JP6698359B2 - Hydraulic system with fail-safe - Google Patents

Description

  The present invention is a hydraulic system that electrically controls a discharge flow rate of a pump using an electromagnetic proportional valve, and when a failure occurs in which a command current is not supplied to the electromagnetic proportional valve due to a disconnection of an electric system or a failure of a control device, The present invention relates to a fail-safe hydraulic system that can function safely during a failure in which a solenoid proportional valve is stuck in a state where no current is flowing.

  BACKGROUND ART Conventionally, a hydraulic system has been known in which the discharge flow rate of a pump is electrically controlled using an electromagnetic proportional valve. In such a hydraulic system, measures are taken when the solenoid proportional valve does not operate during a failure. For example, Patent Document 1 discloses a fail-safe hydraulic system 100 as shown in FIG.

  Specifically, the hydraulic system 100 includes a variable displacement pump 110, a flow rate adjusting device 120 that adjusts a tilt angle of the pump 110, and a solenoid proportional valve 130. The flow rate adjusting device 120 includes a servo piston 121 that changes the tilt angle of the pump 110, a spool 122 that drives the servo piston 121, a horsepower control piston 124 and a flow rate control piston 123 that operate the spool 122, and a switching valve 128. Including. Further, the flow rate adjusting device 120 is formed with two pressure receiving chambers 126 and 127 for the horsepower control piston 124 and a pressure receiving chamber 125 for the flow rate control piston 123.

  The discharge pressure Pd of the pump 110 is introduced into one pressure receiving chamber 126 for the horsepower control piston 124, and the horsepower control piston 124 moves the spool 122 in the flow rate reducing direction when the discharge pressure Pd rises. The other pressure receiving chamber 127 normally communicates with the tank.

  Normally, the secondary pressure of the solenoid proportional valve 130 is introduced into the pressure receiving chamber 125 for the flow control piston 123. The flow rate control piston 123 moves the spool 122 in the flow rate increasing direction when the secondary pressure of the solenoid proportional valve 130 increases. It should be noted that the horsepower control piston 124 and the flow rate control piston 123 are configured to move the spool 122 preferentially when the discharge flow rate is limited to a small value.

  The switching valve 128 is normally maintained in the right position in FIG. 3 by outputting a relatively large secondary pressure from the electromagnetic proportional valve 130. Thereby, as described above, the pressure receiving chamber 127 communicates with the tank, and the secondary pressure of the solenoid proportional valve 130 is introduced into the pressure receiving chamber 125. On the other hand, at the time of failure, the secondary pressure of the solenoid proportional valve 130 becomes substantially zero, so the switching valve 128 is switched to the left position in FIG. As a result, the pilot pressure Pp of the pilot operated valve (operating device) is introduced into the pressure receiving chamber 125 for the flow rate control piston 123, and the primary pressure Psv of the solenoid proportional valve 130 is introduced into the pressure receiving chamber 127 for the horsepower control piston 124. To be done. When the primary pressure Psv is introduced into the pressure receiving chamber 127, the horsepower control line is below the engine horsepower equivalent line, and the upper limit of the discharge flow rate Q with respect to the discharge pressure Pd is kept low.

Japanese Patent No. 4041789

  However, in the hydraulic system 100 shown in FIG. 3, since the flow rate adjusting device 120 includes the switching valve 128, the flow rate adjusting device 120 becomes large and the cost is high.

  Therefore, an object of the present invention is to provide a fail-safe hydraulic system that can realize fail-safe without using a switching valve.

  In order to solve the above problems, a fail-safe hydraulic system of the present invention is a variable displacement pump driven by an engine, which supplies hydraulic oil to a hydraulic actuator via a control valve, and a tilt angle of the pump. And a second end having a diameter larger than that of the first end exposed in the first pressure receiving chamber into which the discharge pressure of the pump is introduced and the first end exposed in the second pressure receiving chamber. A servo piston having a portion, and an adjusting valve for adjusting a control pressure introduced into the second pressure receiving chamber, the spool moving in a flow rate decreasing direction for increasing the control pressure and a flow rate increasing direction for decreasing the control pressure. And a regulating valve including a command current, an electromagnetic proportional valve that shows a negative correlation between the command current and the secondary pressure, and the flow rate through the spool when at least one of the secondary pressure of the electromagnetic proportional valve and the discharge pressure of the pump rises. A horsepower control piston for moving in a decreasing direction, a flow control piston for moving the spool in the flow increasing direction when the secondary pressure of the solenoid proportional valve increases, and an operating lever for operating the control valve. A command current according to the tilt angle of the operating lever between a minimum command current and a maximum command current greater than zero so that the command current decreases as the tilt angle of the operating lever increases. And a control device for feeding to the solenoid proportional valve, the horsepower control piston, the minimum horsepower control line when the minimum command current is sent to the solenoid proportional valve traces the engine horsepower equivalent line. The emergency horsepower control line is configured to exceed the engine horsepower equivalent line and to be below the engine horsepower equivalent line when the secondary pressure of the solenoid proportional valve becomes maximum.

  According to the above configuration, normally, as the tilt angle of the operation lever increases, the secondary pressure of the solenoid proportional valve increases, and the flow control piston moves the spool in the flow increasing direction. As a result, the discharge flow rate of the pump can be increased as the tilt angle of the operation lever increases. At this time, since the minimum horsepower control line traces the engine horsepower equivalent line or exceeds the engine horsepower equivalent line, the discharge flow rate of the pump is hardly limited by the horsepower control piston.

  On the other hand, when the electric system is cut off or the control device fails, the command current is not supplied to the solenoid proportional valve, and the secondary pressure of the solenoid proportional valve becomes equal to the primary pressure of the solenoid proportional valve and becomes maximum. Further, when the solenoid proportional valve is stuck in a state where no current is flowing, the secondary pressure of the solenoid proportional valve becomes maximum even if the command current is supplied to the solenoid proportional valve. Therefore, during such a failure, the maximum secondary pressure from the solenoid proportional valve acts on the flow control piston, and the flow control piston tries to maximize the discharge flow rate of the pump. However, at this time, since the emergency horsepower control line is below the engine horsepower equivalent line, the discharge flow rate is controlled by the horsepower control piston according to the discharge pressure of the pump. Therefore, fail safe can be realized without using a switching valve.

  The emergency horsepower control line may trace a horsepower suppression line corresponding to a horsepower of 30 to 90% of the horsepower of the engine. According to this configuration, it is possible to prevent the engine from stalling even when the machine equipped with the hydraulic system is used at a high altitude and the engine output decreases.

  According to the present invention, there is provided a fail-safe hydraulic system capable of realizing fail-safe without using a switching valve.

It is a schematic structure figure of a fail safe hydraulic system concerning one embodiment of the present invention. (A) is a graph showing the relationship between the tilt angle of the operating lever and the command current I to the solenoid proportional valve, (b) is a graph showing the relationship between the command current I and the secondary pressure Pf of the solenoid proportional valve, (c) 8A is a graph showing the control characteristic of the horsepower control piston, and FIG. 13D is a graph showing the control characteristic of the flow rate control piston. It is a schematic block diagram of the conventional hydraulic system with a fail safe.

  FIG. 1 shows a fail-safe hydraulic system 1 according to an embodiment of the present invention. The hydraulic system 1 is mounted on, for example, a construction machine such as a hydraulic excavator or a hydraulic crane or an industrial machine.

  Specifically, the hydraulic system 1 includes a variable displacement main pump 12 and a flow rate adjusting device 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 14.

  A circulation line 13 extends from the main pump 12 to the tank, and a control valve 41 is arranged on the circulation line 13. The control valve 41 controls supply and discharge of hydraulic oil to and from the hydraulic actuator 42. That is, the main pump 12 supplies hydraulic oil to the hydraulic actuator 42 via the control valve 41. The hydraulic actuator 42 may be a hydraulic cylinder or a hydraulic motor. Further, a plurality of hydraulic actuators 42 and control valves 41 may be provided.

  The control valve 41 is operated by the operation device 5 including an operation lever. In the present embodiment, the operating device 5 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 41 by a pair of pilot lines 51 and 52. However, the operation device 5 may be an electric joystick that outputs an electric signal according to the tilt angle of the operation lever, and a pair of solenoid proportional valves may be connected to the pilot port of the control valve 41. The pilot lines 51 and 52 are provided with pressure gauges 71 and 72 for detecting the pilot pressure output from the operating device 5, respectively.

  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 the angle of the oblique shaft.

  The flow rate adjusting device 2 includes a servo piston 31 that changes the tilt angle of the main pump 12, and an adjustment valve 32 that drives the servo piston 31. The flow rate adjusting device 2 is formed with a first pressure receiving chamber 21 into which the discharge pressure Pd of the main pump 12 is introduced and a second pressure receiving chamber 22 into which the control pressure Pc is introduced. The servo piston 31 has a first end and a second end having a diameter larger than that of the first end. The first end portion is exposed to the first pressure receiving chamber 21, and the second end portion is exposed to the second pressure receiving chamber 22.

  The adjusting valve 32 is for adjusting the control pressure Pc introduced into the second pressure receiving chamber 22. Specifically, the adjustment valve 32 includes a spool 34 and a spool 34 that move in a flow rate decreasing direction that increases the control pressure Pc (rightward in FIG. 1) and a flow rate increasing direction that decreases the control pressure Pc (leftward in FIG. 1). It includes a sleeve 33 for receiving.

  The servo piston 31 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 31. The sleeve 33 is connected to the servo piston 31 by the first lever 3a so as to be movable in the axial direction of the servo piston 31. A pump port, a tank port and an output port (the output port communicates with the second pressure receiving chamber 22) are formed in the sleeve 33, and the output port is determined by the relative position of the spool 34 with respect to the sleeve 33. Shut off or communicate with either pump port or tank port. Then, when the spool 34 is moved in a flow rate decreasing direction or a flow rate increasing direction by a horsepower control piston 35 and a flow rate control piston 36, which will be described later, the forces (pressure×servo piston pressure receiving area) acting from both sides of the servo piston 31 are balanced. Then, the relative position between the spool 34 and the sleeve 33 is determined, and the control pressure Pc is adjusted.

  The flow rate adjusting device 2 also includes a horsepower control piston 35 and a flow rate control piston 36 that operate the spool 34. The horsepower control piston 35 and the flow rate control piston 36 are connected to the spool 34 via the second lever 3b and the third lever 3c, respectively, so as to press the spool 34. However, the movement of the horsepower control piston 35 in the direction away from the spool 34 and the movement of the flow rate control piston 36 in the direction away from the spool 34 are not restricted. Further, the flow rate adjusting device 2 is formed with two pressure receiving chambers 23 and 24 for the horsepower control piston 35 and a pressure receiving chamber 25 for the flow rate control piston 36.

  The discharge pressure Pd of the main pump 12 is introduced into one pressure receiving chamber 23 for the horsepower control piston 35. The other pressure receiving chamber 24 for the horsepower control piston 35 and the pressure receiving chamber 25 for the flow rate control piston 36 are connected to a solenoid proportional valve 62 by a secondary pressure line 61. That is, the secondary pressure Pf of the solenoid proportional valve 62 is introduced into the pressure receiving chambers 24 and 25. The solenoid proportional valve 62 is connected to the sub pump 14 by the primary pressure line 15.

  The horsepower control piston 35 moves the spool 34 in the flow rate reducing direction when at least one of the secondary pressure Pf of the solenoid proportional valve 62 and the discharge pressure Pd of the main pump 12 increases. On the other hand, when the secondary pressure Pf of the solenoid proportional valve 62 or the discharge pressure Pd of the main pump 12 decreases, the horsepower control piston 35 is pushed back by the urging force of the second spring 38, and as a result, the spool 34 moves to the first position. The urging force of the spring 37 moves in the direction of increasing the flow rate.

  The flow rate control piston 36 moves the spool 34 in the flow rate increasing direction when the secondary pressure Pf of the solenoid proportional valve 62 rises. On the other hand, when the secondary pressure Pf of the solenoid proportional valve 62 decreases, the flow rate control piston 36 moves the spool 34 in the flow rate reducing direction by the urging force of the third spring 39. It should be noted that the horsepower control piston 35 and the flow rate control piston 36 are configured such that the one that limits the discharge flow rate Q to a small value (that is, the one that commands a small flow rate) preferentially moves the spool 34.

  As shown in FIG. 2B, the solenoid proportional valve 62 is a negative type in which the command current I and the secondary pressure Pf have a negative correlation. In other words, the secondary pressure Pf of the solenoid proportional valve 62 decreases as the command current I increases.

  The solenoid proportional valve 62 is controlled by the controller 7. As shown in FIG. 2A, the control device 7 increases as the tilt angle of the operation lever of the operation device 5 increases (in this embodiment, as the pilot pressure detected by the pressure gauges 71 and 72 increases). The command current I is sent to the solenoid proportional valve 62 so that the command current I decreases. That is, the command current I is between the maximum command current I2 (when the tilt angle of the operation lever is zero) and the minimum command current I1 (when the tilt angle of the operation lever is maximum) according to the tilt angle of the operation lever. Shift with. However, the minimum command current I1 is larger than zero.

  In the present embodiment, as shown in FIG. 2B, when the command current I is the maximum command current I2, the secondary pressure Pf of the solenoid proportional valve 62 becomes zero, and when the command current I is the minimum command current I1. The secondary pressure Pf of the solenoid proportional valve 62 becomes α. However, when the command current I is the maximum command current I2, the secondary pressure Pf of the solenoid proportional valve 62 may not be zero but may be a predetermined value close to zero. When the command current I is zero, which is smaller than the minimum command current I1, in other words, when the command current I is not supplied to the solenoid proportional valve 62, the secondary pressure Pf of the solenoid proportional valve 62 becomes β equal to the primary pressure. Will be the largest. In order to secure a large difference between α and β, the minimum command current I1 is set slightly higher than near zero.

  In the normal state other than the fail state, the command current I decreases and the secondary pressure Pf of the solenoid proportional valve 62 increases as the tilt angle of the operating lever increases, and the flow rate control piston 36 moves the spool 34 in the flow rate increasing direction. Let As a result, as shown in FIG. 2D, the discharge flow rate Q of the main pump 12 can be increased as the tilt angle of the operation lever increases.

  As shown in FIG. 2C, the horsepower control piston 35 defines a horsepower control line that defines the upper limit of the discharge flow rate Q with respect to the discharge pressure Pd of the main pump 12. Normally, the horsepower control line shifts between the minimum horsepower control line L1 and the maximum horsepower control line L2. The minimum horsepower control line L1 is the horsepower control line when the minimum command current I1 is sent to the solenoid proportional valve 62 (that is, when Pf=α), and the maximum horsepower control line L2 is the maximum command current I2. It is a horsepower control line when the pressure is fed to the solenoid proportional valve 62 (that is, when Pf=0).

  In the present embodiment, the minimum horsepower control line L1 exceeds the engine horsepower equivalent line LH. However, the minimum horsepower control line L1 may trace the engine horsepower equivalent line LH, in other words, be in contact with the engine horsepower equivalent line LH or intersect at at least one point. In this way, if the minimum horsepower control line L1 exceeds the engine horsepower equivalent line LH or traces the engine horsepower equivalent line LH, the discharge flow rate Q of the main pump 12 is almost always restricted by the horsepower control piston 35 in normal times. Absent. That is, normally, the discharge flow rate Q of the main pump 12 is mainly controlled by the flow rate control piston 36 and changes as shown by the solid line in FIG.

  On the other hand, when the electric system is cut off or the control device 7 fails, the command current I is not sent to the solenoid proportional valve 62, and the secondary pressure Pf of the solenoid proportional valve 62 becomes equal to the primary pressure and becomes maximum. . Further, when the electromagnetic proportional valve 62 is stuck in a state where no current is flowing, even if the command current is supplied to the electromagnetic proportional valve 62, the secondary pressure Pf of the electromagnetic proportional valve 62 becomes maximum. Therefore, at the time of such a failure, the maximum secondary pressure Pf from the solenoid proportional valve 62 acts on the flow control piston 36, and the flow control piston 36 causes the main pump to flow as shown by the broken line in FIG. The discharge flow rate Q of 12 is to be maximized.

  However, at the time of the above failure, the maximum secondary pressure Pf from the solenoid proportional valve 62 is also introduced into the pressure receiving chamber 24 for the horsepower control piston 35. Therefore, as shown in FIG. 2C, the emergency horsepower control line L0 defined by the horsepower control piston 35 falls below the engine horsepower equivalent line LH. Therefore, the discharge flow rate Q of the main pump 12 is controlled by the horsepower control piston 35 according to the discharge pressure Pd. It is desirable that the emergency horsepower control line L0 traces a horsepower suppression line LS corresponding to a horsepower of 30 to 90% of the horsepower of the engine (in other words, contact with the horsepower suppression line LS or intersect at at least one point). This is because even when the machine equipped with the hydraulic system 1 is used at a high altitude and the engine output is reduced, the stall of the engine 11 can be prevented.

  As described above, in the hydraulic system 1 of the present embodiment, normally, the discharge flow rate Q of the main pump 12 is controlled by the flow rate control piston 36, while electromagnetic proportionality is generated due to the disconnection of the electric system or the failure of the control device 7. During a failure in which the command current is not supplied to the valve 62 or a failure in which the solenoid proportional valve 62 is stuck in a state where no current flows, the horsepower control piston 35 controls the discharge flow rate Q of the main pump 12. Therefore, fail safe can be realized without using a switching valve.

  That is, the hydraulic system 1 of the present embodiment has a smaller number of parts than the conventional hydraulic system, so that the cost can be reduced and the reliability can be improved. Moreover, it is possible to reduce the size and weight of the flow rate adjusting device 2.

(Modification)
The present invention is not limited to the above-described embodiments, and various modifications can be made without departing from the spirit of the present invention.

1 Hydraulic System 11 Engine 12 Main Pump 31 Servo Piston 32 Regulator Valve 34 Spool 35 Horse Power Control Piston 36 Flow Control Piston 41 Control Valve 42 Hydraulic Actuator 5 Operating Device 62 Electromagnetic Proportional Valve 7 Control Device L1 Minimum Horsepower Control Line L0 Emergency Horsepower Control Line LH Engine horsepower equivalent line LS Horsepower suppression line

Claims (2)

A variable displacement pump driven by an engine, which supplies hydraulic oil to a hydraulic actuator via a control valve,
A servo piston for changing a tilt angle of the pump, the first piston being exposed in a first pressure receiving chamber into which a discharge pressure of the pump is introduced and the first end being exposed in a second pressure receiving chamber. have a second end portion of the large diameter, the control pressure introduced to the second pressure receiving chamber to reduce the tilt angle of the pump when elevated, tilting of the pump when the control pressure drops Servo piston that increases the angle ,
A regulating valve for adjusting the control pressure, the control valve comprising a spool which moves in the flow rate increasing direction to reduce the flow reduction direction and the control pressure increases the control pressure,
A solenoid proportional valve that shows a negative correlation between the command current and the secondary pressure,
A horsepower control piston that moves the spool in the flow rate reducing direction when at least one of the secondary pressure of the solenoid proportional valve and the discharge pressure of the pump rises;
A flow rate control piston that moves the spool in the flow rate increasing direction when the secondary pressure of the solenoid proportional valve increases;
An operating device including an operating lever for operating the control valve;
A command current corresponding to the tilt angle of the operating lever is set between the minimum command current and the maximum command current greater than zero so that the command current decreases as the tilt angle of the operating lever increases. And a control device for feeding to
The horsepower control piston and the flow control piston is configured to move the spool in preference is better to limit small discharge flow rate of the pump among them,
In the horsepower control piston, the minimum horsepower control line when the minimum command current is sent to the solenoid proportional valve traces the engine horsepower equivalent line or exceeds the engine horsepower equivalent line, and the two of the solenoid proportional valve are provided. A fail-safe hydraulic system, wherein an emergency horsepower control line when the next pressure becomes maximum is configured to fall below the engine horsepower equivalent line.
  The fail-safe hydraulic system according to claim 1, wherein the emergency horsepower control line traces a horsepower suppression line corresponding to a horsepower of 30 to 90% of the horsepower of the engine.

Images