safe line is output when the secondary pressure output from the electromagnetic proportional valve is higher than the set pressure. A switching valve for blocking the flow, a portion between the working chamber and the check valve in the signal pressure line, and a portion between the pair of throttles in the failsafe line, from the failsafe line And a relay line provided with a check valve that allows a flow toward the signal pressure line but prohibits a reverse flow. According to this configuration, since the fail safe line is blocked when the electromagnetic proportional valve operates normally, both the flow rate control and the horsepower control can be performed as described above. On the other hand, when a malfunction occurs in which the secondary pressure of the electromagnetic proportional valve is maintained lower than the set pressure due to an electrical system breakage or a failure of the electromagnetic proportional valve, the signal pressure of the common piston is reduced by a pair of negative control pressure and fail-safe line. The pressure between the throttles is higher. Thus, when a failure occurs, the discharge flow rate of the pump can be limited when the relationship between the discharge pressure and the discharge flow rate of the pump exceeds a predetermined horsepower control characteristic line while performing flow rate control. Therefore, it is possible to reliably prevent the engine from being stopped due to overload.

  According to the present invention, it is possible to perform both flow rate control and horsepower control while simplifying the structure of the regulator.

1 is a schematic configuration diagram of a hydraulic drive system according to a first embodiment of the present invention. It is a graph which shows the relationship between the negative control pressure and pump discharge flow volume when the signal pressure to a common piston is a negative control pressure. (A) is a graph showing a horsepower control characteristic line, and (b) shows the relationship between the command current to the electromagnetic proportional valve and the pump discharge flow rate when the signal pressure to the common piston is the secondary pressure of the electromagnetic proportional valve. It is a graph. It is a schematic block diagram of the hydraulic drive system which concerns on 2nd Embodiment of this invention. It is a schematic block diagram of the hydraulic drive system which concerns on 3rd Embodiment of this invention.

(First embodiment)
FIG. 1 shows a hydraulic drive system 1A according to a first embodiment of the present invention. The hydraulic drive system 1A is mounted on, for example, a construction machine or an industrial machine such as a hydraulic excavator or a hydraulic crane. The construction machine may be self-propelled, may be mounted on a ship such as a dredger, or may be installed in a port as a fixed loader or unloader.

  Specifically, the hydraulic drive system 1 </ b> A includes a variable displacement main pump 12 that discharges hydraulic oil at a flow rate corresponding to the tilt angle, and a regulator 2 that adjusts the tilt angle of the main pump 12. The main pump 12 is driven by the engine 11. The engine 11 also drives the sub pump 13. However, the main pump 12 and the sub pump 13 may be driven by an electric motor.

  A center bypass line 14 extends from the main pump 12 to the tank, and at least one control valve 31 is disposed on the center bypass line 14. The center bypass line 14 is provided with a throttle 15 on the most downstream side of at least one control valve 31. Further, a throttle bypass line is connected to the center bypass line 14 so as to bypass the throttle 15, and a relief valve 16 is provided in this throttle bypass line.

  The control valve 31 controls supply and discharge of hydraulic oil to the hydraulic actuator 32. The hydraulic actuator 32 may be a hydraulic cylinder or a hydraulic motor. The number of sets of the hydraulic actuator 32 and the control valve 31 may be one or plural.

  For example, when the hydraulic drive system 1A is mounted on a self-propelled hydraulic excavator, the boom cylinder, arm cylinder, bucket cylinder, swing motor, and travel motor are the hydraulic actuators 32. In this case, two main pumps 12 may be provided.

  The control valve 31 is operated by an operation device 33 including an operation lever. In the present embodiment, the operating device 33 is a pilot operating valve that outputs a pilot pressure according to the tilt angle of the operating lever, and is connected to the pilot port of the control valve 31 by a pair of pilot lines. However, the operation device 33 may be an electric joystick that outputs an electric signal corresponding to the tilt angle of the operation lever, and a pair of electromagnetic proportional valves may be connected to the pilot port of the control valve 31.

  In the present embodiment, the main pump 12 is a swash plate pump whose tilt angle is defined by the angle of the swash plate 12a. However, the main pump 12 may be an oblique shaft pump whose tilt angle is defined by an angle formed by the drive shaft and the cylinder block.

  The regulator 2 includes a servo piston 21 that changes the tilt angle of the main pump 12 and an adjustment valve 23 that drives the servo piston 21. In the regulator 2, a first pressure receiving chamber 2a into which the discharge pressure Pd of the main pump 12 is introduced and a second pressure receiving chamber 2b into which the control pressure Pc is introduced are formed. The servo piston 21 has a first end and a second end having a larger diameter than the first end. The first end is exposed to the first pressure receiving chamber 2a, and the second end is exposed to the second pressure receiving chamber 2b.

  The adjustment valve 23 is for adjusting the control pressure Pc introduced into the second pressure receiving chamber 2b. Specifically, the regulating valve 23 includes a spool 24 that moves in a flow rate decreasing direction (in the left direction in FIG. 1) for increasing the control pressure Pc and a flow rate increasing direction (in the right direction in FIG. 1) that decreases the control pressure Pc, and a spool 24. A receiving sleeve 25 is included.

  The servo piston 21 is connected to the swash plate 12a of the main pump 12 so as to be movable in the axial direction of the servo piston 21. The sleeve 25 is connected to the servo piston 21 by a feedback lever 22 so as to be movable in the axial direction of the servo piston 21. The sleeve 25 is formed with a pump port, a tank port, and an output port (the output port communicates with the second pressure receiving chamber 2 b), and the output port is determined by the relative position between the sleeve 25 and the spool 24. Shut off from the port or communicate with either the pump port or the tank port. Then, when the spool 24 is moved in a flow rate decreasing direction or a flow rate increasing direction by a common piston 26 described later, the spool 24 and the sleeve 25 so that the forces (pressure × servo piston pressure receiving area) acting from both sides of the servo piston 21 are balanced. And the control pressure Pc is adjusted.

  The regulator 2 includes a common piston 26 for driving the spool 24 and a spring 27 disposed on the opposite side of the common piston 26 with the spool 24 interposed therebetween. The common piston 26 is used for both flow rate control and horsepower control. The spool 24 is pressed by the common piston 26 and moves in the flow rate decreasing direction, and moves in the flow rate increasing direction by the urging force of the spring 27.

  Further, the regulator 2 is formed with a working chamber 2c for applying the signal pressure Pp to the common piston 26. That is, the common piston 26 moves the spool 24 in the flow rate reduction direction as the signal pressure Pp increases.

  The working chamber 2 c is connected to the high pressure selection valve 41 by a signal pressure line 42. In the present embodiment, the high pressure selection valve 41 is disposed outside the regulator 2, but the high pressure selection valve 41 may be incorporated in the regulator 2. In this case, an electromagnetic proportional valve 51 described later may also be incorporated in the regulator 2. Alternatively, only the electromagnetic proportional valve 51 may be incorporated into the regulator 2 without incorporating the high pressure selection valve 41 into the regulator 2.

  The high pressure selection valve 41 is connected to a portion between the control valve 31 located on the most downstream side of the at least one control valve 31 in the center bypass line 14 and the throttle 15 by the flow rate control line 43, and has a horsepower. The control line 44 is connected to the electromagnetic proportional valve 51. The electromagnetic proportional valve 51 is connected to the sub pump 13 described above by a primary pressure line 52. A relief line branches off from the primary pressure line 52 and is connected to a tank, and a relief valve 53 is provided in the relief line.

  The electromagnetic proportional valve 51 is controlled by the control device 6. That is, the command current I is supplied from the control device 6 to the electromagnetic proportional valve 51, and the electromagnetic proportional valve 51 outputs the secondary pressure Ps corresponding to the command current I. The electromagnetic proportional valve 51 is a direct proportional type (normally closed type) in which the secondary pressure Ps increases as the command current I increases.

  The high pressure selection valve 41 includes the secondary pressure Ps output from the electromagnetic proportional valve 51 and the pressure between the control valve 31 located on the most downstream side of the at least one control valve 31 in the center bypass line 14 and the throttle 15. The higher one of the negative control pressures Pn that is (the pressure on the most downstream side of at least one control valve 31) is selected. The pressure (Ps or Pn) selected by the high pressure selection valve 41 is supplied to the common piston 26 as the signal pressure Pp through the signal pressure line 42.

  When the signal pressure Pp is the negative control pressure Pn, the discharge flow rate Q of the main pump 12 decreases as the negative control pressure Pn increases as shown in FIG. On the other hand, when the signal pressure Pp is the secondary pressure of the electromagnetic proportional valve 51, since the electromagnetic proportional valve 51 is a direct proportional type as described above, the discharge flow rate Q of the main pump 12 is shown in FIG. Thus, it decreases as the command current I increases. That is, the discharge flow rate Q is minimized when the command current I is equal to or greater than the second command current I2, and the discharge flow rate Q is maximized when the command current I is equal to or less than the first command current I1.

  Further, the center bypass line 14 is provided with a pressure sensor 61 for detecting the discharge pressure Pd of the main pump 12. The pressure sensor 61 is electrically connected to the control device 6. However, in FIG. 1, only a part of the signal lines is drawn for simplification of the drawing.

  For example, the control device 6 includes a memory such as a ROM and a RAM and a CPU, and a program stored in the ROM is executed by the CPU. The control device 6 stores in advance a horsepower control characteristic line as shown in FIG. The horsepower control characteristic line is a relationship line between the discharge pressure Pd of the main pump 12 and the discharge flow rate Q. The horsepower control characteristic line is determined based on the output of the engine 11 or set to be lower than the output of the engine 11 in order to improve fuel consumption. As shown in FIGS. 3A and 3B, the control device 6 supplies hydraulic oil having a discharge flow rate Qx determined by a horsepower control characteristic line in accordance with the discharge pressure Pdx of the main pump 12 detected by the pressure sensor 61. The command current Ix is supplied to the electromagnetic proportional valve 51 so that the main pump 12 discharges.

  As described above, in the hydraulic drive system 1A of the present embodiment, when the negative control pressure Pn is higher than the secondary pressure Ps output from the electromagnetic proportional valve 51, the negative control pressure Pn becomes the signal pressure Pp of the common piston 26. Therefore, when the secondary pressure Ps output from the electromagnetic proportional valve 51 is higher than the negative control pressure Pn, the secondary pressure Ps output from the electromagnetic proportional valve 51 is equal to the signal pressure Pp of the common piston 26. Thus, horsepower control can be performed. In addition, since the regulator 2 includes one common piston 26 as the driving means of the spool 24, the structure of the regulator 2 can be simplified. Furthermore, if a plurality of horsepower control characteristic lines are stored in the control device 6 and one of them is selected according to the situation, the horsepower control can be made variable.

(Second Embodiment)
FIG. 4 shows a hydraulic drive system 1B according to the second embodiment of the present invention. In the present embodiment and the third embodiment to be described later, the same components as those in the first embodiment are denoted by the same reference numerals, and redundant description is omitted.

  In the present embodiment, the regulator 2 is formed with a first working chamber 2d and a second working chamber 2e as working chambers for the common piston 26. The first working chamber 2d is connected to the high-pressure selection valve 41 by a signal pressure line 42, similarly to the working chamber 2c of the first embodiment. That is, the first working chamber 2d is a working chamber in which the signal pressure Pp is applied to the common piston 26. On the other hand, the second working chamber 2 e is connected to the center bypass line 14 by a fail safe line 71. The second working chamber 2e is a working chamber in which the discharge pressure Pd of the main pump 12 acts on the common piston 26. That is, the fail safe line 71 branches from the center bypass line 14 on the most upstream side of at least one control valve 31 and is connected to the second working chamber 2e.

  The fail safe line 71 is provided with a switching valve 72. The switching valve 72 is connected to the tank by a tank line 74.

  The switching valve 72 has a pilot port, and this pilot port is connected to the horsepower control line 44 by a pilot line 73. When the secondary pressure Ps output from the electromagnetic proportional valve 51 is lower than the set pressure Pt, the switching valve 72 opens the fail safe line 71 to guide the discharge pressure Pd of the main pump 12 to the second working chamber 2e, When the secondary pressure Ps is higher than the set pressure Pt, the fail safe line 71 is blocked and the second working chamber 2e is communicated with the tank.

  The set pressure Pt is supplied to the electromagnetic proportional valve 51 by the first command current I1, which is the maximum of the command current (0 to I1 in FIG. 3B) that maximizes the discharge flow rate Q of the main pump 12. Sometimes it is set to a value lower than the secondary pressure Ps1 output from the electromagnetic proportional valve 51. For example, when the secondary pressure Ps1 is 1 MPa, the set pressure Pt is 0.3 to 1.0 MPa.

  In the hydraulic drive system 1B of the present embodiment, when the electromagnetic proportional valve 51 operates normally, the fail safe line 71 is blocked and the second working chamber 2e communicates with the tank. Therefore, the flow rate is the same as in the first embodiment. Both control and horsepower control can be performed. On the other hand, when a malfunction occurs in which the secondary pressure Ps of the electromagnetic proportional valve 51 is maintained lower than the set pressure Pt due to a disconnection of the electrical system or a failure of the electromagnetic proportional valve 51, the negative control pressure Pn is used as the signal pressure Pp of the common piston 26. The discharge pressure Pd of the main pump 12 can be superimposed on (Pp = Pn + Pd × Ad / An). Here, An is a pressure receiving area of the common piston 26 in the first working chamber 2d, and Ad is a pressure receiving area of the common piston 26 in the second working chamber 2e. Thus, when a failure occurs, the relationship between the discharge pressure Pd of the main pump 12 and the discharge flow rate Q is determined in advance for each operation amount of the operating device 33 (discharge pressure Pd and discharge flow rate) while performing flow rate control. The discharge flow rate Q of the main pump 12 can be limited when exceeding the relationship line with Q). Therefore, the stop of the engine 11 due to overload can be reliably prevented.

(Third embodiment)
FIG. 5 shows a hydraulic drive system 1C according to a third embodiment of the present invention. In the present embodiment, the regulator 2 has one working chamber 2c as the working chamber for the common piston 26 as in the first embodiment.

  Furthermore, in this embodiment, the fail safe line 81 branches from the center bypass line 14 on the most upstream side of the at least one control valve 31 and is connected to the tank. The fail safe line 81 is provided with a pair of stops 82 and 83. Further, the fail safe line 71 is provided with a switching valve 84 on the upstream side of the pair of throttles 82 and 83.

  The switching valve 84 has a pilot port, which is connected to the horsepower control line 44 by a pilot line 85. The switching valve 84 opens the fail-safe line 81 when the secondary pressure Ps output from the electromagnetic proportional valve 51 is lower than the set pressure Pt, and fails when the secondary pressure Ps is higher than the set pressure Pt. Block.

  The set pressure Pt is a secondary output from the electromagnetic proportional valve 51 when the first command current I1 that is the maximum of the command current that maximizes the discharge flow rate Q of the main pump 12 is supplied to the electromagnetic proportional valve 51. It is set to a value lower than the pressure Ps1. For example, when the secondary pressure Ps1 is 1 MPa, the set pressure Pt is 0.3 to 1.0 MPa.

  In the present embodiment, a check valve 45 is provided in the signal pressure line 42. The check valve 45 allows the flow from the high pressure selection valve 41 toward the working chamber 2c, but prohibits the reverse flow.

  A portion between the pair of throttles 82 and 83 in the fail safe line 81 is connected to a portion between the working chamber 2 c and the check valve 45 in the signal pressure line 42 by a relay line 86. A check valve 87 is provided in the relay line 76. The check valve 87 allows the flow from the fail safe line 81 toward the signal pressure line 42 but prohibits the reverse flow.

  In the hydraulic drive system 1C of the present embodiment, the failsafe line 81 is blocked when the electromagnetic proportional valve 51 is operating normally. Therefore, both flow control and horsepower control can be performed as in the first embodiment. it can. On the other hand, when a malfunction occurs in which the secondary pressure Ps of the electromagnetic proportional valve 51 is maintained lower than the set pressure Pt due to a disconnection of the electric system or a failure of the electromagnetic proportional valve 51, the signal pressure Pp of the common piston 26 becomes the negative control pressure Pn. The pressure Pm between the pair of throttles 82 and 83 in the fail safe line 81 is the higher one. Thus, when a malfunction occurs, the relationship between the discharge pressure Pd and the discharge flow rate Q of the main pump 12 exceeds a predetermined horsepower control characteristic line (relation line between the discharge pressure Pd and the discharge flow rate Q) while performing flow rate control. The discharge flow rate Q of the main pump 12 can be limited. Therefore, the stop of the engine 11 due to overload can be reliably prevented.

  That is, in the present embodiment, when the above-described malfunction occurs, the discharge flow rate with respect to the discharge pressure Pd of the main pump 12 regardless of the magnitude of the flow control, in other words, regardless of the tilt angle of the operation lever of the operating device 33. Since the upper limit of the horsepower control can be set only by the relationship of Q, the engine horsepower can be used more effectively.

(Other embodiments)
The present invention is not limited to the first to third embodiments described above, and various modifications can be made without departing from the gist of the present invention.

  For example, in the second embodiment, instead of the fail safe line 71 and the switching valve 72, a fail safe line 81 and a switching valve 84 provided with a pair of throttles 82 and 83 similar to the third embodiment are adopted, A portion of the safe line 81 between the pair of throttles 82 and 83 may be connected to the second working chamber 2e by a relay line. According to this configuration, when a malfunction occurs in which the secondary pressure Ps of the electromagnetic proportional valve 51 is maintained lower than the set pressure Pt, the pressure Pm between the pair of throttles 82 and 83 in the fail safe line 71 is changed to the second working chamber. 2e.

1A to 1C Hydraulic drive system 12 Main pump 14 Center bypass line 2 Regulator 21 Servo piston 24 Spool 26 Common piston 2a First pressure receiving chamber 2b Second pressure receiving chamber 2c Working chamber 2d First working chamber 2e Second working chamber 31 Control valve 41 High pressure selection valve 42 Signal pressure line 45 Check valve 51 Proportional solenoid valve 6 Control device 61 Pressure sensor 71 Fail safe line 72 Switching valve 81 Fail safe line 82, 83 Restriction 84 Switching valve 86 Relay line 87 Check valve

Claims (3)

A variable displacement pump that discharges hydraulic oil at a flow rate according to the tilt angle;
At least one control valve disposed on a center bypass line extending from the pump to the tank for controlling supply and discharge of hydraulic fluid to the actuator;
An operating device including an operating lever for operating the at least one control valve;
A regulator for adjusting a tilt angle of the pump, which is larger than the first end exposed in the first pressure receiving chamber and the first end exposed in the second pressure receiving chamber into which the discharge pressure of the pump is introduced. A servo piston having a second end of a diameter, a spool that moves in a flow rate decreasing direction that increases the control pressure introduced into the second pressure receiving chamber, and a flow rate increasing direction that decreases the control pressure, and the higher the signal pressure, A regulator including a common piston that moves the spool in the flow rate reduction direction;
An electromagnetic proportional valve that outputs a secondary pressure according to the command current, and a direct proportional electromagnetic proportional valve that increases the secondary pressure when the command current increases;
The common piston is selected as the signal pressure by selecting the higher one of the negative control pressure that is the pressure on the most downstream side of the at least one control valve in the center bypass line and the secondary pressure that is output from the electromagnetic proportional valve. A high pressure selection valve to supply to
A pressure sensor for detecting a discharge pressure of the pump;
A horsepower control characteristic line, which is a relation line between the discharge pressure and the discharge flow rate of the pump, is stored in advance, and the operation of the discharge flow rate determined by the horsepower control characteristic line according to the discharge pressure of the pump detected by the pressure sensor A control device for supplying a command current to the electromagnetic proportional valve so that the pump discharges oil;
A hydraulic drive system comprising: The regulator is connected to the high pressure selection valve by a signal pressure line, the first working chamber for applying the signal pressure to the common piston, and the common piston connected to the center bypass line by a fail safe line. Including a second working chamber for applying a discharge pressure of the pump;
A switching valve provided in the fail-safe line, wherein when the secondary pressure output from the electromagnetic proportional valve is lower than a set pressure, the fail-safe line is opened to the second working chamber. A switching valve for guiding a discharge pressure and blocking the fail-safe line when the secondary pressure output from the electromagnetic proportional valve is higher than the set pressure and communicating the second working chamber with a tank; Item 2. The hydraulic drive system according to Item 1. The regulator includes a working chamber connected to the high pressure selection valve by a signal pressure line and operating the signal pressure on the common piston,
The signal pressure line is provided with a check valve that allows a flow from the high pressure selection valve to the working chamber but prohibits a reverse flow thereof,
A fail safe line provided with a pair of throttles, branched from the center bypass line to the tank on the most upstream side of the at least one control valve;
A switching valve provided in the fail-safe line upstream of the pair of throttles, wherein the fail-safe line is opened when a secondary pressure output from the electromagnetic proportional valve is lower than a set pressure; A switching valve that blocks the fail-safe line when the secondary pressure output from the electromagnetic proportional valve is higher than the set pressure;
The flow from the fail safe line to the signal pressure line connecting the portion between the working chamber and the check valve in the signal pressure line and the portion between the pair of throttles in the fail safe line is The hydraulic drive system according to claim 1, further comprising a relay line provided with a check valve that permits but prohibits reverse flow.

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