JPH1089304A - Hydraulic driving device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a hydraulic driving device which is excellent in operability on both of low load side and high load side, is stable without generating hunting, becomes high in viscosity of hydraulic operating fluid at the time of a low temperature and is not remarkably reduced in pump flow rate even if pressure loss in a conduit from a pump to a valve becomes excessive, in the hydraulic driving device having a plurality of hydraulic actuators such as hydraulic excavators. SOLUTION: In a hydraulic driving device having a variable pump 2, each of a plurality of hydraulic actuators 10, 20, directional control valves 8, 18 and pressure compensating valves 4, 14, flow rate of the pressure compensating valves leading to the actuators is made to be reduced according to an increase of load pressure of each actuator. Further, a differential pressure control valve 31 generating secondary pressure corresponding to differential pressure of discharge pressure (Pd) of the variable pump 2 and maximum load pressure (Pm) of the actuators 10 and 20 is provided. Action force of a spring 19 is made to act in a direction opening and closing a pump flow control valve 38 and increasing and decreasing displacement volume of the variable pump.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a plurality of direction controls having a flow rate adjusting function capable of controlling pressure oil flowing into each of a plurality of actuators from discharge oil of one hydraulic pump used in a construction machine or the like. The present invention relates to a hydraulic drive device having a valve and a plurality of pressure compensating valves for compensating the pressure of each directional control valve.

[0002]

2. Description of the Related Art This type of hydraulic drive device is mainly used for construction machines and agricultural machines, and has a load sensing function of controlling a variable pump discharge amount in accordance with a load pressure. Also, when driving a plurality of actuators at the same time, a pump compensating valve is provided in a circuit to each actuator so that the speed of the actuators does not change due to a difference in load pressure between the actuators. Are diverted. Further, when the pump discharge amount falls below a predetermined required flow rate of a plurality of drive actuators, a device having a function of distributing the pump discharge amount to each actuator at an appropriate ratio, a so-called anti-saturation function, is also used. For example, in the device disclosed in Japanese Patent Publication No. 4-48967 having such an anti-saturation function, the load sensing function applies a spring force and the maximum load pressure of each actuator in a direction to increase the variable pump discharge amount. A pump flow control unit is provided which applies a discharge pressure in a direction to decrease the variable pump discharge amount in opposition to the acting force, and controls the pump discharge amount according to the load pressure. On the other hand, as described in, for example, Japanese Patent Application Laid-Open No. 19409/1992, a pressure compensating valve is disposed downstream of a directional control valve having a flow rate adjusting function, and the pressure compensating valve is opened in each control oil chamber of the pressure compensating valve. An apparatus having an anti-saturation function is disclosed in which the pressure (Pd ') on the upstream side of the pressure compensating valve is applied, and the maximum load pressure (Pm) of the actuator is applied in the direction in which the pressure compensating valve is throttled. . In Japanese Utility Model Laid-Open Publication No. 4-119604, the pressure (Pz) on the downstream side of the pressure compensating valve is applied to each control oil chamber of the pressure compensating valve in the direction in which the pressure compensating valve is closed so that the variable pump discharge pressure (Pd) is reduced. Secondary pressure (Pc =
Pd-Pm) is provided, and a secondary pressure (Pc) output from the differential pressure control valve in a direction to open the pressure compensating valve and an actuator load pressure (pressure downstream of the directional control valve). PL) is also described. Further, FIG. 6 of JP-A-4-54303 shows a differential pressure detector for detecting the variable pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator, and a control signal in response to the output of the differential pressure detector. And an apparatus that secures an anti-saturation function by a secondary pressure (Pc) output by an electromagnetic proportional valve that is operated by a control signal of the control apparatus.

[0003] The flow characteristics of the flow compensating region of such a conventional pressure compensating valve having an anti-saturation function are such that the actuator load pressure (PL) and the upstream pressure () of the directional control valve are determined by the balance between the open side and the closed side. Pz), that is, the relationship between the differential pressure before and after the directional control valve (ΔP: hereinafter referred to as the directional control valve differential pressure) is represented by ΔP = Pz−PL = Pd−Pm = Pc. The differential pressure is a pressure proportional to the differential pressure between the variable pump discharge pressure and the maximum load pressure of the actuator, that is, the secondary pressure. However, it has been known that, in a hydraulic drive device that performs such pressure compensation, when a load having a large inertia is moved, the system is not stabilized and the operation is accompanied by hunting. For example, when a hydraulic shovel of a construction machine is moved, when a large load cylinder such as a swing motor, a traveling motor, or a boom cylinder is moved, hunting is remarkably exhibited, and operability is impaired. There was a problem. That is, consider a case in which the lever of the direction control valve is moved stepwise by a fixed amount to move an actuator having a large inertia, that is, a swing motor. First, the throttle of the directional control valve is opened and oil flows in, but since the inertia of the actuator is large, the actuator does not move suddenly, so that the flow rate does not flow and the load pressure rises instantaneously. When the load pressure rises, the load pressure acts on the pressure compensating valve and acts to open the pressure compensating valve greatly. As a result, a large flow rate is supplied and the actuator is rapidly accelerated. However, once the actuator starts moving, the supplied flow rate is finite, so that the speed increases but the acceleration gradually decreases. Therefore, the load pressure that has rapidly risen gradually decreases as the acceleration decreases, so that the opening of the pressure compensating valve gradually decreases, and the supplied flow rate decreases. On the other hand, when the actuator loses its acceleration and reaches a constant speed, the speed is accelerated by a large initial acceleration, so that the speed is considerably higher than the target speed, and the acceleration of the load pressure at that time is stopped. After that, the opening degree of the pressure compensating valve has become smaller, and hence the directional control valve differential pressure has also become smaller. This will cause the flow rate to decrease and the actuator to begin to decelerate, but due to the large inertia, the actuator will attempt to maintain its speed, causing the load pressure to decrease more rapidly. Therefore, the opening of the pressure compensating valve becomes smaller and the speed of the actuator decreases more. In the meantime, the deceleration of the actuator stops and the speed becomes constant, but the speed suddenly decreases at the beginning of the deceleration, so that it is considerably lower than the target speed,
Since the load pressure at that point has recovered to a large value since the deceleration has subsided, the opening of the pressure compensating valve is also large, and therefore the directional control valve differential pressure is also large. Thus, the actuator will now start accelerating.

[0004] Once the actuator starts accelerating, it is in the initial state described above. Thus, the phenomenon of repeated rapid acceleration and rapid deceleration does not readily converge, and hunting continues. Actually, the response delay on the pump device side is further added, so that a more complicated aspect is exhibited. In such a circuit that performs pressure compensation, there is a problem that when a load having a large inertia is moved, the system is not stable and hunting is likely to occur. However, as mentioned above,
No. 7 and Japanese Utility Model Application Laid-Open No. 4-119604 disclose that the pressures acting on the control chambers of the pressure compensating valve in the opening and closing directions are made equal and have an anti-saturation function. However, there is no disclosure or suggestion of hunting prevention. In Japanese Patent Application Laid-Open No. 4-54303, hunting and instability of the pressure compensating valve caused by a decrease in the flow rate caused by the flow force generated by the throttle portion of the pressure compensating valve when a plurality of actuators are operated at the same time can be controlled with a very small pressure receiving area. The discharge pressure of the pump acts in the direction in which the pressure compensation valve opens, and when the differential pressure between the discharge pressure of the pump and the load pressure of the actuator increases, the outlet flow rate of the pressure compensation valve is increased to cancel the flow force and flow. Although a technique for obtaining an outlet flow rate that is not affected by force is disclosed, there is no disclosure or suggestion about prevention of hunting for a load having a large inertia. Further, in each of the prior arts described above, the maximum load pressure (Pm) is caused to act in a direction to close the pump flow control valve for changing the displacement of the pump from the valve device via a long and narrow pilot oil passage. Therefore, at low temperatures, the viscosity of the hydraulic oil increases, and if the pressure loss in the pipeline from the pump to the valve device becomes excessive, the pressure on the upstream side of the pressure compensating valve in the valve device decreases by the pressure loss. As a result, the directional control valve differential pressure was reduced, and the flow rate supplied to the actuator was significantly reduced. Also, at least two actuators among a plurality of actuators need to be driven synchronously regardless of the load pressure of the actuators, such as driving two traveling motors that rotationally drive a pair of crawlers of a hydraulic traveling vehicle. At the same time, if each directional control valve is switched by the same stroke, the directional control valve differential pressure before and after each directional control valve is controlled by the pressure compensating valve so that the same flow rate is supplied to each traveling motor. Therefore, the hydraulic traveling vehicle should be able to travel straight, but if there is a processing error in the spool of each directional control valve, even if the directional control valve differential pressure is equal, the throttle opening of each directional control valve will differ. Therefore, the flow rates supplied to the respective traveling motors are not equal, and if there is an error in the pressure receiving area due to a processing error of the pressure compensating valve, the throttle opening of the directional control valve is equal. Also, since not equal the direction control valve differential pressure, there is a problem that the straight running of the hydraulic traveling vehicle can not.

Further, when simultaneously operating at least two hydraulic actuators having greatly different loads, such as a hydraulic motor for turning a hydraulic excavator and a hydraulic boom cylinder, an inertial load of a high-load-side actuator is initially set at the same time. Is too large, an excessive pressure is generated in the inflow-side actuator port, and most of the pressure oil flows out of the overload relief valve installed in the inflow-side actuator port into the tank, thereby reducing the effective discharge flow rate itself However, there is a problem that the driving speed of the boom cylinder, which is a low-load-side hydraulic actuator, becomes extremely slow, and at the same time, the energy loss of the prime mover due to the relief of the pressure oil flowing into the swing motor increases. Thereafter, when the rotation of the swing motor is completed and the swing is performed at a steady speed, the load pressure of the swing motor rapidly decreases. The pressure compensating valve for the swing motor was almost fully opened in the initial stage due to excessive swing load pressure, but its opening rapidly decreases as the load pressure suddenly decreases. Therefore, the swing motor decelerates with a shock. When the pump (effective) discharge amount has a margin with the deceleration, the boom accelerates conversely, causing a jerky movement.

[0006]

SUMMARY OF THE INVENTION The object of the present invention described in claims 1, 3 to 18 of the present invention has been made in view of the above-mentioned conventional problems.
An object of the present invention is to provide a hydraulic drive device having a stable pressure compensating valve that does not cause hunting and has good operability on both the low load side and the high load side even in a combined operation. It is another object of the present invention to provide a hydraulic drive device having a pressure compensating valve which has a simple structure, is low in cost, has high reliability, and can flexibly respond to load conditions. Claim 2 of the present invention
The object of the described invention is to provide a hydraulic oil that prevents the viscosity of hydraulic oil from increasing at low temperatures, resulting in excessive pressure loss in the pipeline from the pump to the valve device, and preventing a large decrease in the supply flow rate to the actuator. It is to provide a driving device. An object of the invention according to claims 7 and 16 of the present invention is to rotate a pair of crawlers of a hydraulic traveling vehicle.
When at least two actuators of a plurality of actuators need to be driven synchronously irrespective of the load pressure of the actuators as in the case of driving a plurality of traveling motors, the processing error of the spool of each directional control valve, pressure An object of the present invention is to provide a hydraulic control device having a pressure compensating valve capable of obtaining a straight traveling of a hydraulic traveling vehicle even when there is an error in a pressure receiving area due to a processing error of a compensating valve. An object of the invention according to claims 8, 9, 17 and 18 of the present invention is to sufficiently supply hydraulic oil to a small load side even if actuators having extremely different load sizes are operated simultaneously. Even if the load pressure of the actuator with the larger load suddenly drops, the speed of each actuator does not suddenly change, enabling smooth operation without shock, energy loss and load on the prime mover It is an object of the present invention to provide a hydraulic drive device having a pressure compensating valve capable of reducing pressure.

[0007]

According to a first aspect of the present invention, a variable pump and a plurality of hydraulic actuators driven by discharge oil of the variable pump are provided. And a plurality of directional control valves having a flow control function capable of controlling the pressure oil flowing into each of the plurality of actuators, and a plurality of pressure compensating valves for compensating the pressure of each of the directional control valves. , Each of the pressure compensating valves is arranged so that the pressure (P
z) and the maximum load pressure (Pm) of the plurality of actuators respectively, and a pump discharge pressure (P) which is a pressure on the upstream side of the pressure compensation valve in the opening direction of the pressure compensation valve.
d) and the actuator load pressure (PL), which is the downstream pressure of the directional control valve, is applied to compensate for the pressure, and the pump flow rate is adjusted so that the discharge oil of the variable pump communicates with the displacement changing means of the variable pump. Having the valve actuate the maximum load pressure (Pm) through the oil passage in the direction of closing the pump flow regulating valve and increasing the displacement of the variable pump together with the acting force of the spring of the pump flow regulating valve; In a hydraulic drive device in which the pump discharge pressure (Pd) is acted on in a direction to reduce the displacement of the variable pump by opening the pump flow regulating valve through another oil passage, the load pressure of each actuator is increased. The present invention provides a hydraulic drive device characterized in that the flow rate of the pressure compensating valve connected to the actuator is reduced in accordance with It was resolved.

[0008]

According to the structure of the first aspect, the flow rate of the pressure compensating valve connected to the actuator is reduced in accordance with the increase in the load pressure of the actuator, that is, the differential pressure of the directional control valve is reduced. Even if there is a sudden change in the self-load pressure, the actuator load pressure is attenuated and the hydraulic control system is stabilized, and the pressure compensation characteristics of the pressure compensation valve which is not affected by the maximum load pressure of the actuator or the discharge pressure of the variable pump can be obtained. In any of the single operation and the combined operation, both the low load side and the high load side have an unprecedented excellent effect of obtaining stable operability without hunting. Furthermore, the pressure compensation characteristic of the pressure compensation valve can be easily changed to a so-called right-downward characteristic, which indicates the degree to which the flow rate of the pressure compensation valve is reduced in accordance with an increase in the load pressure of each actuator by simply changing the components inside the pressure compensation valve. It can be set, and a right-down characteristic that does not cause hunting according to the load characteristic of each actuator can be obtained.
In addition, since the structure of the pressure compensating valve itself is simple, strict accuracy is not required, and cost reduction and high reliability can be obtained.

[0009]

According to a second aspect of the present invention, a variable pump, a plurality of hydraulic actuators driven by discharge oil of the variable pump, and a plurality of hydraulic actuators are provided. A plurality of directional control valves having a flow rate adjusting function capable of controlling the pressure oil flowing into each of the actuators, a plurality of pressure compensating valves for compensating the pressure of each of the directional control valves, and the variable pump discharge pressure ( P
d) a differential pressure control valve for generating a secondary pressure (Pc = Pd-Pm) corresponding to the differential pressure between the maximum load pressure (Pm) of the actuator and the displacement of the variable pump from the discharge oil of the variable pump. And a pump flow regulating valve communicating with the changing means. Each of the pressure compensating valves applies a pressure (Pz) downstream of the pressure compensating valve in a direction to close the pressure compensating valve, and acts in a direction to open the pressure compensating valve. In the hydraulic drive device, the secondary pressure (Pc) output from the differential pressure control valve and the actuator load pressure (PL), which is the downstream pressure of the direction control valve, are applied to compensate for the pressure. The acting force of the spring of the pump flow control valve acts in a direction to close the pump flow control valve and increase the displacement of the variable pump, and the secondary pressure (Pc) is applied to the pump flow of the variable pump via an oil passage. Adjustment Solves the above problems by providing a hydraulic drive device, characterized in that an acting in a direction to reduce the displacement of the variable pump Open.

[0010]

According to the configuration of the second aspect of the present invention, the secondary pressure (Pc) is made to act in the direction of opening the pump flow rate regulating valve of the variable pump via the oil passage to reduce the displacement of the variable pump. Therefore, as in the prior art, the maximum load pressure (Pm) is applied from the valve device to the pump flow regulating valve for changing the displacement of the pump via the long and narrow pilot oil passage in the closing direction, and the pump discharge pressure (Pm) is changed. Compared with Pd) acting in the direction of opening the pump flow control valve via another oil passage, the second invention of the present invention has a higher viscosity of the working oil at low temperatures, and the pressure from the pump to the valve device is increased. Even when the pressure loss in the pipeline becomes excessive, the secondary pressure (Pc) of the oil passage is the differential pressure between the pump discharge pressure (Pd) and the maximum load pressure (Pm) in the valve device. To Generated secondary pressure (Pc), the pump discharge pressure (Pd) of the pump discharge oil passage in the valve device becomes the maximum load pressure (Pm) regardless of the magnitude of the pressure loss of the pump discharge line. ), The pressure is controlled to the pressure corresponding to the acting force of the spring of the pump flow control valve. Therefore, even at a low temperature, unlike the conventional device, the flow rate does not decrease significantly, and the speed of the actuator decreases. Nothing.

Preferably, in the hydraulic drive device according to the third aspect of the invention, instead of applying the secondary pressure (Pc) in a direction to open the pump flow control valve, the hydraulic drive device is used. Pump flow control valve closes the pump flow control valve for driving the displacement changing means of the variable pump through the oil passage by increasing the maximum load pressure (Pm) in the direction of increasing the displacement of the variable pump. The pump discharge pressure (Pd) may be applied in a direction to open the pump flow control valve via another oil passage to decrease the displacement of the variable pump. More preferably, in the invention according to the third aspect, as in the inventions according to the fifth, sixth, and tenth to eleventh aspects, the flow rate of the pressure compensating valve communicating with the actuator according to the increase in the load pressure of the actuator is adjusted. The pressure compensating valve is provided upstream of each of the corresponding directional control valves, and the pressure compensating valve applies its own downstream outlet pressure to the first pressure receiving area of the first oil chamber. Acting in the direction of closing the valve, applying the secondary pressure to the second pressure receiving area of the second oil chamber in the direction of opening the valve, and applying the secondary pressure to the third pressure receiving area of the third oil chamber. The load pressure is applied in a direction to open the valve, and the second and third pressure receiving areas are made substantially the same, and the first pressure receiving area is made larger than the third pressure receiving area.

As in the invention according to claims 7 and 16, at least two of the plurality of actuators are actuators, such as two traveling motor drives that rotationally drive a pair of crawlers of a hydraulic traveling vehicle. When it is necessary to drive in synchronism irrespective of the load pressure, a value obtained by dividing the third pressure receiving area of the two pressure compensating valves connected to the two actuators by the first pressure receiving area. It is desirable to make them equal. As a result, when the supply flow rate to the left and right traveling motors changes and the rotation speed becomes different, the load pressure of the larger flow rate increases, but the flow rate of the pressure compensating valve decreases as the load pressure increases. And the flow characteristics of the pressure compensating valves of the pair of left and right traveling motors are made equal to each other, so that even when there is an error in the pressure receiving area due to the machining error of the spool of the directional control valve and the machining error of the pressure compensating valve. Such errors increase the load pressure on one of the flow rates. Since the differential pressure between the discharge pressure and the maximum load pressure is constant, as the load pressure increases, the pressure compensating valve with the larger flow rate works in the direction to reduce the flow rate by reducing the differential pressure of the directional control valve. The amount decreases, and the rotation of the traveling motor with the higher flow rate decreases. In the other traveling motor, the load pressure and the differential pressure between the discharge pressure and the maximum load pressure do not change, so that the flow rate does not change and the rotation speed does not change, so that good running straightness can be secured. . On the other hand, when changing direction, the load pressure of the traveling motor with the larger flow rate also rises and works to maintain straightness, but when changing direction, the opening of the directional control valve differs greatly between the left and right, so correction is made. The flow rate is supplied to each traveling motor in accordance with the operation stroke of each directional control valve, and the direction can be changed without maintaining the linearity. In the present invention,
Since there is no need for a special attached valve in addition to the improvement of the original pressure compensating valve, there is provided an excellent effect that the overall size of the valve is not increased, the cost is low, and at the same time, the usability is good. Preferably, if the reduction rate of the flow rate is increased, it is likely to meander excessively when straight ahead,
Further, in order to maintain the straightness at the time of direction change, it becomes difficult to operate, or conversely, if the decrease rate of the flow rate is small, it will not be possible to correct and the straightness cannot be secured, so the second pressure compensation valve It is desirable that the value obtained by dividing the pressure receiving area of No. 3 by the first pressure receiving area is 0.99 to 0.95 (99 to 95%).

[0013] More preferably, as in the invention according to claim 5, in the hydraulic drive device according to claim 3, the pressure compensating valve is provided upstream of each of the corresponding directional control valves, and the pressure compensating valve is provided. The direction in which the downstream pressure of the first oil chamber acts on the first pressure receiving area of the first oil chamber in the direction of closing the valve, and the direction in which the secondary pressure opens the valve on the second pressure receiving area of the second oil chamber. And the load pressure of each of the actuators is applied to the third pressure receiving area of the third oil chamber in the direction of opening the valve, and the second and third pressure receiving areas are made substantially the same, and In the hydraulic drive device in which the first pressure receiving area is made larger than the third pressure receiving area, the load pressure of the first actuator of at least two of the plurality of hydraulic actuators is increased by the other second pressure receiving area. Actu When the load pressure is extremely higher than the load pressure of the ator, the value obtained by dividing the third pressure receiving area of the high load side pressure compensating valve communicating with the high load side actuator by the first pressure receiving area is equal to the low pressure passing through the low load side actuator. The third pressure receiving area of the load-side pressure compensating valve was made smaller than a value obtained by dividing the third pressure receiving area by the first pressure receiving area. With this configuration, when the load pressure of the high-load-side actuator suddenly increases, the flow rate of the high-load-side actuator is reduced, and the reduced flow rate is supplied to the low-load-side actuator. Speed can be prevented. Preferably, a value obtained by dividing a third pressure receiving area of the pressure compensation valve of the low load side actuator by a first pressure receiving area is 1 to 0.98, and a third pressure receiving area of the pressure compensation valve of the high load side actuator. Is divided by the first pressure receiving area to be 0.97 to 0.94.

[0014]

According to a third aspect of the present invention, a variable pump, a plurality of hydraulic actuators driven by discharge oil of the variable pump, and a plurality of hydraulic actuators are provided. A plurality of directional control valves having a flow control function capable of controlling the pressure oil flowing into each of the actuators, and each outlet pressure of each of the directional control valves disposed between each of the directional control valves and each of the actuators. A plurality of pressure compensating valves for compensating for the highest load pressure among the actuators, wherein each of the pressure compensating valves acts on a spring of the pressure compensating valve in a direction to close the pressure compensating valve. The maximum load pressure (Pm) of the force and the actuator is applied, and the pressure (Pd ′) on the upstream side of the pressure compensating valve is applied in a direction to open the pressure compensating valve. Secondary pressure corresponding to the differential pressure between the maximum load pressure and the discharge pressure (Pd) (Pm) (Pc = Pd-
Pm), and a pump flow control valve for communicating the discharge oil of the variable pump with the displacement changing means of the variable pump. The secondary pressure (Pc) is controlled via an oil passage. A hydraulic drive device wherein the pump flow control valve is closed to act to reduce the displacement of the variable pump.

[0015]

According to the structure of the thirteenth aspect of the present invention, the same effects as the effects of the first and second aspects of the present invention can be obtained.

[0016]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS A hydraulic circuit diagram of a hydraulic drive device according to a first embodiment of the present invention will be described with reference to FIG. Discharge oil passage 2 of variable displacement hydraulic pump 2 (hereinafter referred to as pump) driven by prime mover 1 such as an engine
A plurality of pressure compensating valves 41 and 42 (only two of which are shown) are connected in parallel to 3, 3 and check valves 40 are provided in output oil passages 6 of each pressure compensating valve for compensating the pressure of each directional control valve. Directional control valves 8 and 18 (only two of them are shown) having a flow control function capable of controlling the pressure oil flowing into each of the plurality of actuators 10 and 20 (only two are shown) via The output sides of these directional control valves are connected to actuators 10 and 20, respectively, so that the return oil from the respective actuators 10 and 20 is returned to the tank 12 again via the respective directional control valves 8 and 18. . The load pressure extracted from the actuator load pressure extraction port 7 of the direction control valves 8 and 18 via the load pressure extraction line 9 is applied to the actuator 1 by the shuttle valve 13.
The maximum load pressure (hereinafter referred to as the maximum load pressure) (Pm) of 0 and 20 is selected. The pressure compensating valves 41 and 42 apply the pressure (Pz) and the maximum load pressure (Pm) on the downstream side of the pressure compensating valves to the control oil chambers of the respective pressure compensating valves in the closing direction, respectively. , 42, a pump discharge pressure (Pd), which is a pressure upstream of the pressure compensating valve, and an actuator load pressure (PL), which is a downstream pressure of the directional control valve, act on the control oil chamber of each pressure compensating valve. When the discharge amount of the pump 2 falls below a predetermined required amount of the actuators 10 and 20, the pressure compensating valves 41 and 42 provide an anti-surgical device that distributes the discharge amount of the pump 2 to the actuator at an appropriate ratio. Has a curation function.

Further, a pump flow regulating valve 45 for communicating the discharge oil of the variable pump 2 with the displacement changing means 17 of the variable pump is provided, and the maximum load pressure (Pm) is controlled by the oil passage 35.
, The pump flow regulating valve 45 is closed, and the displacement of the variable pump 2 is increased so as to act together with the acting force with the spring 46 of the pump flow regulating valve, and the pump discharge pressure (Pd) is changed to another oil passage 23 ′. Through the pump flow regulating valve 4
5 to act in a direction to reduce the displacement of the variable pump 2 so that the pump discharge pressure Pd and the maximum load pressure Pm
Acting force preset by the spring 46 and
Are balanced, the working force of the pump discharge pressure Pd becomes the working force of the maximum load pressure Pm and the spring 46,
On the other hand, when the pump discharge pressure Pd is larger, the acting force of the pump discharge pressure Pd is reduced so that the displacement of the variable pump 2 is reduced.
When the maximum load pressure Pm and the acting force of the spring 46 are smaller than each other, the displacement of the pump 2 is controlled to be increased, whereby the discharge amount of the variable pump 2 is adjusted according to the (highest) load pressure. It has a load sensing function to control. In the first invention of the present invention, in the hydraulic drive device shown in FIG.
42 is to reduce the flow rate.

With such a configuration, in the first invention of the present invention, the pressure receiving area of the closing direction oil chamber of the pressure compensating valves 41 and 42 is made larger than that of the opening direction oil chamber, and the actuator is actuated in accordance with an increase in the actuator load pressure. Since the flow rate of the pressure compensating valve to be passed is reduced, that is, the directional control valve differential pressure is reduced, the actuator load pressure is attenuated even if there is a sudden change in the load pressure of the actuator, and the hydraulic control system is stabilized, Pressure compensation characteristics of the pressure compensation valve that is not affected by the maximum load pressure of the actuator or the discharge pressure of the variable pump can be obtained.
Both the low-load side and the high-load side have an unprecedented excellent effect of obtaining stable operability without hunting.

A hydraulic circuit diagram of a hydraulic drive device according to a second embodiment of the present invention will be described with reference to FIG. The same components as those in the embodiment of FIG. In the hydraulic circuit diagram of FIG. 2A, the shuttle valve 13 is
Of the maximum load pressure (Pm), and generates a secondary pressure (Pc) corresponding to the differential pressure between the variable pump discharge pressure (Pd) and the maximum load pressure (Pm). The pressure compensating valves 4 and 14 provided in the device 22 act on the outlet pressure (Pz) on the downstream side 6 of the pressure compensating valve in a direction to close the pressure compensating valve, and in a direction to open the pressure compensating valves 4 and 14. The secondary pressure (Pc) of the secondary pressure line 32 taken out from the differential pressure control valve 31 and the load pressure (PL) of the load pressure line 34 taken out of the actuators 10 and 20, which are downstream pressures of the directional control valve, act respectively. I'm trying to make it. Further, a pump flow regulating valve 38 is provided for communicating the discharge oil of the variable pump 2 with the displacement changing means 17 of the variable pump. The acting force of the spring 19 of the pump flow regulating valve is used to close the pump flow regulating valve and displace the pump 2. The secondary pressure Pc is caused to increase by opening the pump flow control valve 38 via the oil passage 33 so as to reduce the displacement of the pump 2, and is set in advance by the secondary pressure Pc and the spring 19. When the force exerted by the secondary pressure Pc is greater than the force exerted by the spring 19 by balancing the applied force, the secondary pressure Pc is adjusted by the spring pressure so as to reduce the displacement of the variable pump 2. When the acting force is smaller than the acting force of the pump 19, the load sensing machine which operates the volume changing means 17 so as to increase the displacement of the pump 2 Having.

Here, the operation of the hydraulic drive device shown in FIG. 2A will be described. The pressure of the upstream side 6 of the directional control valves 8, 18 at the respective pressure compensating valves 4, 14 is reduced by the respective actuators at the downstream side. Since the respective pressure compensating valves 4 and 14 act so as to balance with the sum of the load pressure (PL) and the secondary pressure (Pc), the respective directional control valve differential pressures have the same pressure receiving area. Then, the secondary pressure (Pc) becomes equal to the above-mentioned secondary pressure (Pc) regardless of the load pressure of the actuator. That is, it is equal to the differential pressure between the pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator. further,
The secondary pressure (Pc) is guided to the pump flow control valve 38 through the oil passage 33 and is balanced with the acting force of the spring 19 of the pump flow control valve 38, so that the discharge pressure (Pd) of the pump 2 is: The secondary pressure (Pc) is controlled so as to be equal to the pressure corresponding to the acting force of the spring 19. This means that the variable pump discharge pressure (Pd) is
Control is performed so that c) and the pressure corresponding to the acting force of the spring 19 become equal. Therefore, the directional control valve differential pressure of each directional control valve 8, 18 is also controlled to a pressure corresponding to the acting force of the spring 19. With such a configuration, if the pump discharge amount is insufficient, the difference between the pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator, that is, the secondary pressure (P
In c), since it becomes impossible to secure the differential pressure set by the above-mentioned spring 19, the respective directional control valve differential pressures are also smaller than the set values, but since the directional control valve differential pressures become equal, The flow rate to each of the actuators 10 and 20 is diverted to a flow rate equal to the ratio of the throttle opening of the directional control valves 8 and 18, thus having an anti-saturation function.

With such a configuration, in the second aspect of the present invention, the secondary pressure (Pc) is applied to the direction in which the pump flow regulating valve 38 of the variable pump 2 is closed via the pilot oil passage 33, and the displacement of the variable pump 2 is reduced. Is made to act in the direction of decreasing, as in the prior art or FIG.
The maximum load pressure Pm is applied from the valve device 43 via the long and narrow pilot oil passage 35 in the direction in which the pump flow control valve 45 is closed, and the pump discharge pressure Pd is applied in the direction in which the pump flow control valve 45 is opened. According to the second aspect of the present invention, the viscosity of the hydraulic oil increases at low temperatures,
Even when the pressure loss in the pipe line 23 from the pump 2 to the valve device 22 becomes excessive, the secondary pressure (Pc) of the oil passage 32 becomes the highest with the pump discharge passage 3 in the valve device 22. Since the pressure corresponding to the differential pressure (Pc) from the load pressure (Pm) is generated, regardless of the magnitude of the pressure loss in the pump discharge pipe line 23 from the pump 2 to the valve device 22, the pressure in the valve device 22 is increased. The pump discharge pressure (Pd) of the pump discharge oil passage 3 is controlled to a pressure corresponding to the acting force of the spring 19 with respect to the maximum load pressure of the actuator. Is not greatly reduced, and the speed of the actuator is not reduced.

In the hydraulic circuit of the embodiment shown in FIG. 2A, the pressure compensating valve 4 as disclosed in the first invention of the present invention is also provided.
The pressure receiving areas of the closing direction oil chambers 1 and 42 are made larger than those of the opening direction oil chambers, so that the flow rate of the pressure compensating valve communicating with the actuator can be reduced in accordance with the increase of the actuator load pressure. As a result, even if there is a sudden change in the self-load pressure, the actuator load pressure is attenuated, the hydraulic control system is stabilized, and the pressure compensation characteristic of the pressure compensation valve is not affected by the maximum load pressure of the actuator or the discharge pressure of the variable pump. Irrespective of the single operation or the combined operation, an unprecedented excellent effect of obtaining stable operability without hunting on both the low load side and the high load side was obtained.

Referring to FIG. 3, there is shown a pressure compensating valve 4 using the first invention of the present invention in the hydraulic drive device of FIG. 2,
A cross-sectional structural view of the fourteenth embodiment is shown. Since the cross-sectional structure of the pressure compensating valve 4 is the same as that of the pressure compensating valve 4, the pressure compensating valve 4 is representatively shown here. However, as described later, the pressure compensating valves 4 and 14 may have different pressure receiving areas of the respective pressure chambers. The pressure compensating valve 4 includes a main body 101, a main body hole 128 having two inner diameters of a small-diameter main body hole 111 provided in the main body 101 and a large-diameter main body hole 130,
A spool having a small-diameter portion 132 slidably fitted in the small-diameter main body hole 111 (inner diameter d3) and first and second large-diameter lands 133,134 slidably fitted in the large-diameter main body hole 130 (inner diameter d2). 112, a load pressure port 103, a secondary pressure port 104, an outlet port 105, an inlet port 102 communicating with the pump discharge oil passage, and a tank port 106 of the actuator sequentially provided in the main body 101 along the main body hole 128. Have. A small-diameter portion 132 provided at one end of a spool 112 which fits into the small-diameter main body hole 111 forms a third oil chamber 119 communicating with the load pressure port 103 through a spring 118 so as to be able to abut on the main body end face 127 and a spool. 112 other end 114 tank port 1
An oil chamber 124 leading to 06 is formed.

A second oil chamber 113 communicating with the secondary pressure port 104 is formed in a large-diameter main body hole 130 surrounding a joint between the small-diameter portion 132 of the spool 112 and the first large-diameter land 133. The axial hole 116 (the inner diameter d
In 1), a piston 117 is slidably inserted in an oil-tight manner in a nesting manner, and the other end of the piston 117 is the other end face of the body hole.
26 is located in a tank oil chamber 124 which is accessible to 26 and which communicates with tank port 106. A first oil chamber 121 communicating with the outlet port 105 through a pilot oil passage 123 is formed between the spool 112 and the piston 117 in the axial hole 116, and a first oil chamber 121 is formed.
The first pressure receiving area A1 of the oil chamber 121 is determined by the cross-sectional area of the piston 117, and the second pressure receiving area A2 of the second oil chamber 113 is determined by subtracting the cross-sectional area of the small-diameter hole 111 from the cross-sectional area of the large-diameter main body hole 130. And the third pressure receiving area A3 of the third oil chamber 119 is formed by the small-diameter portion 132 in cross-sectional area, respectively, and the spool 112 has the second large-diameter portion facing the first large-diameter land portion 133. Exit port 105 and entrance port 102 on land 134
And an openable / closable squeezing section 115 for narrowing the gap. An outlet pressure Pz acts on the first oil chamber 121 communicating with the outlet port 105 in a direction to close the throttle portion 115 to the left when the spool 112 is viewed in the drawing, and a second pressure receiving area A2 of the second oil chamber 113 is provided. The secondary pressure Pc causes the throttle 112
And the load pressure PL acts on the third pressure receiving area A3 of the third oil chamber 119 so as to open the throttle portion 115 to the right as viewed from the spool 112 in the drawing.

In the embodiment shown in FIG. 3, the third pressure receiving area A
3 and the second pressure receiving area A2 are made substantially equal,
The outer diameter d3 of the small diameter portion 132 of the spool 112 is slightly smaller than the outer diameter d1 of the piston 117 (d3 <d1), so that the third pressure receiving area A3 is smaller than the first pressure receiving area A1.
In addition, when the spool 112 makes a maximum stroke in the left direction as viewed in FIG. 3, the left end surface of the spool is moved to the third oil chamber 119.
Abuts against the left end surface 127 of the main body of the main body, and closes the narrowed portion 115. Conversely, if you make a maximum stroke to the right,
The right end face 114 of the spool and the right end face of the piston 117 abut on the right end face 126 of the main body hole, and the throttle portion 115 is fully opened. In the middle stroke of the spool 112, the opening degree is proportionally increased by the throttle portion 115 of the spool in proportion to the stroke amount of the spool to the right. The spring 118 is provided with the directional control valves 8 and 18.
When the is not operated, the spool 112 is stroked rightward to open the throttle portion 115,
Its working force is extremely weak. FIG. 3 is intended to conceptually show the principle of operation. Although both ends of the main body hole 128 are not open, the main body hole is actually formed as a stepped through hole (not shown) or a machined hole from the right side. Make up,
It can be configured to be closed by a method such as a screw plug (not shown).

Next, the operation of the embodiment shown in FIG. 3 will be described. Consider the balance of the forces acting on the spool 112 of the pressure compensating valve. First, the force acting on the spool 112 in the right direction in the drawing, that is, the direction in which the throttle portion 115 is opened, is represented by PL load pressure, Pd pump discharge pressure, and Pm maximum load pressure.
Assuming that the secondary pressure is Pc (Pc = Pd−Pm), (A3 · PL) + (A2 · Pc) (1) On the contrary, the force acting in the left direction in FIG. If the outlet pressure of the directional control valve upstream side 6, that is, the outlet port 105 is Pz, then (A1 · Pz) (2) Here, when the pressure compensating valve is controlled, since the forces in both directions are balanced, the expressions (1) and (2) are equal. (A3 · PL) + (A2 · Pc) = (A1 · Pz) (3) The following relationship is established. However, the acting force of the spring 118 is so weak that it is ignored. Suppose here that the outside diameter d of the spool small diameter part
Assuming that the outer diameter d3 of the piston 117 is equal to the outer diameter d1 of the piston 117, A3 =
A1 and the directional control valve differential pressure ΔP = (Pz
−PL) is as follows: ΔP = (Pz−PL) = (A2 / A3) · Pc (4) Therefore, the direction control valve differential pressure ΔP is equal to the secondary pressure Pc
And the outer diameter d2 and outer diameter d3 of the spool 112 and the piston 117
Is determined to be a constant value by the dimension of the outer diameter d1 of, and therefore always becomes a constant value regardless of the individual load pressure PL. In the saturation state, the secondary pressure Pc decreases according to the situation, and the directional control valve differential pressure also decreases. However, as described above, since the differential pressures are equal, the flow rate to the respective actuators 10 and 20 is reduced. Is diverted into a flow rate equal to the ratio of the throttle opening degrees of the directional control valves 8 and 18,
It has an anti-saturation function, and the flow rate to each of the actuators 10 and 20 is not affected by the individual load pressure PL. Note that the third pressure receiving area A3 and the second pressure receiving area A2 may or may not be equal.
If A2 = A3, ΔP = Pc, and A2 ≠ A
In the case of 3, the absolute value of ΔP can be variously changed according to the ratio of A2 and A3 as shown in equation (4). Note that the first pressure receiving area is determined in relation to the third pressure receiving area.

According to the first aspect of the present invention, the outer diameter d 1 of the piston 117 is changed to the outer diameter d 3 of the small diameter portion of the spool 112.
Is slightly larger (d1> d3). Therefore, A
When 3 = k · A1 (where k <1) is substituted into equation (3), k · A1 · PL + A2 · Pc = A1 · Pz (5) For convenience in equation (5), k = ｛1− (1) −k)}, {1− (1−k)} · A1 · PL + A2 · Pc = A1 · Pz, which is transformed into A1 · PL−A1 · (1−k) · PL + A2 · Pc = A1 Pz PL- (1-k) PL + (A2 / A1) Pc = Pz- (1-k) PL + (A2 / A1) Pc = Pz-PL Therefore, the directional control valve differential pressure ΔP becomes ΔP = ( Pz−PL) = (A2 / A1) · Pc− (1−k) · PL (6) Alternatively, if A1 = A3 / k is substituted. ΔP = [(k · A2) / A3] · Pc− (1−k) · PL (7) Here, since the constant k is a value smaller than 1, the second term on the right side of the equations (6) and (7) is a negative value.
According to the equations (6) and (7), the directional control valve differential pressure ΔP is a linear equation of the secondary pressure Pc and the actuator load pressure PL, and according to the increase of the actuator load pressure PL,
Each directional control valve differential pressure ΔP decreases, and the flow rate decreases. That is, a downward pressure compensation characteristic in which the flow rate decreases in accordance with an increase in the actuator load pressure PL is obtained. This is the case where only one directional control valve is operated, or even when two or more directional control valves are simultaneously operated, the condition that the maximum discharge amount of the pump is equal to or more than the required flow rate of all actuators, that is, This is always true if the saturation state has not been reached. That is, in this state, the secondary pressure Pc is maintained at a constant pressure set by the acting force of the spring 19 as described above. On the other hand, since the load pressure PL is the load pressure of each actuator, the right-downward pressure that depends only on the load pressure of each actuator consistently regardless of the load pressure of other actuators, the maximum load pressure of the actuator, or the pump discharge pressure. A compensation characteristic is obtained. On the other hand, in the saturation state in which the pump discharge amount is insufficient, the secondary pressure Pc becomes a pressure Pc 'smaller than the acting force set by the spring 19, and the magnitude of the Pc' depends on the degree of insufficient flow rate and is a constant value. No. However, the same secondary pressure Pc 'acts on all the pressure compensating valves, and all of the discharge amount of the pump is distributed to each actuator at an appropriate ratio. Also, if only one directional control valve is operated to be in a saturation state, all the discharge amount flows to one actuator regardless of the actuator load pressure PL.

Next, a case where two directional control valves are operated to reach a saturation state will be described. For the sake of simplicity, it is assumed that there is no change in the opening of both directional control valves.In this state, only the load pressure of the actuator on one side increases and the load pressure on the other side does not change. I do. The directional control valve differential pressure ΔP on the side where the load pressure has increased decreases as the actuator load pressure PL increases according to the equations (6) and (7). However, the differential pressure in the first term is a small pressure P
Since it is c ', the flow rate itself is small. The directional control valve differential pressure on the other side where the load pressure does not change does not change in the second term of the equations (6) and (7) because the actuator load pressure PL does not change, but the flow rate decreases on the side where the load pressure increases. As a result, there is a margin in the flow rate as a whole, and the secondary pressure Pc 'increases, so the first term increases, and the directional control valve differential pressure increases to increase the flow rate. In other words, the entire pump discharge amount is distributed to the actuator as a whole, and the flow rate on the side where the load pressure does not change increases by an amount corresponding to the decrease in the flow rate on the side where the load pressure increases. Accordingly, in the saturation state, the flow rate on the side where the load pressure does not change increases due to the load pressure on the side where the load pressure increases, even though its own load pressure is constant, but is originally required in the saturation state. This characteristic cannot be a cause of hunting because an overshoot to the target speed does not occur because the actual flow rate is insufficient relative to the flow rate. Rather, the increase in the flow rate on the side where the load pressure does not increase has the advantage that the speed that is closer to the target can be made higher than the insufficient flow rate.

On the other hand, when the other load pressure decreases, the opposite phenomenon occurs. That is, the flow rate increases on the side where the load pressure decreases, and the flow rate decreases on the side where the load pressure is constant. Further, when the fluctuation of the load pressure rises or falls at the same rate, the shunt ratio of each flow rate changes without change. This is also true when three or more directional control valves are operated simultaneously. As described above, according to the present invention, stable controllability without hunting can be obtained even when the pump discharge amount is sufficient or in the saturation state. Further, it goes without saying that the pressure compensation characteristic can be set to an arbitrary value by changing the value of the constant k. In other words, the smaller the value of k, the greater the degree of downward pressure to the right of the pressure compensating valve. This means that the degree of rightward descent can be set according to the load characteristics of each actuator. Moreover, the setting is specifically for piston 1
It is only necessary to change the outer diameter d1 of 17 and there is no need to change the main body 101 itself, which can be easily changed. The value of the constant k is determined in accordance with the actual device. However, in an actuator that is likely to cause hunting, the hunting is easily performed when the reduction rate of the compensation flow rate is small, and the original flow rate is fixed when the reduction rate is large. 0.99>k> 0.95 (0.99 to 95)
%) Is appropriate. As described above, not only the degree of downward slope but also various values of k can be easily obtained with the same main body, so that various pressure compensating valves according to load conditions can be easily obtained.

As shown in FIG. 2 (e), in a hydraulic drive having the circuit shown in FIG. 2 (a), two traveling motors 1 for rotatingly driving a pair of crawlers of a hydraulic traveling vehicle
When at least two actuators 15, 15 'of a plurality of actuators need to be driven synchronously regardless of the load pressure of the actuators, as in the case of 5, 15' drive, the two actuators 15, 15 ' 15 '
It is desirable that the values obtained by dividing the third pressure receiving areas A3 of the two pressure compensating valves 28 and 29 leading to the first pressure receiving area A1 are equal to each other. With such a configuration, when the supply flow rates of the left and right traveling motors 15 and 15 'change and the rotation speeds differ, the load pressure of the supply flow rate which increases is increased. Is reduced, and the flow characteristics of the pressure compensating valves of the pair of right and left traveling motors are made equal to each other.
Even if there is an error in the pressure receiving area due to the processing error of the spool and the processing error of the pressure compensating valve, such error increases the load pressure PL of one of the larger flow rates. Here, since the differential pressure Pc between the discharge pressure Pd and the maximum load pressure Pm is constant, the pressure compensating valve having a larger flow rate decreases the directional control valve differential pressure to decrease the flow rate by increasing the load pressure. Therefore, the inflow amount is reduced, and the rotation of the traveling motor having the larger flow rate is reduced. In the other traveling motor, the load pressure and the differential pressure between the discharge pressure and the maximum load pressure do not change, so that the flow rate does not change and the rotation speed does not change, so that good running straightness can be secured. . On the other hand, when changing direction, the load pressure of the traveling motor with the larger flow rate also rises and works to maintain straightness, but when changing direction, the opening of the directional control valve differs greatly between the left and right, so correction is made. The flow rate is supplied to each traveling motor in accordance with the operation stroke of each directional control valve, and the direction can be changed without maintaining the linearity. Since the present invention does not require a special attached valve in addition to the improvement of the original pressure compensating valve, the present invention has an excellent effect that the dimensions of the entire valve are not increased, the cost is low, and the usability is good.

Preferably, when the rate of decrease in the flow rate is increased, the flow rate is excessively corrected when the vehicle goes straight ahead, and the meandering tends to occur.
In order to maintain the straightness at the time of the direction change, it becomes difficult to operate, or conversely, if the decreasing rate of the flow rate is small, the correction cannot be performed and the straightness cannot be ensured. It is desirable that the value obtained by dividing the pressure receiving area by the first pressure receiving area is 0.99 to 0.95 (99 to 95%).

As shown in FIG. 2F, more preferably, in the hydraulic drive device of the first invention of the present invention, the turning hydraulic pressure of at least two of the plurality of hydraulic actuators 11 and 25 among the plurality of hydraulic actuators When the load pressure of the high load side actuator 25 such as a motor is extremely larger than the load pressure of the low load side actuator 11 such as the other boom hydraulic cylinder, preferably the high load side which communicates with the high load side actuator 25 is used. Pressure compensation valve 3
The third pressure receiving area A3 of the low load side pressure compensating valve 30 communicating with the low load side actuator 11 is obtained by dividing the value obtained by dividing the third pressure receiving area A3 of FIG.
Is smaller than the value divided by the pressure receiving area A1. With this configuration, when the load pressure of the high-load-side actuator 25 suddenly increases, the flow rate of the high-load-side actuator decreases, and the reduced flow rate is reduced by the low-load-side actuator 11.
The low-load side actuator 1
1 can be prevented from decreasing. Preferably, the value obtained by dividing the third pressure receiving area A3 of the pressure compensating valve 30 of the low load side actuator 11 by the first pressure receiving area A1 is 1
It is preferable that the value obtained by dividing the third pressure receiving area A3 of the pressure compensating valve 36 of the high load side actuator 25 by the first pressure receiving area A1 is 0.97 to 0.94.

Further, when the swing load pressure of the high load side actuator 25 is excessive, the opening degree of the pressure compensating valve 36 is narrowed and the flow rate supplied to the swing motor 25 side is reduced. The useless relief flow from the overload relief valve that does not flow is reduced, and at the same time, the rise in the load pressure of the swing motor 25 is suppressed. Therefore, a reduction in the speed of the boom hydraulic cylinder, which is the low-load-side actuator 11, can be prevented by an amount corresponding to a reduction in the useless relief flow rate. Thereafter, the load pressure decreases as the speed of the swing motor 25 increases and the acceleration decreases, so that the opening of the pressure compensating valve 36 gradually increases, and the flow rate gradually increases with the decrease in the swing load pressure. Therefore, a gentle acceleration of the turning motor 25 can be obtained. Further, when the rotation of the swing motor is completed and the vehicle turns at a steady speed, the swing load pressure sharply decreases, and the load pressure of the boom cylinder increases. At this time, the opening of the pressure compensating valve 36 on the side of the swing motor 25 is not fully opened, but is in the process of gradually increasing the opening from the state of the throttle, and is still in the state of being throttled. Therefore, even if the load pressure of the swing motor 25 suddenly decreases, the rate at which the opening of the pressure compensating valve 36 on the swing motor 25 side sharply decreases decreases, and the swing motor 25 decelerates with a shock. Nothing. Conversely, the opening degree of the pressure compensating valve 30 on the boom cylinder 11 side is relatively large since the secondary pressure is secured to some extent from the initial stage of turning of the turning motor 25. Even if the rotation of the rotation motor ends and the vehicle turns at a steady speed,
The opening does not suddenly increase unlike the conventional example, and acceleration with a shock does not occur.

Referring to FIG. 4, a description will be given of a pressure compensating valve 4 ′ of an embodiment different from FIG. 3 in which the first invention of the present invention is used in the hydraulic drive device of FIG. The same parts as those in the embodiment of FIG. 3 are denoted by the same reference numerals, and a part of the description is omitted. The pressure compensating valve 4 'in FIG. 4 is different from the embodiment in FIG. 3 in the configuration of the pressure receiving areas A3 and A2 acting in the opening direction of the pressure compensating valve. That is, in FIG. 4, the main body 201 has only a large-diameter hole (inner diameter d2) in the main body hole 228, and a spool 212 having first, second, and third large-diameter lands 209, 210, and 211 is provided in the main body hole 228. It is inserted. And the small diameter hole of FIG.
An auxiliary piston 217 having an outer diameter of d3 instead of 111 (inner diameter d3) is slidably inserted in an auxiliary axial direction hole 202 provided in an outer end 214 of a spool 212 in a nested manner.
A secondary pressure port 204, an actuator load pressure port 203, an outlet port 105, and an inlet port 1 communicating with the pump discharge oil passage are sequentially provided in the main body 201 along the main body hole 228.
02, and a tank port 106. Auxiliary piston 2
The outer end of 17 is adapted to abut against the end face 227 of the body bore 228 to form a second oil chamber 213 communicating with the secondary pressure port 204. Spool 212 and auxiliary piston in auxiliary axial hole 202
A third oil chamber 220 is formed between the second oil chamber 217 and the load pressure port 203 through a spring 218 and an auxiliary pilot oil passage 223. The first pressure receiving area A1 of the first oil chamber 121 depends on the cross-sectional area of the piston 117.
The second pressure receiving area A2 of 213 is obtained by subtracting the cross-sectional area of the auxiliary piston 217 from the cross-sectional area of the main body hole 228, and the third pressure-receiving area A3 of the third oil chamber 220 is obtained by the auxiliary piston 217.
Are formed respectively according to the cross-sectional area of.

According to this configuration, each of the diameters d1, d
If the relationship between 2 and d3 is the same as in the embodiment of FIG.
The secondary pressure Pc, that is, the differential pressure between the pump discharge pressure and the maximum load pressure of the actuator becomes the acting force of the spring 19 of the pump flow control valve 38.
Since the load pressure PL is sufficiently large, the auxiliary piston is pressed against the left end face of the main body hole, and the same operation as the embodiment shown in FIG. 3 is obtained.

Note that in the embodiment of FIG. 4, the secondary pressure P
When the load pressure PL becomes smaller than the pressure c, the auxiliary piston 217 moves away from the left end face 227 of the main body hole and the spool 2
The secondary pressure Pc acts on the pressure receiving area A3 where the load pressure PL acts on the pressure receiving area A3. In this case, the load pressure is assumed to be the same as Pc and is controlled, and the differential pressure of the directional control valve is slightly increased, so that the flow rate is slightly increased. However, in contrast to the embodiment of FIG. 3 having a stepped hole, the embodiment of FIG. 4 has an advantage that the hole processing does not need to be stepped and processing is easy, and the secondary pressure Pc is Since the pressure is inherently small, there are few problems in practical use caused by the secondary pressure Pc acting on the outer end face of the auxiliary piston, and it is more suitable for a device in which the load pressure itself is always a certain pressure or higher. Become.

As described above, the pressure compensating valves 4 and 4 ′ shown in FIGS. 3 and 4 have been described as being used in FIG. 2 (a).
4 'is applicable other than the circuit example of FIG. That is, as described above, the load pressure PL and the secondary pressure Pc act in the opening direction, and the upstream pressure Pz (downstream of the pressure compensating valve) Pz acts in the closing direction to control. What is necessary is just a pressure compensating valve. For example, it can be used for the circuit shown in FIG.

Therefore, instead of applying the secondary pressure Pc to the pump control circuit shown in FIG. 2A via the pilot oil passage 33, the variable pump 2 shown in FIG. And a pump flow control valve 45 for communicating the discharge oil of the variable pump with the displacement changing means 17 of the variable pump, and the maximum load pressure (Pm) is controlled via the oil passage 35 to the pump flow control valve 4.
5 is actuated in a direction to increase the displacement of the variable pump 2 to reduce the pump discharge pressure (Pd) to another oil passage 23 '.
, The pump flow control valve 45 is opened to act in a direction to decrease the displacement of the variable pump 2, and the acting force of the pump discharge pressure Pd is reduced by the maximum load pressure Pm and the spring 46.
And the action force set in advance may be balanced.

As shown in FIG. 2 (c), in FIG. 2 (a), the variable pump discharge pressure (P
d) and the differential pressure between the maximum load pressure (Pm) of the actuator from the maximum load pressure extraction line 16 is detected by the differential pressure detector 60, and the output of the differential pressure detector 60 is input to the control device 61 and the control signal is output. Generating and outputting a control signal 62
Pressure (Pc) output by the solenoid operated proportional valve 63 operating at
Thus, an anti-saturation function may be secured. 64 is a pilot pump.

FIG. 2D is a hydraulic circuit diagram showing an embodiment of a hydraulic drive device different from that of FIG. 2A. The point that a pressure compensating valve is disposed downstream of a directional control valve is disclosed in 19409
This is the same as in the above publication. 2A are denoted by the same reference numerals, and a part of the description is omitted. As shown in FIG. 2D, reference numerals 50 and 51 denote hydraulic actuators. The check valve 40 checks the pressure oil in the discharge oil passage 3, the directional control valves 53 and 54 having a flow control function, and the pressure compensation valves 44 and 48. The return oil of the actuators 50 and 51 is returned from the direction control valves 53 and 54 to the tank T via the tank oil passage 12. The highest load pressure among the load pressures of the actuator is selected by the shuttle valve 13 and both pressure compensating valves 44,
The valve 48 acts in the closing direction together with the springs 44a and 48a, and the downstream pressure Pd 'of the direction control valves 53 and 54 acts in the direction in which the pressure compensating valve opens. The differential pressure across the directional control valves 53, 54 is made to match the differential pressure between the variable pump discharge pressure Pd and the maximum load pressure Pm as in FIG. I have. A differential pressure control valve 31 for generating a pressure Pc corresponding to a differential pressure between the pump discharge pressure Pd of the discharge oil passage 3 in the valve device 24 and the maximum load pressure Pm of the outlet oil passage 16 selected by the shuttle valve 13 is shown. 2 (a), the secondary pressure Pc generated by the differential pressure control valve 31 is supplied through the secondary pressure line 32 and the pilot oil passage 33 to the pump flow control valve 3 of the pump device 21.
8, the discharge amount of the variable pump 2 is opened to operate the displacement changing means 17 so as to reduce the discharge amount of the variable pump 2. Is controlled so as to increase the discharge amount.

The operation of the hydraulic drive device shown in FIG. 2D will be described.
The pump discharge pressure Pd of the discharge oil passage 3 is increased by the maximum load pressure Pm and the acting force of the spring 19 because the pressure of the discharge conduit 3 is higher than the pressure of the discharge conduit 3 by an amount corresponding to the pressure loss in the discharge conduit 23. Only depending on the temperature of the hydraulic oil. That is, the pump flow control valve 38
The balance formula of the internal force is as follows: Pc = acting force of the spring 19, and the secondary pressure Pc is obtained from the differential pressure control valve 31 as Pc = P
Since d−Pm, Pd−Pm = acting force of the spring 19 holds, and the pump discharge pressure becomes Pd = Pm + acting force of the spring 19. Further, from the balance of the forces in the pressure compensating valves 44 and 48, Pd '= Pm + the acting force of the spring 44a. Therefore, the differential pressure between the direction control valves 53 and 54 = P
d−Pd ′ = acting force of spring 19−spring 44
a, the differential pressure across the directional control valve is changed by the pump flow control valve 3
Since the pressure is determined only by the acting force of the spring 19 of FIG. 8 and the acting force of the springs 44a and 48a of the pressure compensating valves 44 and 48 and does not depend on the load pressure of the actuators 50 and 51, the operating oil similar to FIG. The hydraulic drive device is not affected by the temperature of the motor. On the other hand, in the case of the saturation state, the relationship of the above expression is not established because the pump discharge amount is insufficient. However, the direction control valves 53, 5
4 is the sum of the maximum load pressure Pm and the acting force of the springs of the springs 44a and 48a of the pressure compensating valves. If the respective springs are the same, the downstream pressures of all the directional control valves are equal. Are the same. On the other hand, since the upstream side of all the directional control valves communicates in parallel with the discharge conduit 3,
Its pressure is the same as Pd. Accordingly, the differential pressures before and after all the directional control valves are the same, and the variable pump discharge amount is divided at a ratio proportional to the ratio of the opening degrees of the respective directional control valves, as in the case of FIG. It will have a saturation function.

Referring to FIG. 5, a pressure compensating valve 41 used in FIG. 1, which is an embodiment of the first invention of the present invention, will be described. The body 301 of the pressure compensating valve 41 is the first body 30
The first body 301a is divided into a first body 301a and a second body 301b.
Is provided with a small diameter hole 321 and a medium diameter hole 322 following the small diameter hole.
A second spool 312 that fits into the 22 is arranged, and a second body 301b is provided with a large-diameter hole 323 following the medium-diameter hole 322, a small-diameter hole 321 following the large-diameter hole, and an auxiliary small-diameter hole 325 having the same diameter as the small-diameter hole 321. The third spool 310 is fitted into the large-diameter hole 323 and the auxiliary small-diameter hole 325.
Are the first and second large-diameter lands 31 fitted into the large-diameter holes 323.
3,314 and an auxiliary small-diameter portion 315 fitted into the auxiliary small-diameter hole 325, and a spring 350 for pressing each spool is interposed between the first spool 311 and the small-diameter hole 321 end face 320. Further, along the body 301, an auxiliary inlet port 341 communicating with the pump discharge oil passage 3 and communicating with the small-diameter hole 321; an actuator load pressure port 342 communicating with the actuator load pressure line 34 and communicating with the medium-diameter hole 322; 2 spools 3
A tank port 343 communicating with a large-diameter hole 323 surrounding a contact portion between the second spool 12 and the third spool 310; a first and a second large-diameter land 31;
An outlet port 344 communicating with the large-diameter hole 323 between 3,314, an inlet port 345 communicating with the pump discharge oil passage 3 and opening to be throttled by an openable / closable throttle 316 provided in the second large-diameter land 314; Among them, a maximum load pressure port 346 communicating with the maximum load pressure take-out line 16 and communicating with a large diameter hole 323 in a continuous portion of the second large diameter land 314 and the auxiliary small diameter portion 315 is sequentially provided. An oil chamber 334 is provided between the auxiliary small-diameter portion 315 and the auxiliary small-diameter hole end surface 330 and communicates with the outlet port 344 via the pilot oil passage 351. The first body 301a and the second body 301b are assembled together by a suitable method such as bolting (not shown) to form the body 301.
At this time, even if the medium-diameter hole 322 on the first body 301a side and the large-diameter hole 323 on the second body 301b are misaligned, the second spool 312 and the third spool 310 are merely abutted as separate parts. There are no operational problems.

Then, the pressure compensating valve 41 applies the outlet pressure (Pz) of the outlet port 344 in the closing direction to the pilot oil passage 351.
The maximum load pressure (Pm) of the maximum load pressure port 346 is applied to the end surface 340 (pressure receiving area B1) of the auxiliary small diameter portion 315 of the oil chamber 334 from the sectional area of the second large diameter land 314 via the Pressure receiving area B2 of oil chamber 336 with the cross-sectional area subtracted
Respectively. Further, in the direction in which the pressure compensating valve 41 is opened, the pump discharge pressure (P
d) to the pressure receiving area B1 of the first spool 311 of the oil chamber 331,
Actuator load pressure (PL) at load pressure port 342
On the pressure receiving area B3 of the oil chamber 332 obtained by subtracting the sectional area B1 of the first spool 311 from the sectional area of the second spool 312. The cross-sectional area of the oil chamber 333, which is obtained by subtracting the cross-sectional area of the second spool 312 from the cross-sectional area of the first large-diameter land 313, communicates with the tank through the tank port 343. The working force does not work. The sectional area B2 and the sectional area B1 of the first spool 311 are made substantially the same (B1 = B2). In addition, the sectional area B3 is made smaller than the sectional area B1 (= B2) of the first spool (B1> B3). ), The load pressure of each actuator (P
The flow rate of the pressure compensating valve 41 leading to the actuator is reduced in accordance with the increase of L).

In operation, the force of the spring 350 pushing each spool is as weak as the spring 118 of FIG. 3, so ignoring this force, when each spool of the pressure compensating valve 41 is in equilibrium, The balance of the force applied to each spool is: Pz · B1 + Pm · B2 = Pd · B1 + PL · B3 (8) Since B1 = B2, Pz · B1 + Pm · B1 = Pd · B1 + PL · B3 (9) Pz + Pm = Pd + PL · (B3 / B1) B1> B3, and (B3 / B1) = k, Pz + Pm = Pd + PL · k (10) where k <1 k = 1− (1-k) In other words, Pz + Pm = Pd + PL. [1- (1-k)] Pz + Pm = Pd + PL-PL. (1-k) (11) Since ΔP = Pz-PL, from the equation (11), Pz PL = −Pm + Pd−PL · (1-k) ΔP = Pz−PL = (Pd−Pm) −PL · (1−k) (12) where Pc in FIG. 3 is Pc = Pd− Since the pressure is Pm, the pressure compensating valve 41 in FIG. 5 operates in the same manner as the pressure compensating valve 4 in FIG.

For the same reason as described for the pressure compensating valve 4 in FIG. 3, the value obtained by dividing the third pressure receiving area B 3 of the two pressure compensating valves 41 and 42 by the first pressure receiving area B 1 is 0.99. ~ 0.
95 (99-95%) is desirable. Also in the pressure compensating valves 41 and 42 of the hydraulic drive device according to the first invention of the present invention, at least two of the plurality of actuators are driven like two traveling motor drives for rotating a pair of crawlers of a hydraulic traveling vehicle. When the two actuators need to be driven synchronously irrespective of the load pressure of the actuators, the third pressure receiving area B3 of the two pressure compensating valves 41 and 42 communicating with the two actuators is changed to the first pressure receiving area B3. It is desirable to make the values divided by the area B1 equal. Preferably, if the decrease rate of the flow rate is increased, it is likely to be overcorrected when straight ahead and meander easily, and it is difficult to operate while maintaining straightness at the time of turning, and conversely if the decrease rate of the flow rate is small. Since the correction cannot be performed and the straightness cannot be secured, the two pressure compensating valves 4
It is desirable that the value obtained by dividing the third pressure receiving area B3 of 1, 42 by the first pressure receiving area B1 is 0.99 to 0.95 (99 to 95%). For the same reason as described for the pressure compensating valve 4 in FIG. 3, in the pressure compensating valves 41 and 42 used in the hydraulic drive device of the first invention of the present invention, at least two of the actuators 10 and 20 are used. The third pressure receiving area of the high load side pressure compensating valve 42 which communicates with the high load side actuator 20 when the load pressure of the high load side actuator 20 of the actuators is extremely larger than the load pressure of the other low load side actuator 10. A value obtained by dividing B3 by the first pressure receiving area B1 is made smaller than a value obtained by dividing the third pressure receiving area B3 of the low load side pressure compensating valve 41 communicating with the low load side actuator 10 by the first pressure receiving area B1. It is desirable. The value obtained by dividing the third pressure receiving area of the pressure compensating valve of the low load side actuator by the first pressure receiving area is 1 to
0.98, the value obtained by dividing the third pressure receiving area of the pressure compensating valve of the high load side actuator by the first pressure receiving area is 0.97.
It is desirably 0.94.

In each of the embodiments described above, the hydraulic circuit that drives two hydraulic actuators has been described. For example, in a hydraulic shovel, two traveling motors that rotationally drive a pair of crawlers of a hydraulic traveling vehicle, a turning hydraulic motor, And like each hydraulic cylinder of boom, arm and bucket,
At least six actuators are operated. Therefore, in each of the embodiments described above, only two actuators representing these are shown. In the present invention, a plurality of hydraulic actuators are referred to as a plurality of traveling motors, a turning hydraulic motor, a hydraulic cylinder, and the like, respectively. And a plurality of pressure compensating valves and directional control valves, each communicating with a hydraulic actuator.

[Brief description of the drawings]

FIG. 1 is a hydraulic circuit diagram of a hydraulic drive device according to a first embodiment of the present invention.

FIG. 2A is a hydraulic circuit diagram of a hydraulic drive device according to an embodiment including the second invention and the first invention of the present invention,
(B) is a partial hydraulic circuit diagram showing a pump discharge amount control circuit according to an embodiment different from (a), and (c) is a partial hydraulic circuit diagram showing a secondary pressure generating device different from (a) in FIG.
The other parts of (a) are the same, and FIG.
Fig. 2 (e) is a hydraulic circuit diagram in which a pressure compensating valve is arranged downstream of a directional control valve according to a further different embodiment from Fig. 2 (a). FIG. 2 is a partial hydraulic circuit diagram of a hydraulic drive device that drives a traveling motor in synchronization.
FIG. 2F is a partial hydraulic circuit diagram of a hydraulic drive device for driving two actuators having greatly different loads, which is still another embodiment different from FIG. 2A.

FIG. 3 is a sectional structural view (conceptual diagram) of an embodiment of a pressure compensating valve used in the hydraulic drive device of FIG. 2 (a).

FIG. 4 is a sectional structural view (conceptual diagram) of a pressure compensating valve of an embodiment different from FIG. 3 used in the hydraulic drive device of FIG. 2 (a).
Is shown.

FIG. 5 is a sectional structural view (conceptual diagram) of a pressure compensating valve used in the hydraulic drive device of FIG. 1;

[Explanation of symbols]

1 motor 2 variable pump 4, 4 ', 14, 28, 29, 30, 36, 44, 48 pressure compensating valve 8, 18 directional control valve 10, 11, 15, 15', 20, 25, 50, 51,
Hydraulic actuator 13 Shuttle valve 16 Maximum load pressure line 17 Pump volume changing means 19, 46 Spring 31 Differential pressure control valve 33, 35 Pilot oil passage 38, 45 Pump flow control valve

────────────────────────────────────────────────── ───

[Procedure amendment]

[Submission date] May 20, 1997

[Procedure amendment 1]

[Document name to be amended] Drawing

[Correction target item name] All figures

[Correction method] Change

[Correction contents]

FIG.

FIG. 2

FIG. 3

FIG. 4

FIG. 5

FIG. 6

FIG. 7

[Procedure amendment 2]

[Document name to be amended] Statement

[Correction target item name] 0022

[Correction method] Change

[Correction contents]

In the hydraulic circuit of the embodiment shown in FIG. 2A, as disclosed in the first invention of the present invention , the pressure compensating valve
The pressure receiving areas of the closing direction oil chambers 4 and 14 can be made larger than those of the opening direction oil chambers, so that the flow rate of the pressure compensating valve to the actuator can be reduced in accordance with the increase of the actuator load pressure. As a result, even if there is a sudden change in the self-load pressure, the actuator load pressure is attenuated, the hydraulic control system is stabilized, and the pressure compensation characteristic of the pressure compensation valve is not affected by the maximum load pressure of the actuator or the discharge pressure of the variable pump. Irrespective of the single operation or the combined operation, an unprecedented excellent effect of obtaining stable operability without hunting on both the low load side and the high load side was obtained.

[Procedure amendment 3]

[Document name to be amended] Statement

[Correction target item name] 0023

[Correction method] Change

[Correction contents]

Referring to FIG. 3, there is shown a sectional structure of an embodiment of a pressure compensating valve 4 , 14 , 28, 29 30, 36 in which the first invention of the present invention is used in the hydraulic drive device of FIGS . 2 and 7. The figure is shown. Since the sectional structure of the pressure compensating valve is the same as that of the pressure compensating valves 4, 14 , 28 , 29, 30, and 36 , the pressure compensating valve 4 is representatively shown here. However, as described later, the pressure compensating valves 4 and 14 may have different pressure receiving areas of the respective pressure chambers. The pressure compensating valve 4 includes a main body 101 and a small-diameter main body hole 1 provided in the main body 101.
11, a main body hole 128 having two inner diameters of a large-diameter main body hole 130, a small-diameter portion 132 and a large-diameter main body hole 130 slidably fitted in the small-diameter main body hole 111 (inner diameter d3).
A spool 112 having first and second large-diameter lands 133 and 134 slidably fitted to the (inner diameter d2); a load pressure port 103 of an actuator sequentially provided on the main body 101 along the main body hole 128; Secondary pressure port 10
4, an outlet port 105, an inlet port 102 communicating with the pump discharge oil passage, and a tank port 106. A small-diameter portion 132 provided at one end of the spool 112 fitted into the small-diameter main body hole 111 is connected to the load pressure port 103 via a spring 118 so as to be able to abut on the main body hole end surface 127.
Of the spool 112 and the other end 11 of the spool 112
4 forms an oil chamber 124 communicating with the tank port 106.

[Procedure amendment 4]

[Document name to be amended] Statement

[Correction target item name] 0030

[Correction method] Change

[Correction contents]

As shown in FIG. 7 (a), in the hydraulic drive device having a circuit shown in FIG. 2 (a), 2 pieces of the travel motor for rotating the pair of crawler hydraulic traveling vehicle 1
When at least two actuators 15, 15 'of a plurality of actuators need to be driven synchronously regardless of the load pressure of the actuators, as in the case of 5, 15' drive, the two actuators 15, 15 ' 15 '
It is desirable that the values obtained by dividing the third pressure receiving areas A3 of the two pressure compensating valves 28 and 29 leading to the first pressure receiving area A1 are equal to each other. With such a configuration, when the supply flow rates of the left and right traveling motors 15 and 15 'change and the rotation speeds differ, the load pressure of the supply flow rate which increases is increased. Is reduced, and the flow characteristics of the pressure compensating valves of the pair of right and left traveling motors are made equal to each other.
Even if there is an error in the pressure receiving area due to the processing error of the spool and the processing error of the pressure compensating valve, such error increases the load pressure PL of one of the larger flow rates. Here, since the differential pressure Pc between the discharge pressure Pd and the maximum load pressure Pm is constant, the pressure compensating valve having a larger flow rate decreases the directional control valve differential pressure to decrease the flow rate by increasing the load pressure. Therefore, the inflow amount is reduced, and the rotation of the traveling motor having the larger flow rate is reduced. In the other traveling motor, the load pressure and the differential pressure between the discharge pressure and the maximum load pressure do not change, so that the flow rate does not change and the rotation speed does not change, so that good running straightness can be secured. . On the other hand, when changing direction, the load pressure of the traveling motor with the larger flow rate also rises and works to maintain straightness, but when changing direction, the opening of the directional control valve differs greatly between the left and right, so correction is made. The flow rate is supplied to each traveling motor in accordance with the operation stroke of each directional control valve, and the direction can be changed without maintaining the linearity. Since the present invention does not require a special attached valve in addition to the improvement of the original pressure compensating valve, the present invention has an excellent effect that the dimensions of the entire valve are not increased, the cost is low, and the usability is good.

[Procedure amendment 5]

[Document name to be amended] Statement

[Correction target item name] 0032

[Correction method] Change

[Correction contents]

As shown in FIG . 7 (b) , more preferably, in the hydraulic drive device according to the first invention of the present invention, the turning hydraulic pressure of at least two of the plurality of hydraulic actuators 11, 25 is provided. When the load pressure of the high load side actuator 25 such as a motor is extremely larger than the load pressure of the low load side actuator 11 such as the other boom hydraulic cylinder, preferably the high load side which communicates with the high load side actuator 25 is used. Pressure compensation valve 3
The third pressure receiving area A3 of the low load side pressure compensating valve 30 communicating with the low load side actuator 11 is obtained by dividing the value obtained by dividing the third pressure receiving area A3 of FIG.
Is smaller than the value divided by the pressure receiving area A1. With this configuration, when the load pressure of the high-load-side actuator 25 suddenly increases, the flow rate of the high-load-side actuator decreases, and the reduced flow rate is reduced by the low-load-side actuator 11.
The low-load side actuator 1
1 can be prevented from decreasing. Preferably, the value obtained by dividing the third pressure receiving area A3 of the pressure compensating valve 30 of the low load side actuator 11 by the first pressure receiving area A1 is 1
It is preferable that the value obtained by dividing the third pressure receiving area A3 of the pressure compensating valve 36 of the high load side actuator 25 by the first pressure receiving area A1 is 0.97 to 0.94.

[Procedure amendment 6]

[Document name to be amended] Statement

[Correction target item name] 0037

[Correction method] Change

[Correction contents]

Although the pressure compensating valves 4 and 4 ′ shown in FIGS. 3 and 4 have been described as being used in FIGS . 2 and 7 , the pressure compensating valves 4, 4 4 'is applicable other than the circuit examples of FIGS. That is, as described above, the load pressure PL
And the secondary pressure Pc acts, and the upstream pressure of the directional control valve (downstream of the pressure compensating valve) Pz acts in the closing direction to control the pressure compensating valve. For example, FIG.
It can be used for the circuits shown in

[Procedure amendment 7]

[Document name to be amended] Statement

[Correction target item name] 0040

[Correction method] Change

[Correction contents]

FIG . 6 is a hydraulic circuit diagram showing an embodiment of a hydraulic drive device different from that of FIG. 2A. The point that a pressure compensating valve is disposed downstream of a directional control valve is disclosed in Japanese Patent Laid-Open No. 19409/1992. Is the same as 2A are denoted by the same reference numerals, and a part of the description is omitted. As shown in FIG. 6 , reference numerals 50 and 51 denote hydraulic actuators for guiding the pressure oil in the discharge oil passage 3 through a check valve 40, directional control valves 53 and 54 having a flow control function, and pressure compensating valves 44 and 48. The return oil of the actuators 50 and 51 is supplied to the direction control valve 5.
3 and 54 are returned to the tank T via the tank oil passage 12. The highest load pressure among the load pressures of the actuators is selected by the shuttle valve 13, and both pressure compensating valves 44, 48 are actuated together with the springs 44a, 48a in the closing direction, and the downstream pressure Pd 'of the directional control valves 53, 54 is reduced. The pressure compensating valves act in the opening direction. The differential pressure across the directional control valves 53, 54 is made to match the differential pressure between the variable pump discharge pressure Pd and the maximum load pressure Pm as in FIG. I have. A differential pressure control valve 31 for generating a pressure Pc corresponding to a differential pressure between the pump discharge pressure Pd of the discharge oil passage 3 in the valve device 24 and the maximum load pressure Pm of the outlet oil passage 16 selected by the shuttle valve 13 is shown. 2 (a), the secondary pressure Pc generated in the differential pressure control valve 31 is supplied to the secondary pressure line 32, the pilot oil passage 3
3, the discharge amount of the variable pump 2 is opened to the pump flow regulating valve 38 of the pump device 21 and the displacement amount changing means 17 is displaced.
Is operated to decrease the discharge amount of the variable pump 2, and the spring 19 of the pump flow control valve 38 is controlled to close the pump flow control valve and increase the discharge amount of the variable pump 2.

[Procedure amendment 8]

[Document name to be amended] Statement

[Correction target item name] 0041

[Correction method] Change

[Correction contents]

The operation of the hydraulic drive device shown in FIG . 6 will be described. The pressure at the discharge port of the variable pump 2 is equal to the pressure of the discharge conduit 3 in the valve device 24 by the amount corresponding to the pressure loss in the discharge conduit 23. Since the pressure becomes high, the discharge oil passage 3
Is determined only by the maximum load pressure Pm and the acting force of the spring 19, and does not depend on the temperature of the hydraulic oil. That is, the balance formula of the force in the pump flow control valve 38 is as follows: Pc = acting force of the spring 19, and the secondary pressure Pc is obtained from the differential pressure control valve 31 by Pc
= Pd-Pm, Pd-Pm = acting force of the spring 19 holds, and the pump discharge pressure becomes Pd = P
m + spring 19 acting force. Further, from the balance of the forces in the pressure compensating valves 44 and 48, Pd '= Pm + the acting force of the spring 44a. Therefore, the direction control valve 5
The differential pressure before and after 3, 54 = Pd-Pd '= the acting force of the spring 19 The acting pressure of the car spring 44a is applied, and the differential pressure before and after the directional control valve is changed by the spring 19 of the pump flow regulating valve 38.
And the springs 44a of the pressure compensating valves 44 and 48,
48a is determined only by the acting force of the actuator 5
Since it does not depend on the load pressures of 0 and 51, it is possible to provide a hydraulic drive device that is not affected by the temperature of the hydraulic oil, as in FIG. On the other hand, in the case of the saturation state, the relationship of the above expression is not established because the pump discharge amount is insufficient. However, the pressure P downstream of the directional control valves 53, 54
d 'is the maximum load pressure Pm and the pressure compensating valve spring 4
The sum of the acting forces of the springs 4a and 48a is the same. When each spring is the same, the pressures downstream of all the directional control valves are the same. On the other hand, since the upstream side of all the directional control valves communicates in parallel with the discharge conduit 3, the pressure is the same as Pd. Accordingly, the differential pressures before and after all the directional control valves are the same, and the variable pump discharge amount is divided at a ratio proportional to the ratio of the opening degrees of the respective directional control valves, as in the case of FIG. It will have a saturation function.

[Procedure amendment 9]

[Document name to be amended] Statement

[Correction target item name] Brief description of drawings

[Correction method] Change

[Correction contents]

[Brief description of the drawings]

FIG. 1 is a hydraulic circuit diagram of a hydraulic drive device according to a first embodiment of the present invention.

FIG. 2A is a hydraulic circuit diagram of a hydraulic drive device according to an embodiment including the second invention and the first invention of the present invention,
(B) is a partial hydraulic circuit diagram showing a pump discharge amount control circuit according to an embodiment different from (a), and (c) is a partial hydraulic circuit diagram showing a secondary pressure generating device different from (a) in FIG.
The other parts of (a) remain unchanged .

FIG. 3 is a sectional structural view (conceptual diagram) of an embodiment of a pressure compensating valve used in the hydraulic drive device of FIGS . 2 and 7 ;

FIG. 4 is a sectional structural view (conceptual diagram) of a pressure compensating valve of an embodiment different from FIG. 3 used in the hydraulic drive device of FIGS . 2 and 7 ;

FIG. 5 is a sectional structural view (conceptual diagram) of a pressure compensating valve used in the hydraulic drive device of FIG. 1;

FIG. 6 is an embodiment further different from FIG. 2 (a).
Hydraulic circuit diagram with pressure compensating valve located downstream of directional control valve
Is shown.

FIG. 7 (a) is another embodiment different from FIG. 2 (a).
Hydraulic drive unit that drives two traveling motors synchronously
FIG. 2 (a) is a still another implementation of FIG. 2 (a).
Two actuators with very different loads in form
The partial hydraulic circuit diagram of the hydraulic drive that drives the motor
Shown respectively.

[Description of Signs] 1 Prime mover 2 Variable pump 4, 4 ', 14, 28, 29, 30, 36, 44, 48
Pressure compensating valve 8, 18 Directional control valve 10, 11, 15, 1
5 ', 20, 25, 50, 51, hydraulic actuator 13 shuttle valve 16 maximum load pressure line 17 pump volume changing means 19, 46 spring 31 differential pressure control valve 33, 35 pilot oil passage 38, 45 pump flow control valve

 ──────────────────────────────────────────────────続 き Continued on the front page (31) Priority claim number Japanese Patent Application No. 8-210468 (32) Priority date Hei 8 (July 23, 1996) (33) Priority claim country Japan (JP)

Claims (18)

[Claims] 1. A variable pump, a plurality of hydraulic actuators driven by discharge oil of the variable pump, and a plurality of directions having a flow rate adjusting function capable of controlling pressure oil flowing into each of the plurality of actuators. A plurality of pressure compensating valves for compensating respective pressures of the control valve and each of the directional control valves, wherein each of the pressure compensating valves has a pressure (Pz) downstream of the pressure compensating valve in a closing direction and a pressure compensation valve of the plurality of actuators. The maximum load pressure (Pm) is applied to each of them, and the pump discharge pressure (Pd) which is the pressure on the upstream side of the pressure compensating valve in the opening direction of the pressure compensating valve and the actuator load pressure (Pd) which is the downstream pressure of the directional control valve. PL) to actuate the pressure compensation, and to communicate the discharge oil of the variable pump with the displacement changing means of the variable pump. The pump pressure (Pd) is applied by applying a load pressure (Pm) in the direction of closing the pump flow control valve via an oil passage to increase the displacement of the variable pump together with the acting force of the spring of the pump flow control valve. The hydraulic compensator is configured to open the pump flow control valve via another oil path to reduce the displacement of the variable pump and to act in the direction, wherein the pressure compensating valve increases the load pressure of each actuator. A hydraulic drive device characterized in that the flow rate of the pressure compensating valve connected to the actuator is reduced in accordance with the pressure. 2. A variable pump, a plurality of hydraulic actuators driven by oil discharged from the variable pump, and a plurality of directions having a flow rate adjusting function capable of controlling pressure oil flowing into each of the plurality of actuators. A plurality of pressure compensating valves for compensating the pressures of the control valve and the directional control valves; and a secondary pressure (Pc) corresponding to a differential pressure between the variable pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator. = Pd-Pm), and a pump flow regulating valve for communicating the discharge oil of the variable pump with the displacement changing means of the variable pump. Each of the pressure compensating valves is a pressure compensating valve. The pressure (Pz) downstream of the pressure compensating valve acts in the direction in which the valve is closed, and the secondary pressure (Pc) output from the differential pressure control valve and the downstream pressure in the direction control valve in the direction in which the pressure compensating valve is opened. In Actuator load pressure (P
L) acts in the direction in which the pressure compensating valve is opened to perform the pressure compensation, wherein the acting force of the spring of the pump flow regulating valve is closed by closing the pump flow regulating valve and displacing the variable pump. A hydraulic system characterized in that the secondary pressure (Pc) is made to act in a direction to increase the volume, and the secondary pressure (Pc) is made to act to reduce the displacement of the variable pump by opening the pump flow control valve via an oil passage. Drive. 3. The hydraulic drive according to claim 2, wherein
The hydraulic drive device according to claim 1, wherein the pressure compensating valve decreases a flow rate of the pressure compensating valve connected to the actuator in accordance with an increase in a load pressure of each actuator. 4. The hydraulic drive device according to claim 3, wherein
The pump flow control valve acts on the maximum load pressure (Pm) in a direction to close the pump flow control valve via an oil passage to increase the displacement of the variable pump, and to increase the pump discharge pressure (Pd) by another. A hydraulic drive device wherein the pump flow control valve is opened via an oil passage to act in a direction to reduce the displacement of the variable pump. 5. The hydraulic drive device according to claim 3, wherein each of said pressure compensating valves is provided upstream of each of said corresponding directional control valves, and said pressure compensating valves are arranged to receive a first pressure of a first oil chamber. The area of its own downstream pressure in the direction of closing the valve, the second pressure receiving area of the second oil chamber the secondary pressure in the direction of opening the valve, and the pressure of the third oil chamber Applying the load pressure of each of the actuators to the third pressure receiving area in a direction to open the valve, and making the second pressure receiving area substantially equal to the third pressure receiving area, and setting the first pressure receiving area to be equal to the third pressure receiving area. A hydraulic drive device characterized in that it is larger than the third pressure receiving area. 6. The hydraulic drive device according to claim 5, wherein
A value obtained by dividing the third pressure receiving area of the pressure compensating valve by the first pressure receiving area is 0.99 to 0.95 (99 to 95).
%). 7. The hydraulic drive device according to claim 6, wherein
When at least two actuators of the plurality of actuators need to be driven synchronously regardless of the load pressure of the actuators, such as driving two traveling motors that rotationally drive a pair of crawlers of a hydraulic traveling vehicle. A hydraulic drive unit, wherein values obtained by dividing the third pressure receiving area of the two pressure compensating valves communicating with the two actuators by the first pressure receiving area are equal to each other. 8. The hydraulic drive device according to claim 5, wherein
A high-load-side pressure connected to a high-load-side actuator when a load pressure of a first actuator of at least two of the plurality of hydraulic actuators is significantly higher than a load pressure of the other second actuator; The value obtained by dividing the third pressure receiving area of the compensating valve by the first pressure receiving area is divided by the third pressure receiving area of the low load side pressure compensating valve communicating with the low load side actuator by the first pressure receiving area. A hydraulic drive device characterized by having a value larger than the set value. 9. The hydraulic drive device according to claim 8, wherein
The value obtained by dividing the third pressure receiving area of the pressure compensating valve of the low load side actuator by the first pressure receiving area is 1 to 0.98, and the third pressure receiving area of the pressure compensating valve of the high load side actuator is the first pressure receiving area. Wherein the value divided by the pressure receiving area is 0.97 to 0.94. 10. The hydraulic drive device according to claim 5, wherein the pressure compensating valve slides into a main body hole having two inner diameters of a main body, a small-diameter main body hole provided in the main body, and a large-diameter main body hole. A spool having a small-diameter portion and first and second large-diameter lands to be fitted to each other, and a load pressure port, a secondary pressure port, and an outlet port of an actuator sequentially provided in the main body along the main body hole , Having an inlet port and a tank port communicating with the pump discharge oil passage, wherein the small-diameter portion provided at one end of the spool that fits into the small-diameter main body hole is capable of abutting an end face of the main body hole via a spring and The third oil chamber that communicates with the load pressure port is formed between the small-diameter portion of the spool and the small-diameter portion of the spool. Contact with 1 large diameter land The second oil chamber communicating with the secondary pressure port is formed in the large-diameter main body hole surrounding the joint portion, and the piston slides in an oil-tight telescopic manner in an axial hole provided at the other end of the spool. And the other end of the piston is located in the tank oil chamber so as to be in contact with the end face of the other body hole, and a pilot oil passage is provided between the spool and the piston in the axial hole. Through the first port to the outlet port
The first pressure receiving area A1 of the first oil chamber depends on the cross-sectional area of the piston.
The pressure receiving area A2 is formed by subtracting the small-diameter hole cross-sectional area from the large-diameter main body hole cross-sectional area, and the third pressure-receiving area A3 of the third oil chamber is formed by the small-diameter portion cross-sectional area. And the spool has an openable / closable throttle portion for narrowing between the outlet port and the inlet port provided on the second large-diameter land facing the first large-diameter land portion, The second pressure receiving area A2 is substantially equal to the third pressure receiving area A3, and the third pressure receiving area A3 is set to the first pressure receiving area A3.
A hydraulic drive device characterized in that the pressure receiving area A1 is smaller than the pressure receiving area A1, and the flow rate of the pressure compensating valve connected to each actuator is reduced in accordance with an increase in the load pressure (PL) of each actuator. 11. The hydraulic drive device according to claim 5, wherein the pressure compensating valve includes a main body, a main body hole provided in the main body, and first and second slidably fitted in the main body hole. A spool having a third large-diameter land, a secondary pressure port, a load pressure port of the actuator, an outlet port, an inlet port communicating with the pump discharge oil passage, sequentially provided on the main body along the main body hole, and A tank port, and an auxiliary piston is slidably inserted in an oil-tight manner into an auxiliary axial hole provided at one end of the spool, and the other end of the auxiliary piston contacts an end surface of the main body hole. A second oil chamber is formed between the spool and the auxiliary piston in the auxiliary axial hole, and a spring is interposed between the spool and the auxiliary piston in the auxiliary axial hole. Through the load pressure port The third oil chamber is formed, and the other end of the spool forms a tank oil chamber communicating with a tank port between the other end of the main body hole and an axial hole provided at the other end of the spool. A piston is slidably inserted in an oil-tight manner in a telescoping manner, and the other end of the piston is made to be able to abut on the other end face of the body hole. The first oil chamber communicating with the outlet pressure port through a pilot oil passage is formed between the spool and the piston, and a first pressure receiving area of the first oil chamber is determined by a cross-sectional area of the piston. The second pressure receiving area of the oil chamber is formed by the area obtained by subtracting the auxiliary piston cross-sectional area from the large-diameter hole cross-sectional area, and the third pressure receiving area of the third oil chamber is formed by the auxiliary piston cross-sectional area. The spool faces the second large-diameter land portion. An openable / closable throttle portion provided between the outlet port and the inlet port provided on the third large-diameter land, wherein the second pressure receiving area and the third pressure receiving area are substantially the same, and 3. The hydraulic drive wherein the pressure receiving area of No. 3 is smaller than the first pressure receiving area A1, and the flow rate of the pressure compensating valve communicating with the actuator is reduced in accordance with an increase in the load pressure of each actuator. apparatus. 12. The hydraulic drive device according to claim 4, wherein the secondary pressure is the variable pump discharge pressure (Pd).
Differential pressure detector for detecting a differential pressure between the differential pressure and the maximum load pressure (Pm), a control device for inputting an output of the differential pressure detector to generate a control signal, and the control signal output by the control device A hydraulic drive device characterized in that the secondary pressure (Pc) is output by an electromagnetic proportional valve that operates in (1). 13. A variable pump, a plurality of hydraulic actuators driven by discharge oil of the variable pump, and a plurality of directions having a flow rate adjusting function capable of controlling pressure oil flowing into each of the plurality of actuators. A control valve;
A hydraulic drive device having a plurality of pressure compensating valves disposed between each directional control valve and each actuator for compensating the outlet pressure of each directional control valve against the highest load pressure of the actuator. In each of the pressure compensating valves, the acting force of the spring of the pressure compensating valve and the maximum load pressure (Pm) of the actuator act in a direction to close the pressure compensating valve, and the pressure compensating valve acts in a direction to open the pressure compensating valve. A differential pressure (Pc = Pd-Pm) that generates a secondary pressure (Pc = Pd-Pm) corresponding to the differential pressure between the variable pump discharge pressure (Pd) and the maximum load pressure (Pm) by applying the pressure (Pd ′) on the upstream side of the valve. A pressure control valve, a pump flow regulating valve for communicating the discharge oil of the variable pump to a displacement volume changing means of the variable pump, and the secondary pressure (Pc) of the differential pressure control valve through an oil passage. The pump flow Hydraulic drive system is characterized in that by acting to reduce the displacement of the variable pump closed control valve. 14. The hydraulic drive device according to claim 1, wherein the pressure compensating valve includes a body including a first body and a second body connected to each other, a small-diameter hole provided in the first body, A medium diameter hole following the small diameter hole, a first spool fitted to the small diameter hole, a second spool fitted to the medium diameter hole,
A large-diameter hole adjacent to the medium-diameter hole provided in the second body, an auxiliary small-diameter hole having the same diameter as the small-diameter hole following the large-diameter hole, and first and second large-diameter holes fitted into the large-diameter hole. A third spool having a diameter land and an auxiliary small diameter portion fitted into the auxiliary small diameter hole,
An auxiliary inlet port, which is sequentially provided along the body through a spring that presses the respective spools between the first spool and the end face of the body small-diameter hole, communicates with a pump discharge oil passage and communicates with the small-diameter hole, A load pressure port of the actuator communicating with the load pressure line and communicating with the medium diameter hole, a tank port communicating with the large diameter hole at a contact portion between the second spool and the third spool, and the first and second large holes; An outlet port communicating with the large-diameter hole between the large-diameter lands, an inlet port communicating with the pump discharge oil passage, and opening to be throttled by an openable and closable throttle provided in the second large-diameter land, and the actuator; A maximum load pressure port communicating with the maximum load pressure take-out line and communicating with the large diameter hole of a continuous portion between the second large diameter land and the auxiliary small diameter portion; The auxiliary small diameter An oil chamber communicating with the outlet port via a pilot oil passage is provided between the end faces, and the pressure compensating valve applies the outlet port pressure (Pz) in a closing direction to the end face of the auxiliary small diameter portion via the pilot oil passage. (Cross-sectional area B1)
The maximum load pressure (Pm) of the maximum load pressure port is
The pump discharge pressure (P) through the auxiliary inlet port in the direction of opening the pressure compensating valve to the cross-sectional area B2 obtained by subtracting the cross-sectional area of the auxiliary small-diameter portion from the cross-sectional area of the large-diameter land.
d) to the sectional area B1 of the first spool, and the actuator load pressure (PL) of the load pressure port to the sectional area B3 obtained by subtracting the sectional area B1 of the first spool from the sectional area of the second spool. Further, the sectional area B2 is made substantially equal to the sectional area B1 of the first spool (B1 = B2), and the sectional area B3 is made smaller than the sectional area B1 (= B2) of the first spool ( B1> B3), wherein the flow rate of the pressure compensating valve connected to each actuator is reduced in accordance with an increase in the load pressure (PL) of each actuator. 15. The hydraulic drive device according to claim 14, wherein the sectional area B3 of the pressure compensating valve is changed to the sectional area B1.
A hydraulic drive device characterized in that a value obtained by dividing by 0.99 is 0.99 to 0.95 (99 to 95%). 16. The hydraulic drive device according to claim 14, wherein a pair of crawlers of the hydraulic traveling vehicle are rotationally driven.
When at least two actuators of the plurality of actuators need to be driven synchronously irrespective of the load pressure of the actuators, such as when driving two traveling motors, two actuators connected to the two actuators are driven. A hydraulic drive device, wherein values obtained by dividing the sectional area B3 of the pressure compensating valve by the sectional area B1 are equal to each other. 17. The hydraulic drive device according to claim 14, wherein a load pressure of a first actuator of at least two of the plurality of hydraulic actuators is much higher than a load pressure of the other second actuator. When large, the cross-sectional area B3 of the high-load side pressure compensating valve communicating with the high-load side actuator is changed to the cross-sectional area B
A hydraulic drive apparatus wherein a value obtained by dividing by 1 is smaller than a value obtained by dividing the sectional area B3 of the low load side pressure compensating valve communicating with the low load side actuator by the sectional area B1. 18. The hydraulic drive device according to claim 17, wherein a value obtained by dividing the sectional area B3 of the pressure compensating valve of the low load side actuator by the sectional area B1 is 1 to 0.98.
A value obtained by dividing the sectional area B3 of the pressure compensating valve of the high load side actuator by the sectional area B1 is 0.97 to 0.94.
A hydraulic drive device characterized by the following.