JP2000227103A - Hydraulic transmission - Google Patents

Abstract

(57) [Summary] In a hydraulic drive device including a turning control system, acceleration can be performed without a jerky feeling when starting turning alone or in combination, and the change in turning speed when transitioning from turning alone to compound can be achieved. A system which can be suppressed and can be accelerated without excessively lowering the turning speed as compared with other actuators at the time of turning combined start-up, and does not cause a problem of an increase in cost space and a complicated circuit configuration. A pump controller for controlling a discharge flow rate so that a pump discharge pressure is higher than a maximum load pressure of actuators by a predetermined value is provided, and a pressure compensating valve is provided.
Is set as the target compensation differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the actuators 2 to 6, and when the load pressure increases on the pressure compensating valve 12, the target compensation differential pressure decreases. A lower limit setting spring 55 for preventing the target compensation differential pressure from dropping below a predetermined value.
In addition, the pressure compensating valve 12 is provided with a load-dependent characteristic that reduces the target compensation differential pressure when the load pressure increases.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

TECHNICAL FIELD The present invention relates to a hydraulic shovel or the like,
In connection with a hydraulic drive device of a construction machine including a swing control system, the discharge flow rate of the hydraulic pump is loaded when a plurality of actuators including a swing motor are supplied with hydraulic oil from the hydraulic pump via respective directional control valves. The present invention relates to a hydraulic drive device that is controlled by a sensing system and controls a differential pressure across a direction switching valve by a respective pressure compensating valve.

[0002]

2. Description of the Related Art Japanese Patent Application Laid-Open No. 60-11706 discloses a hydraulic drive device for controlling the discharge flow rate of a hydraulic pump by a load sensing system (hereinafter, appropriately referred to as an LS system). Japanese Patent Application Laid-Open No. HEI 10-102, discloses a hydraulic drive device for a construction machine including a turning control system, which is provided with an LS system and realizes independence and operability of the turning control system.
Japanese Patent No. 37907/37. Further, as an open center type hydraulic drive device for a construction machine including a swing control system, a three-pump system mounted on an actual machine is used to realize the independence of the swing control system. Japanese Patent Application Laid-Open No. H10-89304 discloses a hydraulic drive device that controls the discharge flow rate of a hydraulic pump by an LS system, in which a pressure compensating valve has a load-dependent characteristic. JP 6
The hydraulic drive device described in Japanese Patent Application Publication No. 0-11706 is provided with means for setting a differential pressure between a discharge pressure of a hydraulic pump and a maximum load pressure of a plurality of actuators as a target compensation differential pressure in each of the plurality of pressure compensating valves. During a combined operation of simultaneously driving a plurality of actuators, when the discharge flow rate of the hydraulic pump is less than the flow rate required by the plurality of directional control valves, the saturation state reduces the discharge pressure of the hydraulic pump due to this saturation state. As the differential pressure of the maximum load pressure decreases, the target compensation differential pressure of the pressure compensating valve decreases, and the discharge flow rate of the hydraulic pump can be redistributed to the ratio of the flow rate required by each actuator.

[0003] The hydraulic drive device and the three-pump system mounted on the actual machine described in Japanese Patent Application Laid-Open No. H10-37907 each have an open center type independent circuit using an independent hydraulic pump for a swivel section including a swivel motor. Thus, a separate circuit is configured with other actuators, thereby ensuring the independence and operability of the turning control system.

[0004] The hydraulic drive device described in Japanese Patent Application Laid-Open No. 10-89304 discloses a hydraulic drive device for each of a plurality of pressure compensating valves.
In the hydraulic chamber of the pressure compensating valve, the pressure receiving area of the closing-direction hydraulic chamber in which the inlet pressure of the direction switching valve is guided is determined by the pressure receiving area of the opening-direction hydraulic chamber in which the outlet pressure of the direction switching valve is guided. The target compensating differential pressure of the pressure compensating valve is reduced (the pressure compensating valve is throttled) to increase the load pressure of each actuator by increasing the load pressure of each actuator. Therefore, both the low-load side and the high-load side have good operability, do not cause hunting, and can operate stably.

[0005]

However, the above-mentioned conventional hydraulic drive has the following problems with respect to the turning control system.

JP-A-60-11706: The following problems JP-A-10-89304: The following problems JP-A-10-37907: The following problems An open center type three-pump system mounted on a real machine:
The following problems The jerky feeling of operability when turning alone is started The change in turning speed when changing from single turning operation to turning combined operation or vice versa Extreme drop in turning speed during turning combined start-up By providing another circuit Increased cost and space and complicated circuit configuration

(1) Japanese Patent Application Laid-Open No. Sho 60-11706 In a hydraulic drive device provided with an LS system described in Japanese Patent Application Laid-Open No. Sho 60-11706, when this is used for a turning control system, the turning control system has an inertia. Since a load is involved, it is difficult to balance load sensing control of the hydraulic pump (hereinafter referred to as LS control as appropriate) with the flow rate compensation function of the pressure compensation valve.
This is because it is difficult to balance the responsiveness of the pressure compensating valve with the responsiveness of the LS control of the hydraulic pump when controlling the turning drive pressure at the stage of shifting from the turning acceleration to the steady rotation for the following reason. It is mentioned.

(1) In order to maintain a constant flow rate during turning start and acceleration, the pump LS control controls the discharge pressure of the hydraulic pump to be high according to the turning start pressure.

(2) The pressure compensating valve operates in the direction of increasing the flow rate, which tends to decrease as the load pressure increases, in order to keep the differential pressure across the throttle element of the directional control valve constant.

(3) Since the swing drive pressure decreases when the swing reaches a steady speed, the pump LS control does not need to control the discharge pressure of the hydraulic pump as high as during start-up and acceleration. Works.

(4) The pressure compensating valve operates in a direction to decrease the passing flow rate, which tends to increase due to a decrease in the turning drive pressure.

Since the transitions of the above (1) to (4) are steep, the turning operability becomes jerky (described above).

In addition, when the discharge flow rate of the hydraulic pump falls into a saturation state that is less than the flow rates required by the plurality of directional control valves during the combined operation as described above, the target compensation of each of the pressure compensating valves depends on the saturation state. The differential pressure is reduced and the hydraulic pump discharge flow is redistributed to the ratio of the flow required by each actuator. With this function, the speed of the respective actuators is reduced even during the combined operation, but since the actuators operate at a ratio intended for the operation, the operation feeling is not spoiled.

However, this speed reduction similarly occurs in the turning operation, and the turning speed is reduced as in the case of the other actuators in the combined operation including the turning. When the speed is reduced from the combined swing operation to the single swing operation, or vice versa, a change in the swing speed occurs, giving an uncomfortable feeling to the operator (described above).

(2) Japanese Patent Application Laid-Open No. 10-89304 In the hydraulic drive device described in Japanese Patent Application Laid-Open No. 10-89304, the pressure compensating valve has a load-dependent characteristic. The target compensating differential pressure of the pressure compensating valve decreases according to the load pressure, and when the state shifts to the steady state, the target compensating differential pressure of the pressure compensating valve also returns to the original value according to the reduced load pressure of the swing motor, thereby enabling the turning operability. You can start turning without the jerky feeling of. However, when the discharge flow rate of the hydraulic pump becomes saturated during the combined swing operation, the discharge flow rate of the hydraulic pump is redistributed to the ratio of the flow rates required by the respective directional control valves.
This is the same as the hydraulic drive device described in Japanese Patent Application Laid-Open Publication No. H10-207, and when shifting from a combined swing operation to a single swing operation, or vice versa, a change in the swing speed occurs, giving an uncomfortable feeling to the operator (described above).

In addition, since the pressure compensating valve has a load-dependent characteristic, the pressure compensating valve in the swivel section at the time of turning complex start-up operates according to the discharge flow rate of the hydraulic pump. , The target compensation differential pressure also decreases due to the load-dependent characteristic in which the load pressure of the swing motor rises to the relief pressure, and the decrease in the target compensation differential pressure continues until it shifts to a steady state. As a result, the turning speed at the time of the combined swing start is extremely lower than that of the other actuators, and the swing operability of the combined swing start is impaired (described above).

(3) The hydraulic drive described in JP-A-10-37907 and the open center type three-pump system mounted on an actual machine In the hydraulic drive described in JP-A-10-37907, the turning control system is opened. By configuring with a center-type separate circuit, turning operability is ensured in the LS system. In addition, even in the open center type three pump system mounted on the actual machine, the swing control system is a separate circuit of the open center type, and the swing operability is secured.

That is, in the case of the open center type, when the driving pressure rises at the time of turning start, the flow rate flowing back to the tank via the center bypass oil passage increases, so that the pressure oil passing through the throttle of the direction switching valve of the turning section is increased. The flow rate decreases.
For this reason, the flow rate of the pressure oil supplied to the swing motor is
Limited during acceleration. When the turning speed reaches the steady speed,
Since the drive pressure is not as high as at the start, the flow rate is no longer limited, and a flow rate corresponding to the opening of the throttle of the directional control valve in the turning section is supplied to the turning motor. Thereby, the turning start can be smoothly performed without generating the jerky feeling of the operability at the time of starting the turning alone such as the LS control.

The above problem occurs not only in the LS system but also in an open center type system. However, a hydraulic drive device disclosed in JP-A-10-37907 and an open center type three pump system mounted on a real machine are disclosed. By configuring the turning control system with another circuit of the open center type, the independence of the turning control system is realized and the turning speed does not change.

However, in a hydraulic drive device and an open center type three-pump system mounted on an actual machine described in Japanese Patent Application Laid-Open No. Hei 10-37907, the swing control system is configured in parallel with a separate circuit from other actuator systems. Inevitably, the cost increases and the installation space increases, and a hydraulic pump for the swing control system must be separately provided. Particularly, in the system disclosed in JP-A-10-37907, a parallel pump is required. In order to balance the power with the LS system arranged in the system, a signal path is required, and the circuit configuration becomes complicated (described above).

An object of the present invention is to provide a hydraulic drive system including a turning control system, which can be accelerated without a jerky feeling of turning operability and can be shifted to a steady state at the start of turning alone or in combination, and that turning can be performed independently. The change of the turning speed during the transition from the operation to the turning combined operation or vice versa is suppressed, and the turning speed does not become extremely slow compared to other actuators at the time of starting the compound, excellent turning operability and turning independence It is an object of the present invention to provide a hydraulic drive device which can secure the cost and space and increase the complexity of the circuit configuration by providing a separate circuit.

[0022]

(1) In order to achieve the above object, the present invention provides a hydraulic pump, a plurality of actuators including a turning motor driven by hydraulic oil discharged from the hydraulic pump, A plurality of directional control valves that respectively control flow rates of pressure oil supplied from the hydraulic pump to the plurality of actuators; a plurality of pressure compensating valves that respectively control differential pressures before and after the plurality of directional change valves; A pump control means for load sensing control that controls a pump discharge flow rate so that a discharge pressure of a pump is higher than a maximum load pressure of the plurality of actuators by a predetermined value; ,
A first means provided on a pressure compensating valve other than the turning section related to the turning motor, for setting a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators as a target compensation differential pressure; A second means provided on the pressure compensating valve of the section, for setting the target compensation differential pressure; and a pressure compensating valve of at least the turning section of the plurality of pressure compensating valves, wherein the load pressure of the turning motor is When the pressure rises, the target compensation differential pressure set by the second means is reduced, and the pressure compensation valve of the swivel section is provided with a load dependent characteristic. And a fourth means for setting a lower limit of the target compensation differential pressure set by the two means and corrected by the third means.

In the present invention configured as described above,
By providing a third means to the pressure compensating valve of the turning section and having a load dependent characteristic, the pressure compensating valve of the turning section finely adjusts the flow rate according to a change in the load pressure of the turning motor at the start of turning, and the turning motor has It accelerates smoothly and shifts to a steady state.

Further, the second means for setting the target compensation differential pressure of the pressure compensating valve in the swivel section, as in the first means, calculates the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators. Means for setting the target compensation differential pressure may be used. In this case, by providing the fourth means as described above, the fourth means can reduce the target compensation differential pressure itself set by the second means. It functions as the lower limit setting means for both the reduction of the target compensation differential pressure due to the load-dependent characteristic given by the third means (see (2) below).
As a result, when the discharge flow rate of the hydraulic pump is in a saturation state and the target compensation differential pressure of the swing section pressure compensating valve is about to decrease, or when the load pressure of the swing motor becomes high, the target compensation differential of the swing section pressure compensating valve is increased. The fourth means limits the reduction of the target compensation differential pressure when the pressure tends to decrease due to the load-dependent characteristic, or when they occur at the same time, so that the hydraulic fluid is preferentially supplied to the swing motor. As a result, a change in the turning speed during the transition from the single turning operation to the turning combined operation or vice versa is suppressed, and the turning speed is not extremely slower than other actuators at the time of starting the combined operation, and the excellent turning operation is performed. And independence of turning.

The second means for setting the target compensation differential pressure of the pressure compensation valve in the swivel section sets a value which does not change due to the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as the target compensation differential pressure. In this case, the fourth means functions as a lower limit setting means for the reduction of the target compensation differential pressure due to the load-dependent characteristic given by the third means (see (3) below). )reference). As a result, even when the discharge flow rate of the hydraulic pump is in a saturation state, the target compensation differential pressure of the pressure compensation valve of the swing section does not decrease, and the load pressure of the swing motor becomes high, so that the target compensation of the pressure compensation valve of the swing section is increased. When the differential pressure is about to decrease due to the load-dependent characteristic, the fourth means limits the decrease in the target compensation differential pressure so that the reduction of the target compensation differential pressure due to saturation or the load-dependent characteristic occurs singly or simultaneously. , The pressure oil is supplied to the swing motor preferentially. As a result, a change in the turning speed during the transition from the single turning operation to the turning combined operation or vice versa is suppressed, and the turning speed is not extremely slower than other actuators at the time of starting the combined operation, and the excellent turning operation is performed. And independence of turning.

Further, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.

(2) In the above (1), preferably,
The second means is means for setting a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators as the target compensation differential pressure, similarly to the first means,
The fourth means functions as a lower limit setting means for both a decrease in the target compensation differential pressure itself set by the second means and a decrease in the target compensation differential pressure due to the load-dependent characteristic given by the third means. I do.

As a result, as described in the above (1), when the discharge flow rate of the hydraulic pump is in a saturation state and the target compensation differential pressure of the pressure compensation valve in the swing section is about to decrease, or the load pressure of the swing motor is reduced. When the target compensating differential pressure of the pressure compensating valve of the swivel section is going to decrease due to the load dependent characteristic due to the high pressure, or when they occur simultaneously, the fourth means limits the decrease of the target compensating differential pressure, and Pressure oil is supplied preferentially, and excellent turning operability and turning independence can be secured.

(3) In the above (1), the second means sets a value which does not change due to a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators as the target compensation differential pressure. It may be a setting means,
In this case, the fourth means functions as a lower limit setting means with respect to a decrease in the target compensation differential pressure due to the load dependence characteristic given by the third means.

Thus, as described in the above (1), even if the discharge flow rate of the hydraulic pump is in the saturation state, the target compensation differential pressure of the pressure compensation valve in the swing section does not decrease, and the load pressure of the swing motor is reduced. When the pressure becomes high and the target compensation differential pressure of the pressure compensation valve in the swivel section is about to decrease due to the load-dependent characteristic, the fourth means limits the decrease in the target compensation differential pressure, and the target compensation due to the saturation or the load-dependent characteristic. Regardless of whether the differential pressure is reduced alone or simultaneously, pressure oil is preferentially supplied to the turning motor, and excellent turning operability and turning independence can be secured.

(4) Further, in the above (1) to (3), preferably, the fourth means is set by the second means, and the target compensation differential pressure corrected by the third means is a predetermined value. And a biasing means for applying a biasing force in the opening direction to the spool of the pressure compensating valve in the swivel section when the pressure reaches the pressure-reducing valve.

Thus, the fourth means sets the lower limit of the target compensation differential pressure without lowering the target compensation differential pressure of the pressure compensating valve in the turning section to a value equal to or less than the value corresponding to the urging force applied by the urging means. Become.

(5) In the above (4), preferably,
When the target compensation differential pressure set by the second means and corrected by the third means reaches a predetermined value, the biasing means acts on the spool of the pressure compensating valve of the turning section, and opens the spool. It is a lower limit setting spring that urges in the direction.

Thus, the urging means applies an urging force in the opening direction to the spool of the pressure compensating valve of the turning section when the target compensation differential pressure of the pressure compensating valve of the turning section reaches a predetermined value.

(6) In the above (1) and (2), preferably, the fourth means always assists the target compensation differential pressure set by the second means and corrected by the third means. The direction switching valve of the revolving section is configured such that the opening area of the meter-in variable throttle is equivalent to the target compensation pressure of the auxiliary value added by the biasing means. It is configured to be smaller than the opening area of the direction switching valve other than the turning section.

Accordingly, the fourth means limits the decrease in the target compensation differential pressure of the pressure compensating valve in the swivel section by the auxiliary value added by the urging means, and sets the lower limit of the target compensation differential pressure. Becomes

(7) In the above (6), preferably,
The urging means is a turning priority spring which always acts in the opening direction of the spool of the pressure compensating valve of the turning section.

Accordingly, the biasing means always adds an auxiliary value to the target compensation differential pressure of the pressure compensation valve in the turning section.

[0039]

Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 shows a hydraulic drive device according to a first embodiment of the present invention. The hydraulic drive device includes a plurality of hydraulic pumps 1 and a turning motor 2 driven by hydraulic oil discharged from the hydraulic pump 1. Actuators 2 to 6, closed-center type directional control valves 7 to 11 for controlling flow rates of pressure oil supplied from hydraulic pump 1 to actuators 2 to 6, respectively, and directional control valves 7 to 11 Pressure compensating valves 12 to 1 for respectively controlling the differential pressure before and after
6, load check valves 17a to 17e disposed between the directional control valves 7 to 11 and the pressure compensating valves 12 to 16 to prevent the backflow of the pressure oil, and a plurality of actuators 2 to 3 And a pump controller 18 for load sensing control for controlling the pump discharge flow rate so as to be higher than the maximum load pressure of No. 6 by a predetermined value. The actuator line of the swing motor 2 has an overload relief valve 60
a, 60b are provided. Other actuators 3 ~
6, a similar overload relief valve is provided, but is not shown.

The plurality of directional control valves 7 to 11 are provided with self-load pressure detection lines 20 to 24, and the maximum load pressure among the load pressures detected by these detection lines 20 to 24 is signal lines 25 to 29. , Are detected via shuttle valves 30-33 and signal lines 34-36, and are led out to a signal line 37.

The pump control device 18 includes a tilt control actuator 40 connected to a swash plate 1 a which is a variable capacity member of the hydraulic pump 1, and a hydraulic chamber 40 of the actuator 40.
a, a load sensing control valve (hereinafter referred to as LS) for switching and controlling the connection between the hydraulic pump 1 and the discharge oil passage 1 b of the hydraulic pump 1 and the tank 19.
41 (referred to as a control valve). The discharge pressure of the hydraulic pump 1 and the maximum load pressure of the signal line 37 act as control pressures on the LS control valve in opposition. When the pump discharge pressure becomes higher than the total pressure of the maximum load pressure and the set value of the spring 41a (target LS differential pressure), the hydraulic chamber 40a of the actuator 40 is connected to the discharge oil passage 1b of the hydraulic pump 1, and the hydraulic chamber 40a The piston 40b is moved to the spring 40
It moves to the left in the figure overcoming the force of c and reduces the tilt of the swash plate 1a to reduce the discharge flow rate of the hydraulic pump 1. vice versa,
When the pump discharge pressure becomes lower than the total pressure of the maximum load pressure and the set value of the spring 41a (target LS differential pressure), the hydraulic chamber 40a of the actuator 40 is connected to the tank 19,
By reducing the pressure in the hydraulic chamber 40a, the piston 40b is moved rightward in the figure by the force of the spring 40c, thereby increasing the tilt of the swash plate 1a and increasing the discharge flow rate of the hydraulic pump 1. Such LS
By the operation of the control valve, the discharge flow rate of the hydraulic pump 1 is controlled so that the pump discharge pressure becomes higher than the maximum load pressure by the set value of the spring 41a (target LS differential pressure).

The pressure compensating valves 12 to 16 act on the upstream side of the directional control valves 7 to 11 in the closing direction, respectively.
Detection line 2 which is the pressure on the downstream side of the directional control valves 7 to 11
The pressure (load pressure) of 0 to 24 is applied in the opening direction, and the maximum load pressure derived from the signal line 37 is applied in the closing direction, so that the discharge pressure of the hydraulic pump 1 is applied in the opening direction. Pressure between the discharge pressure of the hydraulic pump 1 controlled as described above and the maximum load pressure (hereinafter referred to as L
S-control differential pressure) is used as the target compensation differential pressure to control the differential pressure across the directional control valves 7-11.

The pressure on the upstream side of each of the directional control valves 7 to 11 acting on the pressure compensating valves 12 to 16 is applied to a signal line 50.
The pressures (load pressures) of the detection lines 20 to 24, which are taken out by a to 50e and are downstream of the direction switching valves 7 to 11, are taken out by the signal lines 51a to 51e, and the maximum load pressure of the signal line 37 is a signal. Lines 52 and 52
The discharge pressure of the hydraulic pump 1 is extracted by signal lines 53 and 53a to 53e. In the pressure compensating valves 13 to 16, the signal line 52b
The maximum load pressure taken out by the oil chamber 13a
16a, and the discharge pressure of the hydraulic pump 1 taken out by the signal lines 53b to 53e is changed to the oil chambers 13b to 1b.
6b to set the target compensation differential pressure. The oil chamber for setting the target compensation differential pressure of the pressure compensation valve 12 will be described later.

The pressure compensating valve 12 causes the pressure on the upstream side of the directional control valve 7 to act in the closing direction, and the pressure on the detection line 20 which is the pressure on the downstream side of the directional control valve 7 (the rotation motor 2).
When the load pressure of the swing motor 2 increases when the load pressure is applied in the opening direction, the target compensation differential pressure is reduced so as to limit the flow rate of the pressure oil passing through the direction switching valve 7. In addition to the above configuration, a lower limit setting spring 55 is provided on the opening direction working side which is the target compensation differential pressure setting side. The lower limit setting spring 55 acts on the spool of the pressure compensating valve 12 only when the target compensating differential pressure of the pressure compensating valves 13 to 16 of the other sections becomes lower than the set value of the spring 55, and the target compensating differential pressure is reduced. The lower limit is set so that it does not become smaller than the set value.

FIG. 2 shows the structure of the pressure compensating valve 12.

In FIG. 2, the pressure compensating valve 12 has two bodies, a first body 301a and a second body 301b, and these bodies are appropriately assembled together by bolting or the like (not shown). ing. The first body 301a has a small-diameter hole 321 and a medium-diameter hole 322 following the small-diameter hole 321.
The first spool 311 having a diameter d1 is slidably fitted in the small-diameter hole 321 and the diameter d3 is fitted in the medium-diameter hole 322.
The second spool 312 (> d1) is slidably fitted. The second body 301b is provided with a large-diameter hole 323 following the medium-diameter hole 322, and a small-diameter hole 325 following the large-diameter hole 323 and having the same diameter as the small-diameter hole 321.
The third spool 310 is slidably fitted in the third and small diameter holes 325, and the third spool 310 is slidably fitted in the large diameter hole 323 in the first and second diameters d 2 (> d 3). It has large-diameter portions 313 and 314 and a small-diameter portion 315 having a diameter d1 slidably fitted in the small-diameter hole 325.

A convex portion 321a is provided on an end surface of the small diameter hole 321. An oil chamber 331 is formed around the convex portion 321a, and a concave portion 311a for receiving the convex portion 321a is provided on an end surface of the first spool 311. A weak spring 350 for holding the initial position is provided between the end surface of the convex portion 321a and the bottom of the concave portion 311a to press the spools in the closing direction. The chamber in which the spring 350 is disposed communicates with an external oil chamber 331 through a passage 321b formed in the projection 321a.

The lower limit setting spring 55 is disposed around the convex portion 321 of the oil chamber 331, and faces the end surface of the first spool 311. The lower limit setting spring 55 is merely apart from the end face of the first spool 311 in the initial position shown in the figure, and does not generate a force for pushing the spools in the closing direction.

The body 301a has a pump port 3
41 and a load pressure port 342 are formed.
b, a tank port 343, an outlet port 344, an inlet port 345, and a maximum load pressure port 346 are formed. The pump port 341 communicates with the discharge pressure signal line 53 a of the hydraulic pump 1 and opens to the oil chamber 331, and the load pressure port 342 communicates with the load pressure signal line 51 a and has the small-diameter hole 321 and the medium-diameter hole 322. Are open to the oil chamber 332 formed at the connection portion of the. In addition, tank port 343
Is connected to the tank 19 and the second spool 312 and the third spool
An opening is provided in an oil chamber 333 provided in a large-diameter hole 323 surrounding a contact portion with the spool 310, and an outlet port 344 is connected to the load check valve 17a and provided between the first and second spool large-diameter portions 313 and 314. Oil chamber 32 provided in large-diameter hole 323 of
8 and the inlet port 345 is connected to the pump discharge oil passage 1b.
And the second large diameter portion 314 of the third spool 310
The maximum load pressure port 346 is opened at the entrance side of the openable and closable throttle unit 316 provided at the maximum load pressure signal line 5.
2a and the second large-diameter portion 3 of the third spool 310
The large-diameter hole 323 where the continuous portion of the 14 and the small-diameter portion 315 is located
The oil chamber 336 provided in the portion of the figure is opened.

The pilot oil passage 50 provided in the third spool 310 is provided between the small diameter portion 315 and the small diameter hole end surface 330.
An oil chamber 334 that communicates with the oil chamber 328 of the outlet port 344 via a is provided.

First body 301a and second body 301b
Are appropriately assembled integrally by a method such as bolting (not shown) to form the body 301. At this time, the first body 301
a-side medium diameter hole 322 and second body 301b-side large diameter hole 323
Even if they are misaligned, there is no operational problem since the second spool 312 and the third spool 310 are merely abutted with separate components.

With the above configuration, the pressure compensating valve 12 applies the outlet pressure (Pz) of the outlet port 344 in the closing direction to the pressure receiving area B1 of the end surface 340 of the small diameter portion 315 in the oil chamber 334 via the pilot oil passage 50a. Maximum load pressure port 346
The maximum load pressure (PLmax) is applied to the pressure receiving area B2 of the step portion obtained by subtracting the sectional area of the small diameter portion 315 from the sectional area of the second large diameter portion 314 in the oil chamber 336. Further, the pressure compensating valve 12 applies the pump discharge pressure (Ps) to the pressure receiving area B1 of the end face of the first spool 311 in the oil chamber 331 in the opening direction via the pump port 341 and the load pressure (PL) of the load pressure port 342. From the sectional area of the second spool 312 in the oil chamber 332 to the sectional area B1 of the first spool 311
Is applied to the pressure receiving area B3 of the step portion from which is subtracted. The pressure receiving area of the stepped portion obtained by subtracting the cross-sectional area of the second spool 312 from the cross-sectional area of the first large-diameter portion 313 in the oil chamber 333 is because the oil chamber 33 communicates with the tank 19 through the tank port 343. The operating force for opening and closing the spools does not work.

Then, the pressure receiving area B2 and the pressure receiving area B1 of the first spool 311 are made substantially the same (B1 = B2).
In addition, the pressure receiving area B3 is the pressure receiving area B1 (= B
2) A load-dependent characteristic that is reduced (B1> B3) to reduce the flow rate of the directional control valve 7 that passes through the swing motor 2 in accordance with an increase in the load pressure (PL) of the swing motor 2. It is.

That is, considering the hydraulic balance of the first spool 311, the second spool 312, and the third spool 313, the pressure compensating valve 12 functions by balancing B1Pz-B3PLmax with B1Ps-B2PLmax. Holds.

From B1Ps-B2PLmax = B1Pz-B3PL B1 = B2, B1 (Ps-PLmax) = B2Pz-B3PL Ps-PLmax is the differential pressure between the discharge pressure Ps of the LS-controlled hydraulic pump 1 and the maximum load pressure PLmax ( (LS control differential pressure), so that ΔPc: B1ΔPc = B2Pz−B3PL (1) If the differential pressure across the directional control valve 7 is ΔP, then ΔP = Pz−PL. Also, by transforming equation (1), B1ΔPc + (B3−B2) PL = B2 (Pz−PL) Therefore, ΔP = Pz−PL = (B1 / B2) ΔPc− (1− (B3 / B2)) PL .. (2) Here, if B1 / B2 = α and B3 / B2 = β, ΔP = Pz−PL = αΔPc− (1−β) PL (3) That is, if B2 = B3, (B2 And B3 if there is no area difference), ΔP = αΔPc and ΔP is determined only by ΔPc (LS control differential pressure).
Since there is an area difference in B3, ΔP is affected by the load pressure PL due to the area difference, and ΔP increases as the load pressure PL increases.
And the load dependent characteristic of reducing the flow rate through the directional control valve 7.

FIG. 3 shows load-dependent characteristics of the pressure compensating valve 12. The horizontal axis in FIG. 3 is the load pressure, represented by PL, and the vertical axis is the target compensation differential pressure, represented by ΔPv. Dotted lines indicate target compensation differential pressures of the pressure compensating valves 13 to 16 other than the section of the swing motor 2 (hereinafter, referred to as a swing section). The pressure compensating valves 13 to 16 other than the turning section maintain the target compensation differential pressure ΔPv at the LS control differential pressure ΔPc even if the load pressure PL of the actuators 3 to 6 increases. 12, the target compensation differential pressure ΔPv increases as the load pressure PL increases.
Becomes smaller.

FIG. 4 shows a lower limit setting function of the target compensation differential pressure by the lower limit setting spring 55 when it is assumed that the pressure compensating valve 12 has no load-dependent characteristic. The horizontal axis in FIG. 4 is the sum of the flow rates (valve required flow rates) required by the direction switching valve 7 and the other direction switching valves 8 to 11, and is represented by Qr. This corresponds to the total lever operation amount of an operation lever device (not shown) for switching the direction switching valves 7 to 11, that is, the total required flow rate of the swing motor 2 and its actuator.
The vertical axis is the pressure compensating valve 12 and other pressure compensating valves 13-1.
6 is the target compensation differential pressure ΔPv. Further, the set differential pressure of the lower limit setting spring 55 (the lower limit value of the target compensation differential pressure) is P
b.

In the combined swing operation in which the swing motor 2 and the other actuators are simultaneously driven, the sum Qr of the required flow rates of the direction switching valve 7 and the other direction switching valves 8 to 11 is larger than the maximum discharge flow rate Qpmax of the hydraulic pump 1. When the discharge flow rate of the hydraulic pump 1 is not in the saturation state, the target compensation differential pressure ΔPv of all the pressure compensating valves including the pressure compensating valve 12 is constant at the LS control differential pressure ΔPc.

The sum Qr of the required valve flows is equal to the hydraulic pump 1
Exceeds the maximum discharge flow rate Qpmax, and when the discharge flow rate of the hydraulic pump 1 is in the saturation state, the LS control differential pressure Δ
Until Pc decreases to the set differential pressure Pb of the lower limit setting spring 55 of the pressure compensating valve 12 in the turning section, the target compensation differential pressures ΔPv of all the pressure compensating valves decrease as the LS control differential pressure ΔPc decreases, and the LS control The differential pressure ΔPc is the lower limit setting spring 55
After that, the target compensation differential pressure ΔPv of the pressure compensating valve 12 in the turning section is held at the set differential pressure Pb of the lower limit setting spring 55, and does not decrease below that, and the other than in the turning section , The target compensation differential pressure ΔPv of the pressure compensating valve continues to decrease as the LS control differential pressure ΔPc decreases.

In the drawing, the thick broken line indicates the pressure compensating valves 13 to 1 other than the swivel section in the combined operation including the swivel section.
6 is a change in the target compensation differential pressure ΔPv, and the thin dashed line indicates the pressure compensating valves 13 to 13 in the combined operation that does not include the turning section.
16 shows a change in the target compensation differential pressure ΔPv. The target compensation differential pressure ΔPv of the pressure compensating valves 13 to 16 other than the turning section in the combined operation including the turning section is the target compensation differential pressure ΔPv of the pressure compensating valve 12 of the turning section.
Since the differential pressure Pb does not become smaller than the set differential pressure Pb of No. 5, the pressure compensating valves 13 to 16 in the combined operation not including the swivel section
Is greater than the target compensation differential pressure ΔPv.

The above-described hydraulic drive device is mounted on, for example, a hydraulic shovel. FIG. 5 shows the appearance of the excavator. In FIG. 5, the excavator includes a lower traveling body 20.
0, upper revolving superstructure 201, front work machine 202,
The upper revolving unit 201 is capable of revolving around the axis O on the lower traveling unit 200, and the front work machine 202 is
01 can be moved up and down at the front. Front work machine 202
Is an articulated structure having a boom 203, an arm 204 and a bucket 205, and the boom 203 is a boom cylinder 2
At 06, the arm 204 is driven by the arm cylinder 207 and the bucket 205 is driven by the bucket cylinder 208 in a plane including the axis O. The swing motor 2 shown in FIG. 1 is an actuator that drives the upper swing body 202 to swing on the lower traveling body 200, and three of the actuators 3 to 6 are used as a boom cylinder 206, an arm cylinder 207, and a bucket cylinder 208.

In the above, the oil chambers 13a to 16a and 13b to 16b connected to the signal lines 52b to 52e and 53b to 53e of the pressure compensating valves 13 to 16 are connected to the swing motor 2 among the plurality of pressure compensating valves 12 to 16. The first means is provided in the pressure compensating valves 13 to 16 other than the relevant swivel section and sets a differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as a target compensation differential pressure. Chamber 336 (pressure receiving area B2 = B1) and oil chamber 33 connected to signal lines 52a and 53a of pressure compensating valve 12
1 (pressure receiving area B1) is the pressure compensating valve 1
The oil chamber 334 (pressure receiving area B1> B3) connected to the signal lines 50a and 51a of the pressure compensating valve 12 and the oil chamber 33 constitute second means for setting the target compensation differential pressure.
2 (pressure receiving area B3) is provided at least in the pressure compensating valve 12 of the turning section among the plurality of pressure compensating valves 12 to 16, and when the load pressure of the turning motor 2 increases, the second
A third means for reducing the target compensation differential pressure set by the means and for giving the load compensating characteristic to the pressure compensating valve 12 of the swivel section is constituted. Fourth means is provided in the valve 12 and sets a lower limit of the target compensation differential pressure set by the second means and corrected by the third means.

In the present embodiment, the second means (oil chambers 331, 336) is replaced by the first means (oil chambers 13a to 13a).
6a, 13b to 16b), means for setting the differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as the target compensation differential pressure. ) Is the second means (oil chambers 331, 33)
It functions as a lower limit setting means for both the decrease in the target compensation differential pressure set in 6) and the decrease in the target compensation differential pressure due to the load-dependent characteristics given by the third means (oil chambers 332, 334).

Further, the fourth means (lower limit setting spring 55)
Is set by the second means (oil chambers 331 and 336), and the third
When the target compensation differential pressure corrected by the means (oil chambers 332, 334) reaches a predetermined value, the pressure compensation valve 12
Is an urging means for applying an urging force in the opening direction to the spool 311.

The operation of the present embodiment configured as described above will be described. 1. FIG. 6 shows the operation of the turning direction switching valve 7 and the turning motor 2
The time chart shows the behavior of the pressure compensating valve 12 for turning at the time of the turning alone operation of driving the motor alone.

At the time of starting the swing independent operation, the upper swing body 20
There is an increase in load pressure specific to inertial load. This increase in the load pressure is restricted by a safety valve such as the overload relief valve 60a or 60b provided in the swing motor 2. In this state, the pressure oil supplied to the swing motor 2 is
It is discharged to the tank from the safety valve 60a or 60b.

In the conventional general pressure compensating valve, the upper revolving unit 2 which is an inertial load is released by releasing the pressure oil from the safety valve.
I was adjusting the feeling of acceleration of 01. However, in this case, since the consumption flow rate of the swing motor at the time of startup is small, most of the pressure oil is discharged to the tank, resulting in energy loss. Further, it is difficult to balance the LS control of the hydraulic pump and the flow rate compensation function of the pressure compensating valve, and the turning operability becomes jerky.

On the other hand, in the present embodiment, such a problem does not occur because the pressure compensating valve 12 of the turning section has the load-dependent characteristic as described above.

First, before starting the operation lever of the turning operation lever device, the target compensation differential pressure ΔPv of the pressure compensating valve 12 is controlled to the LS control differential pressure ΔPc (t0 to t1).

When the swing motor 2 is started by operating the operation lever, the load pressure PL is increased by the inertial load at the same time as the start (t1).

Due to the load-dependent characteristics of the pressure compensating valve 12, the target compensation differential pressure ΔPv decreases from the LS control differential pressure ΔPc and stops decreasing at the set differential pressure Pb of the lower limit setting spring 55 (t1). The supply flow rate Qa to the swing motor 2 is controlled to a flow rate corresponding to the set differential pressure Pb of the spring 55. If there is no lower limit setting spring 55,
The target compensation differential pressure ΔPv falls to a pressure lower than Pb (it does not become 0).

When the upper swing body 201 starts rotating and the swing speed increases, the flow rate of the swing motor 2 and the swing motor 2
The supply flow Qa to the balance is balanced, and the load pressure gradually decreases. As a result, the target compensation differential pressure ΔPv of the pressure compensation valve 12 also increases (t2).

When the consumption flow rate of the swing motor 2 and the supply flow rate Qa are not balanced, the load pressure PL rises or falls and is fed back to the pressure compensation valve 12 of the swing section. Due to the load pressure dependent characteristic of the pressure compensating valve 12, when the supply flow rate Qa is too large, the load pressure PL increases, and as a result, the supply flow rate Qa is limited by the pressure compensating valve 12.
Conversely, when the supply flow rate Qa becomes insufficient, the load pressure PL decreases, and the supply flow rate Qa is increased by the pressure compensating valve 12.
By the fine adjustment of the pressure compensating valve 12, the swing motor 2 accelerates slowly without causing hunting as occurs in the conventional LS control.

When the flow rate reaches the original supply flow rate, a steady state is reached (t3), and the load pressure PL becomes a pressure corresponding to the rotation resistance. 2. Activation of other actuators during turning steady rotation FIG. 7 shows the behavior of the pressure compensating valve of each section when another actuator, for example, a boom cylinder is started and combined operation is performed while the rotation is rotating alone. Shown in the time chart. It is assumed that the boom cylinder is the actuator 3.

At the time of the single rotation of the swing, the load pressure PL of the swing motor 2 is reduced to the pressure required for the steady rotation, and the target compensation differential pressure ΔPv of the pressure compensating valve 12 is substantially equal to the LS control differential pressure ΔPv.
It is controlled to Pc (t0 to t1).

When the operation lever of the operation lever device for the boom is additionally operated, the swing motor 2 and the boom cylinder 3
If the required flow rate exceeds the maximum discharge flow rate that can be supplied by the hydraulic pump 1 and saturation occurs,
Due to the decrease in the LS control differential pressure ΔPc in proportion to the supply shortage with respect to the required flow rate Qr, the target compensation differential pressure ΔPv of each of the pressure compensating valves 12 and 13 decreases, and the flow is redistributed (t).
1).

Here, when the degree of saturation is large, the target compensation differential pressure ΔPv greatly decreases, but the decrease in the target compensation differential pressure ΔPv of the pressure compensating valve 12 in the turning section is reduced by the set differential pressure Pb of the lower limit setting spring 55. Limited by
For this reason, the target compensation differential pressure ΔPv of the boom section pressure compensating valve 13 is further reduced by the amount by which the reduction of the turning-side target compensation differential pressure ΔPv is limited.

As a result, at the time of the combined operation including turning,
Pressure oil can be supplied to the swing motor 2 preferentially to some extent. With this function, it is possible to realize independent operability of the turning motor 2 with respect to other actuators in the saturation state, to suppress a change in turning speed during combined operation, and to ensure turning operability.

As a comparative example, in a combined operation that does not include turning, the target compensation differential pressure ΔPv decreases to the same value due to a decrease in the LS control differential pressure ΔPc due to saturation, and the supply flow rate Qa also decreases to the same value (in the combined operation). It is assumed that the opening areas of the related directional control valves are the same). When there is no lower limit setting spring 55 in the pressure compensating valve 12 of the turning section (Japanese Patent Laid-Open No. 10-8 / 1998)
The same applies to the composite operation including the turning of the case of No. 9304). By providing the lower limit setting spring 55, the target compensation differential pressure ΔPv and the supply flow rate Qa of the turning section are reduced by ΔΔPv1 and ΔQa1 compared to that case. , And pressure oil is preferentially supplied to the turning motor 2, so that a change in turning speed during combined operation can be suppressed.

FIG. 8 shows the hydraulic pump 1 in the combined operation.
Is a case where the degree of saturation of the discharge flow rate is small.

When the degree of saturation is small, the decrease in the target compensation differential pressure ΔPv remains at or above the set differential pressure Pb of the lower limit setting spring 55. In this case, the target compensation differential pressure ΔPv and the flow rate Qa are reduced to the same value for both the turning and the boom (assuming that the opening areas of the directional control valves 7 and 8 of the turning and the boom sections are the same).

As described above, by setting the lower limit setting spring 55, the priority of turning can be set by the degree of saturation. 3. Simultaneous Start of Swing and Other Actuator FIG. 9 is a time chart showing the behavior of the pressure compensating valves of the respective sections during a combined operation in which another actuator, for example, a boom cylinder is started simultaneously at the time of start of swing. Also in this case, the boom cylinder is the actuator 3.

First, in a state before starting the operation levers of the turning and boom operation lever devices, the target compensation differential pressure ΔPv of the pressure compensating valves 12 and 13 is controlled to the LS control differential pressure ΔPc ( t0 to t1).

When the swing motor 2 and the boom cylinder 3 are simultaneously operated by simultaneously operating the swing and boom operation levers, the required flow rate of the swing and the boom is adjusted by the hydraulic pump 1.
When the maximum discharge flow rate exceeds the maximum discharge flow rate and saturation occurs, the target compensation differential pressure ΔPv of each of the pressure compensating valves 12 to 16 decreases due to a decrease in the LS control differential pressure ΔPc proportional to the supply shortage with respect to the required flow rate Qr, and the flow is redistributed. Occurs (t1).

In this case as well, the fine adjustment by the load-dependent characteristic of the pressure compensating valve 12 of the swing section allows the swing motor 2 to be adjusted.
Accelerates gently without causing hunting as occurs in the conventional LS control.

When the degree of saturation is large, the target compensation differential pressure ΔPv is greatly reduced. Furthermore,
Regarding the pressure compensating valve 12 in the turning section, the load pressure PL of the turning motor 2 increases due to the inertial load at the same time as the turning motor 2 is started. is there. This decrease in the target compensation differential pressure ΔPv of the pressure compensating valve 12 is limited by the set differential pressure Pb of the lower limit setting spring 55. Therefore, the target compensation differential pressure ΔPv of the pressure compensation valve 13 of the boom section
Is further reduced by the limited amount of decrease in the target compensation differential pressure ΔPv on the turning side.

As a result, the discharge flow rate of the hydraulic pump 1 is preferentially supplied to the swing motor 2 to some extent, and the swing operability can be maintained without the swing speed becoming extremely slow as compared with the boom cylinder 3. You.

As a comparative example, in a combined operation that does not include turning, the target compensation differential pressure ΔPv decreases to the same value due to a decrease in the LS control differential pressure ΔPc due to saturation, as shown by the broken line in FIG. 9, and the supply flow rate Qa also decreases. It decreases to the same value (assuming that the opening areas of the directional control valves related to the combined operation are the same).

In a combined operation including turning when the pressure compensating valve 12 of the turning section does not have the lower limit setting spring 55 (in the case of Japanese Patent Application Laid-Open No. Hei 10-89304), the reduction of the LS control differential pressure ΔPc due to saturation and the pressure compensating valve 12 The target compensation differential pressure ΔPv decreases extremely as shown by the two-dot chain line in FIG. 9, and the supply flow rate Qa also decreases extremely due to the load-dependent characteristics. In the present embodiment, the decrease in the target compensation differential pressure ΔPv of the pressure compensating valve 12 is limited by the set differential pressure Pb of the lower limit setting spring 55. Therefore, as compared with the case where the spring 55 is not provided, the reduction of the target compensation differential pressure ΔPv and the supply flow rate Qa of the turning section is suppressed by ΔΔPv2 and ΔQa2. With this function, the turning operability can be maintained without the turning speed becoming extremely slow as compared with other actuators during the combined operation.

FIG. 10 shows a case where the degree of saturation of the discharge flow rate of the hydraulic pump 1 in the combined operation is small. When the degree of saturation is small, the decrease in the target compensation differential pressure ΔPv of the pressure compensation valve 13 of the boom section remains at or above the set differential pressure Pb of the lower limit setting spring 55 of the pressure compensation valve 12 of the turning section. Due to the load dependency of the pressure compensating valve 12 of the turning section, the target compensation differential pressure ΔPv of the turning section decreases to the set differential pressure Pb of the lower limit setting spring 55.

As the turning speed increases, the turning motor 2
, The target compensation differential pressure ΔPv of the pressure compensation valve 12 in the turning section increases. Eventually, the target compensation differential pressure ΔPv and the supply flow rate Qa are the same for both the turning and the boom sections (assuming that the opening areas of the directional control valves of the turning and the boom sections are the same) (t4).

In the case where the lower limit setting spring 55 is not provided in the pressure compensating valve 12 of the turning section (in the case of Japanese Patent Laid-Open No. Hei 10-89304), as shown by a two-dot chain line in FIG. The compensation differential pressure ΔPv decreases to a pressure lower than Pb, and the supply flow rate Qa to the swing motor 2 is reduced.
However, it drops significantly immediately after startup. By providing the lower limit setting spring 55, ΔΔP
The reduction of the target compensation differential pressure ΔPv of the turning section and the supply flow rate Qa are suppressed by v3 and ΔQa3. Therefore, also in this case, the turning operability can be maintained without extremely lowering the turning speed as compared with other actuators.

As described above, according to the present embodiment, due to the load-dependent characteristics of the pressure compensating valve 12 of the turning section, the turning operation can be performed without any jerky feeling and the steady operation can be achieved regardless of whether the turning operation is started alone or in combination. Can transition to the state. Further, a lower limit setting spring 55 is provided in the pressure compensating valve 12 of the turning section, and pressure oil is preferentially supplied to the turning motor 2 when the discharge flow rate of the hydraulic pump 1 is saturated.
The change in turning speed during the transition from the swing-only operation to the swing-combined operation is suppressed, and the same applies to the transition from the reverse swing-combination to the swing-only operation. Acceleration can be achieved without being extremely slow, and excellent turning operability and turning independence can be secured. Further, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.

FIGS. 11 to 14 show a second embodiment of the present invention.
This will be described below. In the figure, the same reference numerals are given to the member loads equivalent to the members shown in FIGS. 1 and 2. In the present embodiment, the turning priority spring always acts on the spool of the pressure compensating valve.

In FIG. 11, the pressure compensating valves 13 to 16 other than the swivel section are the same as those in the first embodiment.

The pressure compensating valve 12A in the turning section causes the pressure on the upstream side of the directional control valve 7A to act in the closing direction, and the detection lines 20 to 2, which are the pressures on the downstream side of the directional control valve 7A.
4 is applied in the opening direction, the maximum load pressure derived from the signal line 37 is applied in the closing direction, and the discharge pressure of the hydraulic pump 1 is applied in the opening direction.
As a result, the LS control differential pressure (LS controlled hydraulic pump 1
The differential pressure between the directional control valve 7A and the directional control valve 7A is controlled using the differential pressure between the discharge pressure and the maximum load pressure as the target compensation differential pressure. And has a load-dependent characteristic of reducing the target compensation differential pressure so as to limit the flow rate of the pressure oil passing through the pressure compensating valve. This point is also the same as the pressure compensating valve 12 of the first embodiment.

The pressure compensating valve 12A has a turning priority spring 55A on the opening direction acting side which is the setting side of the target compensation differential pressure, and the turning priority spring 55A constantly performs pressure compensation during the operation of the pressure compensating valve 12A. This acts on the spool of the valve 12A to set a fixed auxiliary target compensation differential pressure for turning priority which is added to the target compensation differential pressure due to the LS control differential pressure.
That is, the target compensation differential pressure of the pressure compensating valve 12A is larger than that of the pressure compensating valves 13 to 16 other than the turning section by the turning priority spring 5.
It is increased by the amount set by 5A.

Also, the directional control valve 7A of the swivel section
Corresponds to the setting of the target compensation differential pressure which is larger than that of the pressure compensating valve 12A, so that the flow rate characteristic as designed can be obtained when the discharge flow rate of the hydraulic pump 1 is not in the saturation state. The opening area of 57b is set smaller than usual.

FIG. 12 shows the relationship. In the drawing, M1 is a change in the opening area (opening area characteristic) of the meter-in variable throttles 57a and 57b with respect to the spool stroke of the direction switching valve 7A, and M2 is a swing priority spring 55A for the pressure compensating valve.
4 shows a change in the opening area (opening area characteristic) of the meter-in variable throttle with respect to the spool stroke of the directional switching valve (for example, the directional switching valve 7 in the first embodiment shown in FIG. 1) under rated conditions, without using. M1 is set to have a larger opening area for the same spool stroke than M2.

FIG. 13 shows the structure of the pressure compensating valve 12A.
In FIG. 13, a small-diameter hole 321 having an end face 320 is formed in the first body 301a.
The first spool 311, the second spool 312, and the third spool 310 are closed between the first spool 311 fitted into the small-diameter hole 321 and the end surface 320 of the small-diameter hole 321 in the oil chamber 331 </ b> A of the first end surface 320. The above-mentioned turning priority spring 55A which pushes in the direction is arranged. Oil chambers 331A, 33
The relationship between the pressure receiving areas B1, B3, B1, and B2 in the pressure chambers 2, 334, 336 is the same as the oil chambers 331, 33 shown in FIG.
The relationship between the pressure receiving areas B1, B3, B1, and B2 in 2,334,336 is the same. The other configuration of the pressure compensating valve 12A is the same as that of the first embodiment shown in FIG.

Turning priority spring 5 in pressure compensating valve 12A
The operation principle of 5A will be described.

The lower limit setting spring 55 in the pressure compensating valve 12 of the first embodiment sets the lower limit of the target compensation differential pressure so that the target compensation differential pressure does not become smaller than a predetermined value. Assuming that the lower limit value of the target compensation differential pressure is the aforementioned Pb, in the present embodiment, the turning priority spring 55A is always applied to the spool, and the target compensation differential pressure corresponding to the lower limit value Pb is set to the target compensation by the LS control differential pressure. Set to be added to the differential pressure. As a result, the target compensation differential pressure of the pressure compensating valve 12A becomes larger than the other pressure compensating valves 13 to 16 by Pb. That is, the target compensation differential pressure of the pressure compensating valves 13 to 16: Ps−PLmax The target compensation differential pressure of the pressure compensating valve 12A: Ps−PLmax +
Pb When the target compensation differential pressure of the pressure compensating valve 12A is set as described above, if the opening area of the meter-in variable throttle of the directional control valve of the turning section is set to the same size as before, only the turning motor 2 has Pb amount. Will flow more. Therefore, it is necessary to reduce the opening area of the meter-in variable throttle of the directional control valve of the turning section by Pb so that the same flow rate as before flows through the turning motor 2.

That is, the opening area of the directional control valve for turning at the target compensation differential pressure under the original rated condition is As, and the directional control valve 7
Assuming that the aperture area of the meter-in variable diaphragm of A is Aso, Aso = As√ ((Ps−PLmax) / (Ps−PLmax + P
b))

A change in the flow rate supplied to the swing motor 2 during saturation when such a pressure compensating valve 12A and directional switching valve 7A are used is compared with other actuators. The opening area of the directional control valve relating to the other actuator is set to As which is the same as the opening area of the directional control valve for turning at the target compensation differential pressure under the rated condition, the flow rate supplied to the turning motor 2 is set to Qa, If the supply flow rate is Qb,
Qa and Qb can be expressed as follows.

Qb = c × As√ ((2 / ρ) (Ps−PLmax)) = c × As√ ((2 / ρ) ΔPc) Qa = c × Aso × √ ((2 / ρ) (ΔPc + Pb) ) = C × As√ ((Ps−PLmax) / (Ps−PLmax + P
b)) × √ ((2 / ρ) (ΔPc + Pb)) where As√ ((Ps−PLmax) / (Ps−PLmax + P
b)) is a value (constant) under rated conditions.

The rating conditions are set as follows.

Ps-PLmax = 15 kgf / cm2 Pb = 3 kgf / cm2 Qa = Qb = 85 (liter / min) Therefore, {((Ps-PLmax) / (Ps-PLmax + Pb)) =}
(15 / (15 + 3)) ≒ 0.91 c × As√ (2 / ρ) = Q / √ΔPc ≒ 21.94 These values are substituted into the above equations of Qb and Qa.

Qb = 21.94√ΔPc Qa = 21.94 × 0.91√ (ΔPc + Pb) FIG. 14 shows the relationship between the above Qa, Qb and the LS control differential pressure ΔPc. As can be seen from this figure, when the LS control differential pressure ΔPc becomes 15 kgf / cm 2 or less, that is, when the discharge flow rate of the hydraulic pump 1 is in a saturation state where the required flow rate is less than the required flow rate, the supply flow rate Qa of the swing motor 2 is changed to an actuator other than the swing. , And the pressure oil is supplied to the swing motor 2 preferentially. The degree of priority (difference in flow rate) increases as the LS control differential pressure ΔPc decreases.

In the above, the oil chambers 13a to 16a and 13b to 16b connected to the signal lines 52b to 52e and 53b to 53e of the pressure compensating valves 13 to 16 are connected to the swing motor 2 among the plurality of pressure compensating valves 12 to 16. The first means is provided in the pressure compensating valves 13 to 16 other than the relevant swivel section and sets a differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as a target compensation differential pressure. , Signal lines 52a and 53a of the pressure compensating valve 12A
Chamber 336 (pressure receiving area B2 = B1) and oil chamber 3 connected to
31A (pressure receiving area B1) is provided on the pressure compensating valve 12 of the swivel section and constitutes second means for setting the target compensation differential pressure, and the signal lines 50a and 50a of the pressure compensating valve 12A.
The oil chamber 334 (pressure receiving area B1> B3) and the oil chamber 332 (pressure receiving area B3) connected to 1a are provided with a plurality of pressure compensating valves 12 to
16, at least the pressure compensating valve 1 of the swivel section
2A, when the load pressure of the swing motor 2 increases,
The third means for reducing the target compensation differential pressure set by the second means and giving the load compensating valve 12A of the turning section a load-dependent characteristic is provided. The turning priority spring 55A of the pressure compensating valve 12A And a fourth means for setting a lower limit of the target compensation differential pressure set by the second means and corrected by the third means.

In the present embodiment, the second means (oil chambers 331A, 336) is connected to the first means (oil chambers 13a to 13a).
16a, 13b to 16b), a means for setting the differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as the target compensation differential pressure. ) Is the second means (oil chamber 331A,
336) functions as a lower limit setting means for both the reduction of the target compensation differential pressure itself and the reduction of the target compensation differential pressure due to the load-dependent characteristic given by the third means (oil chambers 332, 334).

Further, the fourth means (turning priority spring 55)
Is a biasing means which constantly adds an auxiliary value to the target compensation differential pressure set by the second means (oil chambers 331A, 336) and corrected by the third means (oil chambers 332, 334). The direction switching valve 7A of the section has direction opening valves 8 to 11 other than the swivel section in which the opening areas of the meter-in variable throttles 57a and 57b correspond to the target compensation pressure of an auxiliary value added by the urging means. Is configured to be smaller than the opening area.

Therefore, also in the present embodiment, the load-dependent characteristic of the pressure compensating valve 12A of the swivel section causes
At the start of either the turning alone or the compound turning, it is possible to accelerate and shift to the steady state without the jerky feeling of the turning operation. In addition, since the turning priority spring 55A is provided in the pressure compensating valve 12A of the turning section to supply the pressurized oil to the turning motor 2 preferentially at the time of the saturation of the discharge flow rate of the hydraulic pump 1, the turning single operation to the turning combined operation are performed. The change in turning speed during the transition to is suppressed, and the same applies to the transition from the reverse turning composite to the turning alone operation.
As compared with other actuators, the turning speed can be accelerated without being extremely slow, and excellent turning operability and turning independence can be secured. Further, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.

FIGS. 15 and 1 show a third embodiment of the present invention.
6 will be described. In the figure, the same reference numerals are given to the member loads equivalent to the members shown in FIGS. 1 and 2. In this embodiment,
The turning priority is given to the pressure compensating valve of the turning section without setting the target compensation differential pressure by the LS control differential pressure.

In FIG. 15, the pressure compensating valves 13 to 16 other than the swivel section are the same as those in the first embodiment.

The pressure compensating valve 12B of the swivel section
Makes the pressure on the upstream side of the direction switching valve 7 act in the closing direction, and the detection line 20 which is the pressure on the downstream side of the direction switching valve 7
When the load pressure of the swing motor 2 increases when the pressure (load pressure of the swing motor 2) is applied in the opening direction, the target compensation differential pressure is reduced so as to limit the flow rate of the pressure oil passing through the pressure compensation valve 12B. This is the same as the pressure compensating valve 12 of the first embodiment.

The pressure compensating valve 12B has means for setting a normal target compensation differential pressure, for example, a setting spring 60, on the opening direction acting side which is the setting side of the target compensation differential pressure. The target compensation differential pressure having the same magnitude as the target compensation differential pressure by the LS control differential pressure when the discharge flow rate of the pump 1 is not in the saturation state is set.
That is. When the discharge flow rate of the hydraulic pump 1 is in the saturation state, the pressure compensation valves 13 to 16 other than the swivel section that sets the target compensation differential pressure by the LS control differential pressure become smaller in accordance with the degree of saturation. On the other hand, in the pressure compensating valve 12B of the turning section, the target compensation differential pressure set by the setting spring 60 is substantially unchanged even in the saturation state, and the target compensation differential pressure changes due to the load-dependent characteristic. .

Further, the pressure compensating valve 12B is provided with a lower limit setting spring 55 for setting the lower limit of the target compensation differential pressure of the pressure compensating valve 12B as in the first embodiment.

FIG. 16 shows the structure of the pressure compensating valve 12B.
16, oil chambers 331 and 336 in the first embodiment shown in FIG. 2 are oil chambers 331B and 336B, respectively.
These oil chambers 331B and 336B communicate with tanks through tank ports 341B and 346B, respectively, and are provided with oil chambers 33 provided by the first spool 311.
1B and the pressure receiving area B2 of the oil chamber 336B given by the step between the second large-diameter portion 314 and the small-diameter portion 325 of the third spool 310 respectively.
It is configured such that no hydraulic pressure acts on the eleventh and third spools 310. A spring 60 for setting the above-mentioned target compensation differential pressure is disposed in a concave portion 311a formed on the end surface of the first spool 311 instead of the weak spring 350 for holding the initial position. The relationship between the pressure receiving areas B3 and B1 located in the oil chambers 332 and 334 is the same as that of the first embodiment (B1> B3), whereby the swing motor 2 is turned on in response to an increase in the load pressure (PL) of the swing motor 2. Has a load-dependent characteristic that reduces the flow rate of the directional control valve 7 that leads to.

In the above, the oil chambers 13a to 16a and 13b to 16b connected to the signal lines 52b to 52e and 53b to 53e of the pressure compensating valves 13 to 16 are connected to the swing motor 2 among the plurality of pressure compensating valves 12 to 16. The first means is provided in the pressure compensating valves 13 to 16 other than the relevant swivel section and sets a differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as a target compensation differential pressure. The setting spring 60 of the pressure compensating valve 12B is provided on the pressure compensating valve 12B of the turning section, constitutes a second means for setting the target compensation differential pressure, and is connected to the signal lines 50a and 51a of the pressure compensating valve 12B. The chamber 334 (pressure receiving area B1> B3) and the oil chamber 332 (pressure receiving area B3) are provided with at least one of the pressure compensating valves 12 to 16 in the swivel section. B, a third means for reducing the target compensation differential pressure set by the second means when the load pressure of the swing motor 2 rises and for giving the load compensating valve 12B of the swing section a load-dependent characteristic. The lower limit setting spring 55 of the pressure compensating valve 12 is provided on the pressure compensating valve 12 of the turning section, and is set by the second means.
A fourth means for setting the lower limit of the target compensation differential pressure corrected by the means is constituted.

In the present embodiment, the second means (setting spring 60) sets a value which does not change due to the difference between the discharge pressure of the hydraulic pump 11 and the maximum load pressure of the plurality of actuators 2 to 6 to the target compensation difference. Means to set as pressure
The fourth means (lower limit setting spring 55) functions as a lower limit setting means for a decrease in the target compensation differential pressure due to the load-dependent characteristic given by the third means (oil chambers 332, 334).

Further, the fourth means (lower limit setting spring 55)
When the target compensation differential pressure set by the second means (setting spring 60) and corrected by the third means (oil chambers 332, 334) reaches a predetermined value, the spool 311 of the pressure compensating valve 12B of the swivel section becomes This is an urging means for applying an urging force in the opening direction.

In the present embodiment configured as described above, the setting spring 60 sets the target compensation differential pressure having the same magnitude as the target compensation differential pressure due to the LS control differential pressure when the discharge flow rate of the hydraulic pump 1 is not in the saturation state. Since the pressure is set, before the discharge flow rate of the hydraulic pump 1 saturates, the discharge flow rate of the hydraulic pump 1 is distributed according to the ratio of the required flow rates of the plurality of actuators as in the first embodiment. When the discharge flow rate of the hydraulic pump 1 is in a saturation state, the target compensation differential pressure is set so as to satisfy the target compensation differential pressure, and the target compensation differential pressure is corrected by the load-dependent characteristic of the pressure compensating valve 12B in the turning section. The target compensation differential pressure of the pressure compensating valves 13 to 16 decreases in accordance with the decrease in the LS control differential pressure, whereas the target compensation differential pressure decreases. Setting of the pressure compensating valve 12B spring 60
Does not change depending on the degree of saturation, the target compensation differential pressure of the pressure compensating valve 12B changes only by the load-dependent characteristic, and a lower limit is set for a decrease in the target compensation differential pressure due to the load-dependent characteristic. The spring 55 functions, and also in this case, similarly to the first and second embodiments, the pressurized oil is supplied to the swing motor 2 preferentially.

Therefore, according to the present embodiment, the load-dependent characteristic of the pressure compensating valve 12B of the swivel section is
At the start of either the turning alone or the compound turning, it is possible to accelerate and shift to the steady state without the jerky feeling of the turning operation. Also, a lower limit setting spring 55 and a setting spring 60 are provided in the pressure compensating valve 12B of the turning section, and the pressure is preferentially applied to the turning motor 2 at the time of the saturation of the discharge flow rate of the hydraulic pump 1 and at the time of reduction of the target compensation differential pressure due to the load-dependent characteristic. Since the oil is supplied, the change in the turning speed at the time of transition from the single swing operation to the combined swing operation is suppressed, and the same applies to the transition from the reverse combined swing to the single swing operation, and also at the time of the start of the combined swing operation. As compared with other actuators, the turning speed can be accelerated without being extremely reduced, and excellent turning operability and turning independence can be secured. Further, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.

In the above embodiment, an example is shown in which a before orifice type pressure compensating valve located upstream of the direction switching valve is used. However, an after orifice type pressure compensating valve located downstream of the direction switching valve is used. Even if a valve is used, a system having the same effect can be configured.

In the above embodiment, the lower limit setting spring 55 and the turning priority spring 55 serve as means for controlling the target compensation differential pressure so that the pressure compensating valve in the turning section has priority.
A. Although the setting spring 60 is provided, an oil chamber for guiding the control pressure may be provided in the same manner as the oil chamber for guiding the pressure upstream and downstream of the direction switching valve to apply a hydraulic control force. in this case,
By changing the control pressure according to the purpose, more complicated and advantageous control can be performed.

Further, in the above embodiment, the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators is set as the target compensation differential pressure. The differential pressure generating valve that generates a secondary pressure corresponding to the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators is provided, and the output pressure is adjusted. The pressure compensating valve may be guided to the end in the opening direction of the spool.

[0128]

According to the present invention, in the hydraulic drive system including the turning control system, the turning operation can be accelerated without a jerky feeling and can be shifted to a steady state, regardless of whether the turning operation is started alone or in combination. Excellent turning operation because the change in turning speed during the transition from turning only operation to turning combined operation or vice versa can be suppressed, and the turning speed can be accelerated without extremely slowing down compared to other actuators when starting up the compound, and excellent turning operation In addition to ensuring the independence and turning independence, it is possible to provide a system that does not cause a problem of an increase in cost and space and a complicated circuit configuration due to provision of another circuit.

[Brief description of the drawings]

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

FIG. 2 is a cross-sectional view showing details of a structure of a pressure compensating valve in a swivel section.

FIG. 3 is a diagram showing load-dependent characteristics of a pressure compensating valve in a swivel section.

FIG. 4 is a diagram showing a lower limit setting function of a target compensation differential pressure by a turning priority spring in a pressure compensating valve of a turning section.

FIG. 5 is a diagram showing an external appearance of a hydraulic shovel using the hydraulic drive device of the present invention.

FIG. 6 is a time chart showing a change in a target compensation differential pressure of a pressure compensating valve in a turning section during a single turning operation.

FIG. 7 is a time chart for explaining the operation of the pressure compensating valve in the turning section when the degree of saturation is large when another actuator is activated during the turning steady rotation.

FIG. 8 is a time chart for explaining the operation of the pressure compensating valve in the turning section when the degree of saturation is small when another actuator is started during the turning steady rotation.

FIG. 9 is a time chart for explaining the operation of the pressure compensating valve in the turning section when the degree of saturation is large when the turning and other actuators are simultaneously activated.

FIG. 10 is a time chart for explaining the operation of the pressure compensating valve in the turning section when the degree of saturation is small when turning and simultaneous activation of another actuator are performed.

FIG. 11 is a circuit diagram showing a hydraulic drive device according to a second embodiment of the present invention.

FIG. 12 is a diagram showing an opening area characteristic of a directional control valve of a swivel section.

FIG. 13 is a cross-sectional view showing details of the structure of the pressure compensating valve in the swivel section.

FIG. 14 is a diagram illustrating a priority characteristic of a flow rate of a swirling section in a saturation state.

FIG. 15 is a circuit diagram showing a hydraulic drive device according to a third embodiment of the present invention.

FIG. 16 is a cross-sectional view showing details of the structure of the pressure compensating valve in the swivel section.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Hydraulic pump 2-6 Actuator (2: turning motor) 7-11 Direction switching valve 12-16 Pressure compensation valve 13a-16a Oil chamber (1st means) 13b-16b Oil chamber (1st means) 18 Pump control device 20 -24 detection line 25-29 signal line 34-36 signal line 37 signal line 40 tilt control actuator 41 load sensing control valve 50a-50e signal line 51a-51e signal line 52,52a-52e signal line 53,53a-52e signal Line 55 Lower limit setting spring (fourth means; biasing means) 60 Setting spring (second means) 200 Lower traveling body 201 Revolving body 202 Front work machine 331 Oil chamber (pressure receiving area B1) (second means) 332 Oil chamber ( Pressure receiving area B3) (third means) 334 Oil chamber (pressure receiving area B1> B3) (third means) 336 oil Chamber (pressure receiving area B2 = B1) (second means) 7A Direction switching valve 12A pressure compensating valve 55A turning priority spring (fourth means; biasing means) 57a, 75b meter-in variable throttle 331A oil chamber (pressure receiving area B1) ( 2nd means) 12B pressure compensation valve

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[Procedure amendment]

[Submission date] December 14, 1999 (1999.12.
14)

[Procedure amendment 1]

[Document name to be amended] Statement

[Correction target item name] 0108

[Correction method] Change

[Correction contents]

Ps−PLmax = 15 kgf / cm ² Pb = 3 kgf / cm ² Qa = Qb = 85 (liter / min) Therefore, {((Ps−PLmax) / (Ps−PLmax + Pb)) =}
(15 / (15 + 3)) ≒ 0.91 c × As√ (2 / ρ) = Q / √ΔPc ≒ 21.94 These values are substituted into the above equations of Qb and Qa.

 ──────────────────────────────────────────────────の Continuing on the front page (72) Inventor Takashi Kanai 650, Kandamachi, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. Inside the Tsuchiura Plant of Shikiura Company (72) Kenichiro Nakaya 650, Kandamachi, Tsuchiura-shi, Ibaraki Prefecture Hitachi Construction Machinery Co., Ltd. Inside the Tsuchiura Plant (72) Inventor Satoshi Hamamoto 1-1-1, Fujikoshi Honcho, Toyama City, Toyama Prefecture Fujiuchi Co., Ltd. (72) Inventor Koji Okazaki 1-1-1, Fujikoshi Honcho, Toyama City, Toyama Prefecture 72) Inventor Yukiaki Nagao 1-1-1 Fujikoshi Honcho, Toyama City, Toyama Pref.

Claims (7)

[Claims] 1. A hydraulic pump, a plurality of actuators including a swing motor driven by hydraulic oil discharged from the hydraulic pump, and a flow rate of hydraulic oil supplied from the hydraulic pump to the plurality of actuators. A plurality of directional control valves, a plurality of pressure compensating valves respectively controlling a pressure difference between the front and rear of the plurality of directional control valves, and a discharge pressure of the hydraulic pump is higher than a maximum load pressure of the plurality of actuators by a predetermined value. And a pump control means for load sensing control for controlling a pump discharge flow rate. The hydraulic drive device further comprising: a pressure compensating valve provided in a pressure compensating valve of the plurality of pressure compensating valves other than a turning section related to the turning motor; First means for setting a differential pressure between a discharge pressure of a pump and a maximum load pressure of the plurality of actuators as a target compensation differential pressure; A second means provided on the pressure compensating valve of the turning section to set a target compensation differential pressure; and a load of the turning motor provided on at least the pressure compensating valve of the turning section among the plurality of pressure compensating valves. When the pressure rises, a third means for reducing the target compensation differential pressure set by the second means and giving the pressure compensation valve of the swivel section a load-dependent characteristic, and a pressure compensation valve of the swirl section, The second
And a fourth means for setting a lower limit of the target compensation differential pressure set by the means and corrected by the third means. 2. The hydraulic drive device according to claim 1, wherein the second means, like the first means, sets a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators to the target pressure. Means for setting as a compensation differential pressure, wherein the fourth means reduces the target compensation differential pressure itself set by the second means and decreases the target compensation differential pressure by the load-dependent characteristic given by the third means. A hydraulic drive device functioning as a lower limit setting means for both of them. 3. The hydraulic drive device according to claim 1, wherein the second means sets a value which does not change due to a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators. The hydraulic drive device according to claim 4, wherein the fourth means functions as a lower limit setting means for a decrease in the target compensation differential pressure due to the load-dependent characteristic given by the third means. 4. The hydraulic drive device according to claim 1, wherein said fourth means is set by said second means, and said target compensation differential pressure corrected by said third means is a predetermined value. A hydraulic drive device comprising: urging means for applying an urging force in the opening direction to the spool of the pressure compensating valve of the turning section when the value reaches the value. 5. The hydraulic drive device according to claim 4, wherein said urging means is set by said second means, and said turning is performed when a target compensation differential pressure corrected by said third means reaches a predetermined value. A hydraulic drive device, which is a lower limit setting spring that acts on a spool of a pressure compensation valve of a section and urges the spool in an opening direction. 6. The hydraulic drive device according to claim 1, wherein the fourth means always sets an auxiliary value to the target compensation differential pressure set by the second means and corrected by the third means. The directional control valve of the turning section is provided with an opening area of the meter-in variable throttle, the opening area of the meter-in variable throttle being equal to a target compensation pressure of an auxiliary value added by the pressing means, and being provided in a portion other than the turning section. A hydraulic drive device configured to be smaller than an opening area of a direction switching valve. 7. The hydraulic drive device according to claim 6, wherein the biasing means is a turning priority spring that always acts in a direction in which a spool of a pressure compensating valve of the turning section opens. .