JPH09296803A - Hydraulic driving device and proportioning pressure reducing valve for hydraulic driving device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a hydraulic driving device and a proportioning pressure reducing valve, by which delicate operability of a part of a swing motor of a hydraulic shovel is improved, the speed of an actuator is not decreased, and work efficiency is not deteriorated. SOLUTION: A hydraulic driving device using pressure compensating valves 45 is provided with two pressure receiving areas in the opening directions of the pressure compensating valves 4, 5 so that loading pressures 14, 15 may act on one of the pressure compensating valves, secondary pressure 32 of differential pressure between maximum loading pressure 16 and pump delivery pressure 3 may act on the other one, and pressures on directional control valve upstream sides 6, 7 may act against the acting force. The secondary pressure 32 to act on the pressure compensating valve 4 for an actuator 10 requiring delicate operability is reduced through a proportioning pressure reducing valve 41, and the operation of the proportioning pressure reducing valve 41 is switched by using a selector valve 46.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic drive system for supplying the discharge oil of one or a plurality of hydraulic pumps to a plurality of actuators used in a construction machine or the like, in which only the flow rate of a specific actuator is used. The present invention relates to a hydraulic drive device and a constant ratio pressure reducing valve that are suitable for reducing the pressure.

[0002]

2. Description of the Related Art Conventionally, a hydraulic drive system for a construction machine is disclosed in, for example, Japanese Patent Publication No. 4-48967 as shown in FIG. Explaining with the hydraulic circuit diagram shown in FIG. 5, a plurality of pressure compensating valves 404 and 405 are connected in parallel to the discharge oil passages 23 and 3 of the variable displacement hydraulic pump 2 driven by the prime mover 1 such as an engine. Directional control valves 8 and 9 are connected to the output oil passages 6 and 7 of the pressure compensation valve via check valves 26 and 27, respectively, and the output side of each directional control valve is connected to the actuator 41.
0, 411, respectively, and the return oil from each actuator is returned to the tank 12 again via the direction control valves 8, 9. The pressure compensating valves 404 and 405 are actuated in the opening direction by the pump discharge pressure, that is, the pressure of the pump discharge oil passage 3 and the respective load pressures 14 and 15 detected by the respective directional control valves 8 and 9, so that the inlet pressure of the directional control valve is increased. 6, 7 and the maximum load pressure 16 detected by the shuttle valve 13 act in the closing direction, and the maximum load pressure 16 acts on the flow rate adjusting valve 418 for driving the displacement volume changing means 417 of the variable displacement pump 2. Then, the differential pressure between the pump discharge pressure and the maximum load pressure is determined by the spring 419 of the flow rate adjusting valve 418.
It is designed to control the pressure set by.

In such a hydraulic system, the differential pressure before and after the respective directional control valves 8 and 9, that is, the differential pressure and pressure between the output oil passage 6 of the pressure compensating valve 404 and the load pressure 14 on the downstream side of the directional control valve 8. The differential pressure between the output oil passage 7 of the compensating valve 405 and the load pressure 15 on the downstream side of the directional control valve 9 is determined by the differential pressure between the pump discharge pressure and the maximum load pressure 16.

Each of these actuators 41
When the flow rate to 0, 411 is relatively small and the total of the flow rates does not reach the maximum discharge flow rate of the variable displacement pump 2, the differential pressures before and after the respective directional control valves 8, 9 are the pump discharge pressure and the maximum load. Pressure difference with pressure 16, in other words, flow control valve 4
Eighteen springs 419 equal to the preset differential pressure. Therefore, even if there is a difference in the load pressure of each actuator, the flow rate to each actuator does not depend on the load pressure, and the flow rate to each actuator differs between the throttle opening degree of each directional control valve 8 and 9 and the spring 419. Determined by pressure.

Next, the respective actuators 410, 411
A so-called saturation state in which the total flow rate to the variable displacement pump 2 reaches the maximum discharge flow rate of the variable displacement pump 2 will be described. In this case, since the pressure difference between the pump discharge pressure and the maximum load pressure 16 cannot be ensured to be the pressure difference set by the spring 419, the pressure difference before and after the directional control valves 8 and 9 is also increased by the spring. Although the desired differential pressure set by 419 is not obtained, the differential pressures before and after the respective directional control valves become equal, so that the respective actuators 41
The flow rate to 0, 411 is diverted to a flow rate equal to the ratio of the throttle opening of the directional control valves 8, 9 and has a so-called anti-saturation function.

As described above, according to the structure of Japanese Patent Publication No. 4-48967, since the anti-saturation function is provided, the total required flow rate required by each directional control valve exceeds the maximum discharge flow rate of the variable displacement pump. In this case, good simultaneous operability can be obtained.

However, in this type, when the flow rate supplied to the actuators 410 and 411 does not reach the saturation state, the throttle opening degree of the directional control valves 8 and 9, that is, regardless of the load pressure of the actuator and the pump discharge flow rate, It is determined only by the stroke amount of the directional control valves 8 and 9. Therefore, the speed of the actuator is always constant with respect to the stroke amount regardless of the load pressure and the engine speed. However, in a construction machine such as a hydraulic excavator, this may be disadvantageous. For example, in excavation work on a fence, etc., it is necessary to perform a turning motion slowly, but if the speed of the actuator is always constant with respect to the stroke amount, it will be slow unless a fine operation is performed in a narrow stroke range. There was a problem that speed could not be obtained and fine control could not be performed.

Therefore, in Japanese Patent Laid-Open No. 5-99126, a fixed throttle is provided at a discharge port of a fixed displacement pump driven by an engine together with a variable displacement pump, and a differential pressure before and after the fixed throttle is changed. The differential pressure between the pump discharge pressure and the maximum load pressure is controlled so as to be a pressure proportional to the differential pressure before and after the fixed throttle by operating the flow rate adjusting valve for driving the displacement volume changing means. With such a configuration, the differential pressure across the directional control valve is proportional to the differential pressure across the fixed throttle, so the flow rate supplied to the actuator increases or decreases according to the engine speed. Therefore, the engine operability is improved by lowering the engine speed and grading the flow rate gradient with respect to the stroke amount.

Further, in Japanese Patent Laid-Open No. 6-81804, a spring acting on a flow rate adjusting valve for driving a displacement volume changing means of a variable displacement pump and a throttle lever for controlling an engine speed are linked by a link. By interlocking with each other, when the engine speed increases, the spring force increases so that the differential pressure between the pump discharge pressure and the maximum load pressure is controlled to be a pressure proportional to the engine speed. With this configuration as well, the differential pressure across the directional control valve is linked to the movement of the throttle lever, so that the flow rate supplied to the actuator can be increased / decreased in accordance with the engine speed as well. The flow rate gradient with respect to the stroke amount is loosened to improve fine operability.

Further, in Japanese Patent Application Laid-Open No. 6-300002, the pressure compensating valve is of a so-called after-orifice type located between the directional control valve and the actuator, and a spring for biasing the pressure compensating valve in the closing direction is used. The acting force is made weaker as the engine speed increases and becomes stronger as the engine speed decreases. That is, according to the engine speed detected by the engine speed sensor, the electromagnetic proportional pressure control valve is driven by the control device,
The output pressure of the electromagnetic proportional pressure control valve is made to act on the spring that urges the pressure compensating valve in the closing direction, and the differential pressure across the directional control valve is controlled according to the engine speed to improve fine operability. .

[0011]

However, in these systems, the engine speed is reduced to loosen the flow rate gradient with respect to the stroke amount of the directional control valve to improve the fine operability. Since the differential pressure across the directional control valve changes as the number of revolutions increases and decreases, the flow rate gradient with respect to the stroke amount of all actuators becomes gentle when the engine number of revolutions decreases, resulting in a slower overall speed. was there. In particular, as described above, in excavation work on the wall, fine operability of turning and swinging (offset) can be improved, but on the other hand, there is a problem that the flow rate gradient of other actuators becomes gentle and the speed is slow and work efficiency is reduced. there were.

The present invention has been made in view of the above-mentioned conventional problems, and its object is to improve only the flow rate of a specific actuator such as a swing motor according to the work content to improve the fine operability. For the actuator of
An object of the present invention is to provide a hydraulic drive system in which the work efficiency is not reduced by keeping the speed high without making the flow rate gradient with respect to the stroke amount gentle. Another object is to provide a valve used in this hydraulic drive system.

[0013]

In order to solve the above problems, the present invention provides a variable pump, a plurality of hydraulic actuators driven by pressure oil discharged from the variable pump, and a plurality of hydraulic actuators which flow into the plurality of hydraulic actuators. A directional control valve having a flow rate adjusting function capable of controlling each pressure oil, a pressure compensating valve capable of respectively compensating pressures of the plurality of directional control valves, a variable pump discharge pressure and a plurality of hydraulic actuators. Means for outputting a secondary pressure proportional to the differential pressure from the maximum load pressure, the pressure compensating valve being provided upstream of the directional control valve and being closed by the pressure downstream of the pressure compensating valve. Hydraulic drive in which the load pressure and the secondary pressure of the actuator downstream of the directional control valve act independently in the opening direction. In the installation, the secondary pressure is introduced into the inlet port between the secondary pressure introducing port and the secondary pressure oil passage that actuate the pressure compensating valve of the specific actuator requiring fine operability in the opening direction. A constant-ratio pressure reducing valve was provided which was made selectable to output the next pressure or reduce the secondary pressure at a constant ratio. For other actuators, without reducing the secondary pressure with the constant ratio pressure reducing valve,
The pressure compensating valve of each actuator is operated as it is.

The constant-ratio pressure reducing valve has a main body having an inlet port, an outlet port and a drain port, and an inlet port and an outlet port inserted into a body hole provided in the main body, or an outlet port and a drain port. Selectively communicate with the spool, an outlet pilot chamber where the outlet port pressure is guided and urges the spool in the direction to communicate the outlet port and the drain port, and an inlet port pressure is introduced into the spool. An inlet pilot chamber having a pressure receiving area smaller than that of the outlet pilot chamber, which is adapted to urge the port and the outlet port to communicate with each other, and the inlet port pressure is reduced at a constant ratio. In addition to the constant ratio pressure reducing valve that is configured to output, the inlet port pressure introduction and the pressure oil release to the drain port are selectable and the spool is inserted. A port and an outlet port is provided selective pilot chamber which is adapted to bias in a direction that communicates.

As the selectable means for the constant ratio pressure reducing valve, a switching valve such as a solenoid valve, a manually operated valve or a pilot operated valve, which is made selectable between an inlet port and a drain port for the selected pilot chamber, is used.

(Operation) The secondary pressure acting on the pressure compensating valve of the specific actuator, which requires fine operability, is made to act through the constant ratio pressure reducing valve for reducing the pressure at a constant ratio, and the constant ratio pressure reducing valve is operated. Since the valve is selected by the switching valve, the secondary pressure acting on the pressure compensating valve of the specific actuator is reduced by operating the switching valve, and the differential pressure before and after the directional control valve is reduced, and the stroke amount The flow rate gradient with respect to is gentle and the fine operability is improved. For other actuators, the secondary pressure is not reduced by the constant-ratio pressure reducing valve, and the pressure compensation valve of each actuator is acted on as it is, so the speed of the actuator does not decrease and the work efficiency decreases. Never.

Selection of the constant ratio pressure reducing valve When the pilot chamber and the inlet port are communicated with each other, the spool is urged so that the inlet port into which the secondary pressure flows and the outlet port are directly communicated with each other. The secondary pressure is output as is. On the other hand, when the selected pilot chamber of the stoichiometric pressure reducing valve is opened to the drain port with no pressure oil, there is no effect to bias the spool. Is output.

Since either the inlet port or the drain port for the selected pilot chamber can be selected by a switching valve such as a solenoid valve, a manually operated valve or a pilot operated valve, the constant ratio pressure reducing valve can be electrically or manually operated. The output can be easily selected, and the characteristics according to the requirements of normal operation and fine operation can be easily obtained while the speed is increased without loosening the flow rate gradient with respect to the stroke amount.

[0019]

BEST MODE FOR CARRYING OUT THE INVENTION A first embodiment of the present invention will be described with reference to the drawings. FIG. 1 is a hydraulic circuit diagram according to an embodiment of the present invention, and FIG. 2 is a sectional structural view (conceptual diagram) of a constant pressure reducing valve according to the first embodiment and a switching valve (solenoid valve) for driving the constant pressure reducing valve. ) Is an explanatory view showing the relationship with FIG.
In, a plurality of pressure compensating valves 4, 5 are connected in parallel to the discharge oil passages 23, 3 of the variable displacement hydraulic pump 2 driven by the prime mover 1 such as an engine, and the output oil passage 6 of each pressure compensating valve is connected. , 7 are connected to directional control valves 8 and 9 via check valves 26 and 27, respectively, the output sides of these directional control valves are connected to actuators 10 and 11, respectively, and return from the respective actuators 10 and 11 is performed. The oil is returned to the tank 12 again via the respective directional control valves 8 and 9.

Further, the pressure compensating valves 4, 5 are actuated in the closing direction by the pressure on the upstream side of the respective directional control valves 8, 9, that is, the pressure of the output oil passages 6, 7, and further the directional control valves 8, 9 are closed. The pressure on the downstream side, that is, the load pressures 14 and 15 of the actuators 10 and 11 act in the opening direction. Further, among the respective load pressures, a maximum load pressure is selected by the shuttle valve 13, and a pressure control valve 31 for generating a secondary pressure 32 proportional to the differential pressure between the maximum load pressure and the pump discharge pressure is provided in the valve device 22. It is provided.

Here, the specific actuator 10 that requires fine operability is a swing motor, and the actuator 11 that does not require fine operability is a boom cylinder. The secondary pressure 32 from the pressure control valve 31 acts on the pressure compensation valve 5 as it is in the boom cylinder 11 in which fine operability is not required, and the constant pressure reducing valve is used in the actuator 10 in which fine operability is required. The pressure compensation valve 4 is actuated via 41.

The constant ratio pressure reducing valve 41 has an inlet port 42 for introducing the secondary pressure 32, an outlet port 43 for reducing and outputting the secondary pressure 32, and a drain port 44 communicating with the tank port 12, A first pressure-receiving area 41a that causes the output pressure of the outlet port 43 to act in a direction that allows communication between the port 43 and the drain port 44 and blocks communication between the secondary pressure 42, and the secondary pressure 42 and the outlet port. The second pressure receiving area 41b that allows the pressure selected by the solenoid valve 46 to act in a direction that allows the communication with 43 and blocks the communication with the drain port 44, and the second pressure receiving area 41b in the same direction as the second pressure receiving area 41b. It has a third pressure receiving area 41c that constantly applies pressure, and the acting force of the spring 45 acts in the same direction as the second and third pressure receiving areas. The solenoid valve 46 changes the pressure acting on the second pressure receiving area 41b to
It is for communicating with either the secondary pressure or the tank pressure.

Further, the secondary pressure 32 is made to act on the flow rate adjusting valve 18 for driving the displacement volume changing means 17 of the variable displacement pump 2 via the oil passage 33, and is set in advance by the secondary pressure and the spring 19. When the secondary pressure is larger than the acting force of the spring 19 by balancing with the acting force, the secondary pressure is smaller than the acting force of the spring 19 so as to reduce the displacement volume of the variable displacement pump 2. In this case, the displacement of the variable displacement pump 2 is controlled to be large.

The operation of the hydraulic drive system will now be described. When the constant ratio pressure reducing valve 41 is not operating, the pressure compensating valves 4 and 5 are used to control the pressure 6 on the upstream side of the directional control valves 8 and 9. , 7 acts so as to balance the load pressures 14, 15 of the respective downstream actuators and the sum of the secondary pressure 32, the differential pressure across the respective directional control valves 8, 9 is It becomes equal to the secondary pressure 32 regardless of the load pressure. That is, it becomes equal to the differential pressure between the pump discharge pressure and the maximum load pressure.

Further, the secondary pressure 32 depends on the pump device 21.
The secondary pressure 32 of the variable displacement pump 2 is equal to the pressure corresponding to the acting force of the spring 19 because the secondary pressure 32 is guided to the balance of the acting force of the spring 19 of the flow rate adjusting valve 18. Controlled as. This means
The secondary pressure, that is, the differential pressure between the pump discharge pressure and the maximum load pressure is controlled to be equal to the pressure corresponding to the acting force of the spring 19. That is, the pump discharge pressure is controlled to be higher than the maximum load pressure by a pressure corresponding to the acting force of the spring 19. Therefore, the differential pressure across the directional control valves 8 and 9 is controlled to a pressure corresponding to the acting force of the spring 19.

With such a configuration, now,
If the pump discharge flow rate becomes insufficient, the differential pressure between the pump discharge pressure and the maximum load pressure 16, that is, the secondary pressure 32.
Since it becomes impossible to secure the differential pressure amount set by the spring 19, the differential pressure before and after the directional control valves 8 and 9 is also smaller than the set value, but Since the differential pressures are equal to each other, the flow rates to the respective actuators 10 and 11 are divided into the flow rates equal to the ratio of the throttle opening degrees of the directional control valves 8 and 9, and the anti-saturation function is provided. .

Further, when the viscosity of the hydraulic oil becomes high at a low temperature and the pressure loss in the pipe line 23 becomes excessive, the secondary pressure 32 becomes the pump discharge passage in the valve device 22. 3 is generated in proportion to the pressure difference between the maximum load pressure 16 and the maximum load pressure 16, the pump pressure of the pump discharge passage 3 in the valve device 22 is the maximum load regardless of the pressure loss of the pipeline 23. Since the pressure is controlled to a pressure corresponding to the acting force of the spring 19 described above, the flow rate does not decrease significantly even at low temperature, and the actuator speed does not slow down. In addition, the pump device 21
Since it is not necessary to increase the number of pipe lines connecting the valve device 22 and the valve device 22, it is easy to use.

The directional control valves 8 and 9 may be operated by the pilot pressure of the pressure control valve, which increases the pilot pressure according to the operation amount widely used in construction machinery, or may be operated by electromagnetic proportional. It may be operated by a pressure control valve or a high-speed response solenoid valve.

In the hydraulic drive device having such a configuration, the constant ratio pressure reducing valve 4 used in the first embodiment of the present invention.
1 is configured as follows. In FIG. 2, a body hole 102 is formed in the body 101 of the constant ratio pressure reducing valve 41, and a spool 103 is slidably inserted. Into the main body 101, the output pressure reduced by the constant ratio pressure reducing valve from the main body hole end surface 104 is introduced from the left end in FIG. 2, and the main body hole 102 and the outlet pilot chamber constituted by the main hole end surface and the end surface of the spool 103 106, a first large-diameter portion 107 (inner diameter D) on which the spool slides in an oil-tight manner, an inlet port 42 into which the secondary pressure is introduced, and a second large-diameter portion that communicates with the output pressure outlet port 43. 110, a drain port 44 connected to the tank line 12 (T), a third large diameter portion 112 on which the spool slides in an oil-tight manner, and a spring 45 provided in contact with the other end surface 113 of the spool. Interpolated and selected port 1
A selection pilot chamber 116 communicating with 15 is sequentially formed.

The selection port 115, which communicates with the selection pilot chamber 116, is connected to the solenoid valve 46 so that the solenoid valve 46 communicates with either the secondary pressure 32 or the tank line 12 that communicates with the drain port 44. . That is, in FIG. 2, when the solenoid valve 46 is not excited, the secondary pressure 32 is guided to the selected pilot chamber 116,
When the solenoid valve 46 is excited, the selected pilot chamber communicates with the tank line 12.

Further, a small-diameter portion 117 in which a piston 114 having a smaller diameter than the large-diameter portion 112 can be slidably inserted in an oil-tight manner.
(Inner diameter d) is provided so that the piston can come into contact with the spool 103. Following the small diameter portion 117, an inlet pilot chamber 119 into which the secondary pressure is constantly introduced is formed by the end portion of the piston 114 and the main body. The inlet pilot chamber 119 is in communication with the secondary pressure via the inlet pilot port 118. The large diameter portions 107, 11
0 and 112 have the same diameter.

The spool 103 is a land portion corresponding to the first large-diameter portion 107 provided in the main body hole and can be opened and closed between the inlet port 42 for introducing the secondary pressure and the outlet port 43. 120 having a narrowed portion 108
And a small diameter portion 121, and a land portion corresponding to the third large diameter portion 112, and a land portion having a narrowed portion 111 that can be opened and closed between the outlet port 43 and the drain port 44 communicating with the tank line 12. 122 are sequentially formed.

Further, the spool 103 is provided with a communication passage 123 which opens to the small diameter portion 121 and communicates the outlet pilot chamber 106 with the outlet port 43. 2, the spring 45 is in contact with the right end surface 113 of the spool 103 as described above, but the right end surface of the spring is also in contact with the end surface 125 of the selection pilot chamber 116, and at the same time, the spring 45 is contacted. The piston 114 is located inside the
Is interpolated. Further, the piston 114 is pushed leftward by the secondary pressure introduced to the inlet pilot chamber and is in contact with the right end surface 113 of the spool 103.

Here, the communication relationship of each port will be described in detail.
If the spool 103 makes a maximum stroke to the left,
The end surface 105 of the spool 103 comes into contact with the end surface 104 of the main body hole, the communication opening degree of the secondary pressure inlet port 42 and the outlet port 43 becomes maximum, and the drain port 44 communicating with the tank line 12 is blocked. On the contrary, when the maximum stroke is made in the right direction, the right end surface 126 of the piston 114 contacts the end surface 127 of the inlet pilot chamber 119 of the main body 101, and the communication opening degree between the outlet port 43 and the drain port 44 becomes the maximum. The secondary pressure inlet port 42 is shut off.
The communication between the inlet port 42 and the drain port 44 with respect to the outlet port 43, that is, the positions of the throttles 108 and 111 are such that one is in communication and the other is blocked at the same time, that is, a zero lap state, or a state where there is some overlap, Or, it is in a state with some underlap.

Note that FIG. 2 is for conceptually showing the operating principle. Although both ends of the main body hole 102 are not open, in reality, a stepped through hole or a stepped process from the left side surface is formed. It goes without saying that the structure is formed as a hole and is appropriately closed by a method such as a screw plug.

Next, the operation of the first embodiment will be described. First, consider a state in which the solenoid valve 46 is not excited. In this case, the exit pilot room 106
The output pressure of the outlet port 43, the inlet pilot chamber 11
The secondary pressure 32 is introduced into the 9 and the selected pilot chamber 116. In FIG. 2, the spool 103 of the constant ratio pressure reducing valve
Considering the balance of the forces acting on the spool 103, the force Fa acting on the spool 103 in the right direction is reduced by the constant ratio pressure reducing valve with the first pressure receiving area 41a due to the outer diameter D of the spool 103 of the outlet pilot chamber 106 being A1. If the output pressure of the outlet port 43 is Pco, then Fa = A1.Pco (1).

The force Fb acting in the leftward direction is determined by the outer diameter D of the spool 103 in the selected pilot chamber 116 and the piston 114.
The second pressure receiving area 41b due to the diameter difference of the outer diameter d of A is set to A2, and the outer diameter d of the piston 114 in the inlet pilot chamber 119 is set to A2.
When the third pressure receiving area 41c by A3 is A3 and the acting force of the spring 45 is W, Fb = A2.Pc + A3.Pc + W (2)

If the forces acting on the spool 103 in both directions are balanced, the equations (1) and (2) are equal, so Fa = Fb, that is, A1.Pco = A2.Pc + A3.Pc + W ....
(3). Here, by substituting the relationship of A1 = A2 + A3 and rearranging, Pco = Pc + W / A1 (4), but if the output pressure Pco is the secondary pressure P of the supply pressure,
Even if it becomes equal to c, the formula (4) is not established, and the force acting on the spool 103 in the left direction is increased by the acting force W of the spring 45, and as a result, the spool 103 makes a maximum stroke in the left direction. Therefore, the inlet port 42 for the secondary pressure and the outlet port 43 for the output pressure 43 communicate with each other at the maximum opening degree, and Pco = Pc (5) That is, in this case, the secondary pressure is applied to the pressure compensating valve without being reduced.

Next, consider a state in which the solenoid valve 46 is excited. In this case, the output pressure of the outlet port 43 is introduced into the outlet pilot chamber 106, the secondary pressure is introduced into the inlet pilot chamber 119, and the selected pilot chamber 116 is opened to the tank, and there is almost no pressure. The force Fa that acts on the spool 103 in the right direction is the same as that in the expression (1). The force Fc acting on the spool 103 in the leftward direction is generated by the selected pilot chamber 116.
Since the (second pressure receiving area A2) is connected to the tank line by the solenoid valve 46, Fc = A3.Pc + W (6) Here, from equation (1) = equation (6), that is, Fa = Fc, Pco = (A3 · Pc + W) / A1 (7) is obtained. Therefore, the output pressure Pco is the first pressure receiving area A1 due to the outer diameter D of the spool 103, the third pressure receiving area A3 due to the outer diameter d of the piston 114, and the secondary pressure Pc.
The pressure is reduced to a value determined by the acting force W of the spring 45.

Therefore, as for the pressure acting in the direction of opening the pressure compensating valve 4 of the swing motor 10 which requires fine operability, the secondary pressure Pc acts as it is when the solenoid valve 46 is not excited, When 46 is excited, the output pressure Pco, which is the secondary pressure reduced, acts, and the flow rate gradient with respect to the stroke amount of the directional control valve 8 is loosened, the speed of the swing motor 10 is slowed, and fine operability is improved. Can be improved. At this time, the ratio of the flow rate gradient when the solenoid valve 46 is not excited and when it is excited is determined by the spool 1 described above.
The first pressure receiving area A1 due to the outer diameter D of 03 and the piston 1
It is determined by the third pressure receiving area A3 by the outer diameter d of 14 and the acting force W of the spring 45. If the ratio of the flow rate gradient is 1: 2, those values may be selected so that the ratio of Pco and Pc is 1: 4.

The spool 103 and the piston 114 may be integrated or separated. In addition, the integrated pilot chamber 115 and the entrance pilot chamber 118 are integrated.
And may be interchanged in position. Further, the acting force of the spring does not need to be so strong and may be such that it overcomes the sliding resistance of the spool 103 and the fluid reaction force generated in the throttle portion 108.

Next, a second embodiment of the present invention will be described with reference to the drawings. FIG. 3 is an explanatory diagram showing a relationship between a sectional structure diagram (conceptual diagram) of the constant ratio pressure reducing valve of the second embodiment and a switching valve (solenoid valve) for driving the constant ratio pressure reducing valve.
The same parts as those in the first embodiment described above are designated by the same reference numerals, and a part of the description will be omitted. The second embodiment is
In contrast to the first embodiment, a communication method for guiding the output pressure Pco to the outlet pilot chamber 206 (pressure receiving area A1),
The configuration of the inlet pilot chamber 219 (pressure receiving area A3) and the communication method for guiding the secondary pressure to the selected pilot chamber 216 are changed. In the first embodiment, in order to guide the output pressure Pco to the outlet pilot chamber, the spool 103
Although the output pressure Pco is guided by the communication hole 123 opened in the axis center of the shaft, in the second embodiment, the output pressure introducing port 207 provided in the main body 201 is provided to guide the output pressure Pco to the outlet pilot chamber 206. .

Further, the inlet pilot chamber 219 is provided with a shaft hole 208 on the right end surface of the spool 203, and a piston 209 is slidably inserted into the shaft hole 208, so that the shaft hole 208 and the piston 20 are connected.
9 is formed on the left end face. Further, the selection pilot chamber 216 (pressure receiving area A2) is formed by bringing the right end surface of the piston 209 into contact with the right end surface 211 of the main body 201 and inserting the piston 209 to provide the spring 45. Further, a small diameter portion 212 is provided at a position corresponding to the secondary pressure inlet port 42 of the spool 203, and a communication passage 213 is provided at the shaft center of the spool, one of which is opened to the small diameter portion and the other of which is opened to the inlet pilot chamber 219. And guides the secondary pressure to the inlet pilot chamber 219. The small diameter portion 212 has a positional relationship in which the communication with the inlet port 42 for the secondary pressure is not blocked even if the spool 203 has a maximum stroke in either the left or right direction. Further, in the spool 203, the left end surface of the piston 209 contacts the right end surface 215 of the shaft hole 208 of the spool 203 when the maximum stroke is made in the right direction.

With this structure, too, the outer diameter D of the spool 203, the outer diameter d of the piston 209, and the spring 4
If the acting force W of 5 is the same as in the first embodiment,
The same effect as described above can be obtained. According to this one,
Since the sliding part of the piston 209 is the end of the spool shaft and there is no need to make a step in the main body hole, the main body can be easily processed, and various pressure-receiving areas can be easily replaced by replacing the spool and piston. Can be obtained.

Next, a third embodiment of the present invention will be described with reference to the drawings. FIG. 4 is an explanatory diagram showing a relationship between a sectional structure diagram (conceptual diagram) of the constant ratio pressure reducing valve of the third embodiment and a switching valve (solenoid valve) for driving the constant ratio pressure reducing valve.
The same parts as those in the first embodiment described above are designated by the same reference numerals, and a part of the description will be omitted. The third embodiment is
The configuration of the selection pilot chamber 316 is changed from that of the first embodiment. In the first embodiment, a three-way solenoid valve is used to select the inlet port (secondary pressure) and the drain port, but in the third embodiment, a two-way solenoid valve is used. The secondary pressure is guided to the selective inlet port 304 of the selective pilot chamber 316 via 313, and the selective outlet port 302 of the selective pilot chamber is made to communicate with the tank line 12 via the two-way solenoid valve 346.

According to this structure, when the electromagnetic valve 346 is not excited, the pressure in the selective pilot chamber 316 becomes equal to the secondary pressure because the selective outlet port 302 is closed by the electromagnetic valve 346, and the electromagnetic valve When 346 is excited, the selected pilot chamber communicates with the tank line, so the secondary pressure is throttled by the fixed throttle 313 and the selected pilot chamber becomes low pressure. Therefore, the outer diameter D of the spool 103, the outer diameter d of the piston 114, and the acting force W of the spring 45 are
If the same operation as that of the first embodiment is performed, the same operation as described above can be obtained.

In the second embodiment, as in the third embodiment, the secondary pressure is communicated with the selected pilot chamber through the throttle, and the drain port or the tank line is connected to the two-way solenoid valve 346. It goes without saying that the selected pilot room may be communicated with via the selected pilot room. Although the solenoid valve is used as the switching valve that applies pilot pressure, the switching valve may be a solenoid valve, a manually operated valve, or a pilot operated valve. Good.

Although the basic circuit example shown in FIG. 1 is shown, the pressure compensation valve of the present invention can be applied to other than the above circuit example. That is, as described above, if the load compensating valve and the secondary pressure Pc act in the opening direction, and the upstream side pressure of the directional control valve acts in the closing direction to control, the secondary pressure is the secondary pressure. Pc may be means such as an electromagnetic proportional pressure control valve. An example of such a circuit is
For example, it can be applied to those disclosed in JP-A-1-266301. Further, the maximum load pressure may be applied instead of the secondary pressure applied to the pump control circuit.

In the embodiment of the invention described above, a specific actuator which requires particularly fine operability is shown as one, but this is the same even if there are two or more actuators.
That is, the output pressure Pco of one constant ratio pressure reducing valve may be connected in parallel to the pressure compensating valve of the actuator that requires fine operability. In some cases, it may be connected to the pressure compensation valves of all actuators. Further, when the timings at which the fine operability is required for the individual actuators are different, the constant ratio pressure reducing valve and the switching valve may be provided independently of the pressure compensating valve of each actuator. In this case, it is also possible to set the outer diameter D of the spool, the outer diameter d of the piston, and the acting force W of the spring by changing each constant ratio pressure reducing valve to make the flow rate gradient different.

Further, the constant ratio pressure reducing valve and the switching valve of the present invention are
The above-mentioned JP-A-5-99126 and JP-A-6-81.
It is also possible to use it together with Japanese Patent No. 804. That is, in this case, in addition to making the flow rate gradient with respect to the stroke amount of the directional control valve gentle as the engine speed increases and decreases, the flow rate gradient is made gentler by the configuration of the present invention as necessary. It is possible to improve operability.

[0051]

According to the present invention, there are two pressure receiving areas in the opening direction of the pressure compensating valve, one of which exerts a load pressure and the other of which exerts a secondary pressure Pc to counteract their acting forces. In a hydraulic drive system using a pressure compensating valve configured to actuate the pressure on the upstream side of the directional control valve, the secondary pressure acting on the pressure compensating valve for a particular actuator that requires particularly fine operability is fixed Since the operation is performed via the pressure reducing valve and the operation of the constant ratio pressure reducing valve is switched by the switching valve, the secondary pressure acting on the pressure compensating valve of the specific actuator is reduced by operating the switching valve. The differential pressure before and after the directional control valve becomes small, the flow rate gradient with respect to the stroke amount becomes gentle, and the fine operability is improved. With respect to the other actuators, the secondary pressure is not reduced by the constant ratio pressure reducing valve and is allowed to act on the pressure compensating valve of each actuator as it is, so that the speed of the actuator does not decrease and the working efficiency is improved. It has an excellent effect of not dropping.

Further, the constant ratio pressure reducing valve and the switching valve of the present invention may be connected in parallel to the pressure compensating valves of two or more actuators, or in some cases, may be connected to the pressure compensating valves of all the actuators. Good. Further, when the timings at which fine operability is required for the individual actuators are different, the pressure compensation valves of the individual actuators may be individually provided. In addition, it can be used in combination with an example of a circuit in which the differential pressure across the directional control valve increases or decreases in conjunction with the engine speed, such as the conventional hydraulic drive system. You can get it.

[Brief description of drawings]

FIG. 1 is a hydraulic circuit diagram of a hydraulic drive system according to first to third embodiments of the present invention.

FIG. 2 is an explanatory diagram showing a conceptual view of a constant ratio pressure reducing valve according to the first embodiment of the present invention.

FIG. 3 is an explanatory diagram showing a conceptual view of a constant ratio pressure reducing valve according to a second embodiment of the present invention.

FIG. 4 is an explanatory diagram showing a conceptual diagram of a constant ratio pressure reducing valve according to a third embodiment of the present invention.

FIG. 5 is a hydraulic circuit diagram of a conventional hydraulic drive system.

[Explanation of symbols]

2 Variable pump 3 Variable pump Discharge pressure 4, 5 Pressure compensation valve 6, 7 Pressure compensation valve downstream side (direction control valve upstream side) 8, 9 Directional control valve 10 Hydraulic actuator (swing motor) 11 Hydraulic actuator (boom cylinder) 14 , 15 Actuator load pressure (Direction control valve downstream side pressure) 16 Maximum load pressure 31 Means for outputting secondary pressure (pressure control valve) 32 Secondary pressure (secondary pressure oil passage) 41 Constant ratio pressure reducing valve 42 Inlet Port 43 Outlet port 44 Drain port 46, 346 Switching valve (selectable means) 52 Secondary pressure inlet 101, 201, 301 Main body 102, 202 Main body hole 103, 203 Spool 106, 206 Outlet pilot chamber 116, 216, 316 Selection pilot room 119, 219 Entrance pilot room

Claims (4)

[Claims] 1. A variable pump, and a plurality of hydraulic actuators driven by pressure oil discharged from the variable pump,
A directional control valve having a flow rate adjusting function capable of controlling the pressure oil flowing into each of the plurality of hydraulic actuators,
A pressure compensating valve that enables pressure compensation of each of the plurality of directional control valves; a means for outputting a secondary pressure proportional to a differential pressure between the variable pump discharge pressure and the maximum load pressure of the plurality of hydraulic actuators; The pressure compensating valve is provided on the upstream side of the directional control valve, and is operated in the closing direction by the pressure on the downstream side of the pressure compensating valve, and further, the load pressure of the actuator on the downstream side of the directional control valve. And in the hydraulic drive device in which the secondary pressure is made to act independently in the opening direction, the secondary pressure of the secondary pressure that acts in the opening direction of the pressure compensating valve of the specific actuator that requires fine operability is A constant-ratio depressurization made selectable to introduce the secondary pressure into the inlet port between the inlet and the oil passage of the secondary pressure, and output the secondary pressure or output the secondary pressure in a reduced pressure at a constant ratio. Hydraulic drive system, wherein a provided. 2. A body having an inlet port, an outlet port, and a drain port, and the inlet port and the outlet port inserted into a body hole provided in the body, or the outlet port and the drain port. And an outlet pilot chamber adapted to urge the spool in a direction in which the outlet port pressure is guided so that the spool communicates with the outlet port and the drain port, and the inlet port pressure. And an inlet pilot chamber having a pressure receiving area smaller than that of the outlet pilot chamber, the inlet pilot chamber being configured to urge the spool in a direction in which the inlet port and the outlet port communicate with each other. In the constant-ratio pressure reducing valve, which is configured to output a pressure reduced at a constant ratio to the outlet port, introducing the inlet port pressure and drain port Hydraulic oil discharge device having a selection pilot chamber configured to be capable of selecting pressure oil release to a port and biasing the spool in a direction in which the inlet port and the outlet port communicate with each other. Constant ratio pressure reducing valve. 3. The selectable means of the constant ratio pressure reducing valve for a hydraulic drive system is a switching valve capable of selecting either an inlet port or a drain port for a selected pilot chamber. The constant ratio pressure reducing valve according to item 2. 4. The constant ratio pressure reducing valve according to claim 1 is the constant ratio pressure reducing valve according to claim 2 or 3.
The hydraulic drive described.