JPH0231004A - Hydraulic drive device - Google Patents

Abstract

PURPOSE:To prevent the generation of shocks by limiting the drive of a diversion compensating valve corresponding to a low pressure actuator in the closing direction when the actuator is transferred from the single drive to the composite one. CONSTITUTION:A hydraulic drive device is provided with a differential pressure detector 53 for detecting differential pressure between pump pressure and the maximum load pressure of an actuator to be output as a signal. In the other drive section 38b of a diversion compensating valve 38 corresponding to a low pressure actuator according to the signal from the differential pressure detector 53 is provided a limiting means 52 provided with a controlling force adding means 54 to give a controlling force for restricting the drive of the diversion compensating valve 38 in the closing direction. Hence when the actuator is transferred from the single drive to the double-acting drive, the fluctuation of flow supplied to the low pressure actuator can be restrained, so that the generation of large shocks can be prevented to improve the maneuverability.

Description

[Detailed Description of the Invention] <Industrial Application Field> The present invention is directed to a high-pressure actuator and a low-pressure actuator that are provided in correspondence with a plurality of branch compensating valves for discharging the pressure oil of one main hydraulic pump through a plurality of branch compensating valves. The present invention relates to a hydraulic drive device capable of supplying a divided flow to each of a plurality of actuators including a plurality of actuators, and driving these actuators in a combined manner to perform a desired combined operation.

<Prior Art> FIG. 14 is a circuit diagram showing a hydraulic drive system for a hydraulic excavator as an example of this type of conventional hydraulic drive system.

The hydraulic drive device shown in FIG. 14 includes a prime mover 1, a variable displacement hydraulic pump, that is, a main hydraulic pump 2, which is driven by the prime mover 1, and a turning movement (not shown) that is driven by pressure oil discharged from the main hydraulic pump 2. It is equipped with a turning motor 3 that turns the body, and an actuator including a boom cylinder 4 that turns a boom (not shown). In this case, driving the swing motor 3 requires a large drive pressure, but compared to this, the drive pressure of the boom cylinder 4 is relatively small. Therefore, the swing motor 3 is a high-pressure actuator, and the boom cylinder 4 is a high-pressure actuator. It becomes a low pressure actuator.

In addition, there is also a flow control valve that controls the flow of pressure oil supplied from the main hydraulic pump 2 to the swing motor 3, that is, a swing direction control valve 5, and a flow branch compensation that controls the differential pressure across the swing direction control valve 5. A valve 6, a flow control valve that controls the flow of pressure oil supplied from the main hydraulic pump 2 to the boom cylinder 4, that is, a boom directional control valve 7, and a pressure difference across the boom directional control valve 7. A branch compensating valve 8 is provided.

A control force Fal based on the pressure on the upstream side of the branch flow compensation valve 6 and the load pressure is applied to one drive part 6a of the branch flow compensation valve 6 so that the branch flow compensation valve 6 opens, and the other drive part 6b
The pressure on the downstream side of this branch compensation valve 6 and the shuttle valve 9 are
．． A control force Fa2 based on the maximum load pressure of the circuit led through the circuit 10 is applied to close the branch compensating valve 6, and similarly, one drive section 8a of the branch compensating valve 8 Control force Fb due to upstream pressure and load pressure
1 is applied to open the branch compensating valve 8, and a control force Fb2 due to the downstream pressure of the branch compensating valve 8 and the maximum load pressure of the circuit is applied to the other drive unit 8b to open the branch compensating valve 8. is given to close. The displacement of the main hydraulic pump 2 is controlled by a control actuator 12 driven by a flow rate regulating valve 11 that is switched according to the discharge pressure of the main hydraulic pump 2 and the maximum load pressure of the circuit. For example, swing motors 3 having different driving pressures
When the combined drive of the boom cylinder 4 and the shunt compensation valve 6
．． 8, the pressure difference between the front and rear of the swing direction control valve 5 and the boom direction control valve 7 can be maintained equally, thereby diverting the pressure oil discharged from the main hydraulic pump 2 to the swing motor 3.
It can also be supplied to the boom cylinder 4, making it possible to realize combined operations such as turning and raising the boom.

<Problems to be Solved by the Invention> By the way, in this conventional hydraulic drive system, when the boom cylinder 4, which is a low-pressure actuator, is being driven, the swing direction control valve is not operated with the intention of combined operation with the boom. 5
When switched, the high-pressure oil supplied to drive the swing motor 3, which is a high-pressure actuator, is also guided to the drive part 8b of the branch compensation valve 8 of the boom cylinder 4 via the shuttle valve 9.10. , actuate this branch compensating valve 8 in the closing direction. That is, at the initial stage of transition from the boom-only drive to the combined operation of the boom and swing as described above, the phenomenon in which the shunt compensating valve 8 is temporarily completely closed and then the shunting compensating valve 8 begins to open is prevented. As a result, fluctuations in the flow rate of the pressure oil supplied to the boom cylinder 4 via the diversion compensation valve 8 and the boom directional control valve 7 become large. This may cause a large shock and reduce operability.

The present invention has been made in view of the above-mentioned actual situation in the prior art, and its purpose is to reduce the flow rate supplied to the low-pressure actuator when transitioning from the single drive of the low-pressure actuator to the combined drive of the low-pressure actuator and the high-pressure actuator. An object of the present invention is to provide a hydraulic drive device that can suppress fluctuations.

Means for Solving the Problems To achieve this object, the present invention provides a main hydraulic pump and a plurality of actuators including a high-pressure actuator and a low-pressure actuator driven by pressure oil supplied from the main hydraulic pump. , a flow rate control valve that controls the flow of pressure oil supplied to these actuators, a branch flow compensation valve that controls the differential pressure across these flow rate control valves, and a flow rate discharged from the main hydraulic pump. In a hydraulic drive device that is equipped with a flow rate control valve, supplies pressure oil from the main hydraulic pump to each actuator through a flow compensation valve and a flow rate control valve, and is capable of combined operation of these actuators. The structure includes a restriction means for restricting the driving of the branch compensating valve corresponding to the low-pressure actuator in the closing direction when transitioning from independent driving to combined driving of the actuator and the high-pressure actuator.

Effects> Since the present invention is configured as described above, when the low pressure actuator is moved from a single drive state to a combined drive of the low pressure actuator and the high pressure actuator, the flow dividing compensation valve related to the low pressure actuator is controlled by the limiting means. The drive in the closing direction of the low pressure actuator is limited so that it does not completely close, and therefore less flow can be supplied to the low pressure actuator during the transition phase than is necessary to obtain a predetermined minimum speed of the low pressure actuator; Fluctuations in the flow rate supplied to the low pressure actuator can be suppressed.

<Embodiment> FIG. 1 is a circuit diagram showing a first embodiment of the hydraulic drive device of the present invention. This first embodiment is applied to a hydraulic excavator, and includes a prime mover 21, one variable displacement hydraulic pump driven by the prime mover 21, that is, a main hydraulic pump 22, and pressure oil discharged from the main hydraulic pump 22. A plurality of actuators driven by, ie, a swing motor 23, a left travel motor 24, and a right travel motor 25,
It includes a boom cylinder 26, an arm cylinder 27, and a packet cylinder 28. In addition, the swing motor 23
drives a revolving body (not shown), a left traveling motor 24 and a right traveling motor 25 drive a crawler track (not shown), that is, a traveling body, and a boom cylinder 26, an arm cylinder 27, and a packet cylinder 28 respectively drive a boom, an arm, and a rotating body (not shown).
Drive the packet.

Also, a swing motor 23, a left travel motor 24, a right travel motor 25, a boom cylinder 26, an arm, a cylinder 27,
A flow control valve that controls the flow of pressure oil supplied to each of the packet cylinders 28, that is, a swing direction control valve 2
9. Direction control valve for right travel 30, Direction control valve for right travel 31
, boom directional control valve 32, arm directional control valve 33,
Packet directional control valve 34 and branch flow compensation valves 35, 36, 37, 38 provided corresponding to these flow rate control valves.
39.40.

Further, the displacement volume of the main hydraulic pump 22 described above is controlled by a control actuator 41, and the drive of this control actuator 41 is controlled by a flow rate adjustment valve 42. The flow rate adjustment valve 42 adjusts the pressure difference ΔP between the pump pressure guided through the pipe line 43 and the maximum load pressure guided through the pipe line 44.
Driven by LS. These control actuator 41 and flow rate adjustment valve 42 control the flow rate discharged from the main hydraulic pump 22 by adjusting the differential pressure Δ between the pump pressure and the maximum load pressure.
A flow rate control means is configured to control according to PL5.

One drive section 35a of the above-described branch compensation valves 35 to 40,
36a, 37a, 38a, 39a, and 40a, each force of a spring 45, 46, 47, 48, 49, 50 and a control force due to each load pressure causes these branch compensation valves 35 to 40 to open. The other driving parts 35b, 36b, 37b, 38b, 3.9b, 40b are given as follows.
Each of these flow compensation valves 35, 36, 37
．． 38, 39, and 40 and a control pressure, which will be described later, led through the pipe line 51, a control force is applied to close these branch compensating valves 35 to 40.

In this first embodiment, a swing motor 23, a left travel motor 24, a right travel motor 25, a boom cylinder 26
, arm cylinder 27, and packet cylinder 28, from a single drive state of a low-pressure actuator that operates at a relatively low drive pressure to a combined drive of this low-pressure actuator and a high-pressure actuator that operates at a relatively high drive pressure. A restricting means 52 is provided for restricting the driving in the closing direction of the branch compensating valve provided corresponding to the low pressure actuator at the time of transition. This limiting means 52 is controlled by a pressure difference Δ between the pressure of the pressure oil discharged from the main hydraulic pump 22, that is, the pump pressure, and the maximum load pressure of the actuator described above.
Differential pressure detection device 53 that detects PL5 and outputs it as a signal
Then, in response to the signal output from the differential pressure detection device 53, the other drive section 38 of the flow compensating valve 38 is installed corresponding to the low pressure actuator, for example, the boom cylinder 26 which becomes the low pressure actuator during combined drive of swing and boom.
b includes a control force adding means 54 for applying a control force to restrict the driving of the branch compensating valve 38 in the closing direction.

The control force adding means 54 is connected to the differential pressure detection device 53, for example, and includes an input section 55, a calculation section 56, a storage section 57 that stores in advance the functional relationship between the above-mentioned differential pressure ΔPLS and the control force F, and an output section 54. 58, and the other drive section 35b of the shunt compensation valves 35 to 40 according to control force signals output from the output section 58 of the five controllers.
- 40b. This control pressure generating means 60
is arranged, for example, in the pilot hydraulic pressure source 61 and the pipe line 51, that is, the pilot hydraulic pressure source 61 and the branch compensating valve 35.
- 40, and one electromagnetic valve 62 that is arranged between the other drive parts 35b to 40b and operates in response to a control force signal output from the output part 58 of the controller 59.

Then, in the storage section 57 of the controller 59 described above,
For example, the functional relationship shown in FIG. 3, ie, the relationship between the differential pressure ΔPLs between the pump pressure and the maximum load pressure and the control force F, is stored. Here, PLSX is the load sensing compensation differential pressure, and A is the minimum pressure so that the low-pressure actuator can be driven at the minimum necessary speed when transitioning from single drive of the low-pressure actuator to combined drive of the low-pressure actuator and high-pressure actuator. The differential pressure that makes it possible to supply a flow rate of , that is, the minimum flow rate compensation differential pressure, fc, is the minimum flow rate compensation control force corresponding to this minimum flow rate compensation differential pressure A, and f is the force of the spring that biases the branch compensation valve. .

As shown in FIG. 3, this functional relationship shows that when the differential pressure ΔPLS is larger than the minimum flow compensation differential pressure A, the control force F gradually decreases as the differential pressure ΔPLs increases, and the differential pressure When ΔPL5 becomes less than the minimum flow compensation differential pressure A, the differential pressure Δ
The relationship is such that a constant control force F, that is, a minimum flow compensation control force fc is output regardless of the decrease in PLs.

In the first embodiment configured as described above, by selectively operating any two or more of the direction control valves 29 to 34, the swing motor 23, the left travel motor 24, and the right travel motor are controlled. 25, the corresponding branch compensation valve 3 related to the corresponding actuator among the actuators such as the boom cylinder 26, the arm cylinder 27, and the packet cylinder 28.
Directional control valve 2 in which any one of 5 to 40 operates in the closing direction or in the opening direction and relates to the corresponding actuator.
The differential pressure across any one of the directional control valves 9 to 34 becomes the same regardless of changes in the load pressure of each actuator, and therefore the directional control valves 29 to 34 related to the corresponding actuators
A flow rate proportional to the opening ratio of each of these is supplied from the main hydraulic pump 22, and a combined operation of the actuating body related to the corresponding actuator, that is, the revolving body, traveling body, boom, arm, and packet (not shown) is performed. , turning, traveling, excavating earth and sand, etc.

At this time, as shown in step S1 in FIG. 2, the differential pressure ΔPL
A signal corresponding to s is input to the calculation unit 56 via the input unit 55 of the controller 59, and as shown in step III S2, the function relationship shown in FIG. 3 stored in the storage unit 57 is read out by the calculation unit 56. For example, during normal combined drive of the swing motor 23 and boom cylinder 26, the differential pressure ΔPLs is greater than the minimum flow rate compensation differential pressure A. Also, the process moves to step S3. In step S3, the calculation unit 56 calculates a control force F that is equal to or less than the minimum flow rate compensation control force fC that changes depending on the differential pressure Δpts. Next, the process moves to step S4, in which a control force signal corresponding to the control force F having a magnitude less than or equal to the control force fc is output from the output section 58 of the five controllers to the solenoid valve 6.
2 is output. As a result, the solenoid valve 62 is opened as appropriate, and the above-mentioned control force F output from the pilot hydraulic pressure source 61 is
The pilot pressure, that is, the control pressure corresponding to
8b, these branch compensating valves 35, 38 actuate in the closed direction. As a result, as described above, the pressure oil discharged from the main hydraulic pump 22 is divided and applied to the swing motor 23 and the boom cylinder 26, and the swing motor 2
3 and the boom cylinder 26 are combinedly driven.

For example, when the boom cylinder 26 is driven to operate the boom alone, when the boom cylinder 26 is driven and the swing motor 23 is simultaneously driven to move to a combined operation of the boom and swing, the high pressure The driving pressure of the swing motor 23 constituting the actuator becomes the maximum load pressure, and the differential pressure between the pump pressure and the maximum load pressure ΔPLs
A situation occurs in which the pressure decreases significantly and becomes equal to or less than the minimum flow compensation differential pressure A, but at this time, the judgment in step S2 in FIG. 2 described above is not satisfied and the process moves to step S5. In step S5, the calculation unit 56 calculates the differential pressure ΔPL from the functional relationship shown in FIG.
If S is less than the minimum flow compensation differential pressure A, the differential pressure ΔPL
, or even further decreases, a constant control force fc smaller than the spring force f is required.

A control force signal corresponding to the minimum flow rate compensation control force fc is applied from the output section 58 of the controller 59 to the solenoid valve 62, the solenoid valve 62 is driven, and the control pressure corresponding to the minimum flow rate compensation control force fc is applied to the branch flow compensation valve. 35. The other drive part 3 of 38
5b, given to 38b. Therefore, at this time, the minimum flow rate compensation control force fc is applied to the other driving part 38b of the branch flow compensation valve 38 related to the boom cylinder 26, and as a result, the branch flow compensation valve 38 is activated as the differential pressure ΔPLS significantly decreases. When the boom cylinder 26 is about to be completely closed, it is limited to a slightly open state, that is, an amount of restriction that can supply a flow rate that can secure the desired minimum speed of the boom cylinder 26.

In the first embodiment configured in this manner, when transitioning from single operation of the boom to combined operation of the boom and swing, that is, when transitioning from single operation of the low pressure actuator to combined operation of the low pressure actuator and high pressure actuator, the boom cylinder 26 Since the branch flow compensation valve 38 related to
It is possible to transition from a boom alone to a combination of a boom and swing without causing a stoppage of the flow rate supplied to the boom cylinder 26, and it is possible to suppress fluctuations in the flow rate supplied to the boom cylinder 26 before and after this transition,
Therefore, a large shock is not caused to the body of the hydraulic excavator equipped with this embodiment, and excellent operability can be obtained.

FIG. 4 is an explanatory diagram showing the main part of the second embodiment of the present invention.

This second embodiment differs from the first embodiment shown in FIG. 1 in the configuration of the flow rate control means for controlling the displacement of the main hydraulic pump 22.

The flow rate control means in this second embodiment is a hydraulic power source 63.
and a solenoid valve 64 connected to and connected between the head side and the rod side of the control actuator 41, and a solenoid valve 64 connected between the solenoid valve 64 and the tank and connected to the head side of the control actuator 41. It is connected to a differential pressure detection device 53 that detects a differential pressure ΔPLS between the pump pressure and the maximum load pressure, and has an input section 66, a calculation section 67, a storage section 68, and an output section 69. Control device 70
Contains.

In this flow rate control means, in the storage section 68 of the control device 70,
The differential pressure between the desired pump pressure and the maximum load pressure, that is, the differential pressure corresponding to the spring force of the spring that biases the flow rate regulating valve 42 shown in FIG. The value detected at 53 is compared in a calculation section 67, and a drive signal corresponding to the difference is determined by this calculation section 67. This drive signal is selectively sent from an output section 69 to the drive section of the electromagnetic valve 64, 65. is output to.

Here, if the differential pressure ΔPL detected by the differential pressure detection device 53 is
When s is larger than the set differential pressure, a signal is output from the control device 70 to the drive section of the solenoid valve 64, and this solenoid valve 64
is switched to the lower position, and pressure oil from the hydraulic source 63 is supplied to both the head side and the rod side of the control actuator 41. At this time, due to the difference in pressure receiving area between the head side and the rod side of the control actuator 41, the control actuator 4
The piston No. 1 moves to the left in the drawing, the displacement volume is changed so that the flow rate discharged from the main hydraulic pump 22 is reduced, and the differential pressure ΔPLs is controlled to be small so that it approaches the set differential pressure. Further, when the differential pressure ΔPLs detected by the differential pressure detection device 53 is smaller than the set differential pressure, the control device 7
0 outputs a signal to the driving part of the solenoid valve 65, this solenoid valve 65 is switched to the lower position, the head side of the control actuator 41 and the tank are communicated, and the pressure oil of the hydraulic source 63 is transferred to the control actuator 41. The piston of the control actuator 41 moves to the right in the figure, the displacement volume is changed so that the flow rate discharged from the main hydraulic pump 22 increases, and the differential pressure ΔPLs approaches the set differential pressure. It is controlled to a large extent.

The other configurations are the same as those of the first embodiment described above.

Even in the second embodiment configured in this manner, control based on the load sensing differential pressure can be performed in the same way as in the first embodiment, and the same effects as in the first embodiment can be achieved.

FIG. 5 is an explanatory diagram showing main parts of a third embodiment of the present invention.

This third embodiment also differs from the first and second embodiments in the configuration of the flow rate control means for controlling the displacement of the main hydraulic pump 22. The flow control means in this third embodiment includes, for example, a hydraulic power source 63, electromagnetic valves 64 and 65, an input section 66, and a calculation section 6, which are the same as those in the second embodiment described above.
7. includes a storage unit 68 and a control device 70 including six output units, detects a tilt angle that determines the displacement of the main hydraulic pump 22, and outputs a tilt angle signal to the input unit 66 of the control device 70; The tilt angle detector 71 and the main hydraulic pump 22
A command device 72 outputs a signal for commanding a target flow rate, that is, a target tilt angle, to the input section 66 of the control device 70.

In this flow rate control means, the value of the command signal generated by the operation of the command device 72 and the value detected by the tilt angle detector 71 are compared in the calculation unit 67 of the control device 70, and a drive signal corresponding to the difference is output. The main hydraulic pump 22 outputs a flow rate corresponding to the amount of operation of the command device 72. The other configurations are the same as those of the first and second embodiments described above.

In this third embodiment, the flow rate of the main hydraulic pump 22 can be determined without depending on the load sensing differential pressure. Other effects are similar to those of the first embodiment.

FIG. 6 is an explanatory diagram showing the main part of the fourth embodiment of the present invention.

This fourth embodiment is based on the control force adding means 5 of the limiting means 52.
The control pressure generating means 60 constituting 4 is different from that in the first embodiment described above. The rest of the structure is the same as that shown in FIG. 1 described above. The control pressure generating means 60 in this fourth embodiment is interposed between a pilot hydraulic pressure source 73 and the pilot hydraulic pressure source 73 and the tank, and is outputted from the output section 58 of the controller 59 shown in FIG. A variable throttle member 74 that operates in response to a control force signal, a throttle valve 75 interposed between the variable throttle member 74 and the pilot oil pressure source 73, and a pipe between the throttle valve 75 and the variable throttle member 74. and a conduit 77 which connects the line 76 to the driving parts 35b to 40b of the branch compensating valves 35 to 40 shown in FIG.

Even in the fourth embodiment configured in this way, the variable throttle member 74 is driven in accordance with the signal output from the output section 58 of the controller 59, the amount of throttle is determined, and the output from the pilot hydraulic source 73 is controlled. As a control pressure by appropriately changing the magnitude of the pilot pressure to be applied, the driving parts 35b to 4 of the branch compensating valves 35 to 40 shown in FIG.
0b, and has the same effect as the first embodiment.

FIG. 7, FIG. 8, and FIG. 9 are explanatory diagrams showing main parts of fifth, sixth, and seventh embodiments of the present invention, respectively. These fifth, sixth, and seventh embodiments differ from the first embodiment shown in FIG. 1 in the structure of the driving portion of the branch compensating valve. The other configurations are the same as those of the first embodiment.

Diversion compensation valve 38 shown in FIG. 7, which is the main part of the fifth embodiment
A is provided corresponding to the boom cylinder 26 that serves as a low pressure actuator during, for example, combined operation of swing and boom,
One of the driving portions 38Aa constitutes a receiving portion that receives a control force due to the control pressure guided through the conduit 51 shown in FIG. ing.

In this fifth embodiment, the controller 59 shown in FIG.
The functional relationship shown in FIG. 10 is set in advance in the storage unit 59 of. The functional relationship shown in FIG. 10 is that the differential pressure ΔPL
If S is larger than the maximum flow rate compensation differential pressure A, the differential pressure Δp
The control force F gradually increases as ts increases, and the differential pressure ΔP
When LS becomes less than the minimum flow compensation differential pressure A, the differential pressure ΔPL5
The relationship is such that a constant control force F, that is, a minimum flow rate compensation control force fc is output regardless of the decrease in the flow rate.

In the fifth embodiment configured as described above, when transitioning from independent drive of the boom cylinder 26 to combined drive of the swing motor 23 and the boom cylinder 26, the swing motor 23
Even when the driving pressure becomes the maximum load pressure and the differential pressure ΔPLs decreases below the minimum flow rate compensation differential pressure A, the control force F applied to one drive part 35Aa of the branch compensating valve 35A in FIG.
is the minimum flow rate compensation control force fc which is a force weak enough to compensate for the required opening amount. Therefore, the branch compensation valve 35A is not completely closed, and as in the first embodiment, the boom Fluctuations in the flow rate supplied to the cylinder 26 are suppressed, producing the same effect as the first embodiment.

Particularly, in the fifth embodiment, the main part of which is shown in FIG. 7, the structure is simple because a spring for biasing the drive section 38Aa is not required, and therefore manufacturing errors can be kept small. , along with this, the control accuracy is excellent.

Further, a shunt compensating valve 38B shown in FIG. 8, which is a main part of the sixth embodiment, is also provided corresponding to, for example, the boom cylinder 26, and is attached to one of the driving parts 38Ba in the direction of opening the shunt compensating valve 38B. The spring 38B1 is provided with a spring 38B1 that applies a biasing force, and a preset force variable means 38B2 that varies the preset force of the spring 38B1 in accordance with the control force generated by the control pressure introduced through the conduit 51 shown in FIG.

In this sixth embodiment, the controller 59 shown in FIG.
The functional relationship shown in FIG. 11 is set in advance in the storage unit 57 of . The functional relationship shown in FIG. 11 is almost the same as that shown in FIG. 10 described above, so a detailed explanation will be omitted.

In the sixth embodiment configured in this manner, in almost the same way as the fifth embodiment, the functional relationship shown in FIG. The minimum flow rate compensation control force fc is applied to the preset force variable means 38B2, and accordingly, the preset force of the spring 38B1 is changed so as not to completely close the branch compensation valve 35B.

In addition to achieving the same effects as the first embodiment, the sixth embodiment can set the pressure receiving area of the preset force variable means 38B2 regardless of the size of the pressure receiving area of the drive section 38Ba of the branch compensating valve 38B. , a large degree of freedom in design and production.

Further, a shunt compensating valve 38C shown in FIG. 9, which is a main part of the seventh embodiment, is also provided corresponding to, for example, the boom cylinder 26, and is connected to one of the driving parts 38Ca. A pressure supply means 3803 is provided to supply a constant pressure from the hydraulic pressure source 38C2 defined by the relief valve 38C1 so that the valve operates in the opening direction, and the other drive section 38Cb is connected to the conduit 51 shown in FIG. and is configured to provide a control pressure guided through the control pressure.

In this seventh embodiment, the functional relationships shown in FIG. 12 are preset in the storage units 57 of the five controllers shown in FIG. 1. The functional relationship shown in FIG. 12 is based on the first
Since this is almost the same as FIG. 3 showing the characteristics in the embodiment, detailed explanation will be omitted.

In the seventh embodiment configured in this way, in almost the same way as the first embodiment, the driving pressure of the swing motor 23 is changed when shifting from the independent drive of the boom cylinder 26 to the combined drive of the swing motor 23 and the boom cylinder 26. When becomes the maximum load pressure and the differential pressure ΔPLs decreases, the minimum flow rate compensation control force fc based on the relationship shown in FIG. is driven so that it does not close.

In this seventh embodiment, in addition to achieving the same effect as the first embodiment shown in FIG. 1, if a failure occurs in the signal system including the controller 59, the pressure Since the branch compensation valve 38C is controlled to be driven in the opening direction, pressure oil can be supplied to the boom cylinder 26 from the main hydraulic pump 22, and the boom cylinder 26 can be driven even in this emergency.

FIG. 13 is a circuit diagram showing an eighth embodiment of the present invention.

In the eighth embodiment shown in FIG. 13, the main hydraulic pump 22
a is a constant displacement hydraulic pump, and a discharge amount control means for controlling the flow rate discharged from the main hydraulic pump 22a is configured to control the pump pressure led through the pipe line 43a and the pipe line 4.
It consists only of a flow rate regulating valve 42a that is driven in accordance with a pressure difference ΔPL5 between the maximum load pressure and the maximum load pressure introduced through the flow rate adjusting valve 42a. Similarly to the first embodiment shown in FIG.
5. Actuators such as boom cylinder 26, arm cylinder 27, bucket cylinder 28, etc., two swing direction control valves, right travel direction control valve 30, right travel direction control valve 31, boom direction control valve 32, arm Directional control valve 3
3. It is equipped with a flow rate control valve such as a packet directional control valve 34 and branch compensation valves 35 to 40.

Further, the control force adding means 60a constituting the control force adding means 54 of the restriction means 52 includes six electromagnetic valves 62a, 62b provided corresponding to each of the branch compensation valves 35 to 40,
62c, 62d, 62e, 62f and these solenoid valves 6
The configuration includes a pilot pump 61a that supplies pilot pressure to 2a to 62f, and a relief valve 61b that defines the magnitude of the pilot pressure output from the pilot pump 61a. The solenoid valve 62a and the drive section 35b of the branch compensation valve 35 are connected via a pipe 51a, and similarly, each of the solenoid valves 62b to 62f and each of the drive sections 36b to 40b of the branch compensation valves 36 to 40 are connected to each other. are connected via each of the conduits 51b to 51f. Further, the solenoid valves 62a to 62f are driven by drive signals a, b, c, d, e outputted from the output section 58 of the controller 59.
, f, respectively.

The storage section 57 of the controller 59 stores the solenoid valve 6.
Corresponding to each of 2a to 62f, namely, the turning direction control valve 29, the left travel direction control valve 30, the right travel direction control valve 31, the boom direction control valve 32, the arm direction control valve 33, and the packet. Corresponding to each of the branch compensating valves 35 to 40 that control the differential pressure across the directional control valve 34, the functional relationship between the differential pressure ΔPL5 and the control force F, which takes into account the actuator speed that performs various operations, is determined individually. remembered.

In the eighth embodiment configured in this way, for example, when the swing motor 23 and the boom cylinder 26 are driven in combination, different control forces corresponding to changes in the differential pressure Δpts are used as drive signals a and d to control the solenoid valve. 62a, 62d, whereby the pilot pressure outputted from the pilot pump 61a is applied to the drive unit 35b of the branch compensation valve 35 related to the swing motor 23 as pilot pressures of different magnitudes via the electromagnetic valves 62a, 62d. , is given to each of the driving parts 38b of the branch flow compensation valve 38 related to the boom cylinder 26, and the branch flow compensation valves 35 and 38 are driven so that the differential pressures across the two swing direction control valves and the boom direction control valve 32 are mutually Compared to the case of the first embodiment shown in FIG. Similarly to the first embodiment, depending on each function relationship stored in the section 57, the operation changes from independent driving of the boom cylinder 26 to the swing motor 23 and the boom cylinder 2.
6, when shifting to the combined drive of 6, it has the effect of suppressing fluctuations in the supply flow rate to the boom cylinder 26 due to a significant decrease in the differential pressure ΔPL5, and the speed of the swing motor 23 and the boom that are considered optimal depending on the type of work are achieved. A combination of cylinder 26 speeds can be obtained.

Further, in the above embodiment, the combined drive of the swing motor 23 and the boom cylinder 26 was explained as an example, but the combination of actuators is not limited to these, and any combination of actuators may be used as a single drive of the low pressure actuator. At the time of transition from drive to combined drive of the low pressure actuator and high pressure actuator, the fluctuation in the flow rate supplied to the low pressure actuator due to the decrease in differential pressure ΔPLs due to the high pressure actuator drive pressure reaching the maximum load pressure is suppressed in the same manner as described above. Of course it is possible.

<Effects of the Invention> Since the hydraulic drive device of the present invention is configured as described above, when transitioning from single drive of the low pressure actuator to combined drive of the low pressure actuator and the high pressure actuator, the flow rate supplied to the low pressure actuator is reduced. Therefore, it is possible to prevent the occurrence of large shocks that conventionally occur, and it has the effect of improving operability compared to the conventional method.

[Brief explanation of the drawing]

FIG. 1 is a circuit diagram showing a first embodiment of the hydraulic drive device of the present invention, FIG. 2 is a flowchart showing a processing procedure in a controller provided in the first embodiment shown in FIG.
FIG. 3 is a diagram showing the functional relationship between the differential pressure set by the controller and the control force corresponding to the branch compensating valve provided in the first embodiment shown in FIG. 1, and FIG. FIG. 5 is an explanatory diagram showing the essential parts of the third embodiment of the present invention, and FIG. 6 is an explanatory diagram showing the essential parts of the fourth embodiment of the present invention. , FIG. 7 is an explanatory diagram showing the main part of the fifth embodiment of the invention, FIG. 8 is an explanatory diagram showing the main part of the sixth embodiment of the invention, and FIG. 9 is an explanatory diagram showing the main part of the sixth embodiment of the invention. FIG. 10 shows the functional relationship between the differential pressure set by the controller and the control force corresponding to the branch compensating valve provided in the fifth embodiment shown in FIG. 7. 11 is a diagram showing the functional relationship between the differential pressure set by the controller and the control force corresponding to the branch compensating valve provided in the sixth embodiment shown in FIG. 8, and FIG. 12 is the diagram shown in FIG. 9. 13 is a diagram showing the functional relationship between the differential pressure set by the controller and the control force corresponding to the shunt compensation valve provided in the seventh embodiment shown in FIG. 13, and FIG. 13 is a circuit diagram showing the eighth embodiment of the present invention. , FIG. 14 is a circuit diagram showing an example of a conventional hydraulic drive device. 22.22a... Main hydraulic pump, 23...
...Swivel motor, 26...Boom cylinder, 2
9...Swivel directional control valve, 32...Boom directional control valve, 35.38.38A, 38B, 38
C... Diversion compensation valve, 35a, 35b, 38a,
38b, 38Aa, 38Ba, 38Ca, 38Cb-
...-Drive unit, 41... Control actuator, 42.42a... Flow rate adjustment valve, 38B1.4
5.48... Spring, 38B2... Preset force variable means, 38C1... Relief valve,
38C3...Pressure supply means, 51.76.77
...Pipeline, 52...Restriction means, 53.
... Differential pressure detection device, 54 ... Control force adding means, 5 ... Controller, 60.60a.
...Control pressure generation means, 38C2, 61, 63,
73...Hydraulic source, 61a...Pilot hydraulic pump, 62.62a, 62b, 62c, 62d
, 62e, 62f, 64.65... solenoid valve, 7
0...Control device, 71...Tilt angle detector, 72...Command device, 74...Variable aperture member, 75... - Throttle valve. Fig. 2 Fig. 4 bσ Fig. 5 Fig. 3 Fig. 1O Fig. hjsl1 Fig. 12 jI6 Fig.■ Fig. 7 Fig. 8 Fig. 9 Fig. 1-- □-, J Fig.

Claims (13)

[Claims] (1) One main hydraulic pump, multiple actuators including a high-pressure actuator and a low-pressure actuator driven by pressure oil supplied from the main hydraulic pump, and controlling the flow of pressure oil supplied to these actuators. A flow rate control valve that controls the pressure oil of the main hydraulic pump, a branch flow compensation valve that controls the differential pressure before and after these flow control valves, and a flow control valve that controls the flow rate discharged from the main hydraulic pump. In a hydraulic drive device that supplies power to each of the above actuators via a flow compensation valve and a flow rate control valve, and is capable of combined driving of these actuators, it is possible to convert from single driving of the actuator to combined driving of the low pressure actuator and the high pressure actuator. 1. A hydraulic drive device, comprising: a restriction means for restricting the driving of the branch compensating valve corresponding to the low pressure actuator in the closing direction when the low pressure actuator moves to the closing direction. (2) A flow rate control means for controlling the flow rate discharged from the main hydraulic pump according to the pressure difference between the pressure of the pressure oil discharged from the main hydraulic pump and the maximum load pressure of the actuator. The hydraulic drive device according to claim (1). (3) The flow rate control means includes a command device that commands a target flow rate of the main hydraulic pump, and a discharge amount control means that controls the discharge amount of the main hydraulic pump in accordance with a command signal output from the command device. The hydraulic drive device according to claim 1, characterized in that: (4) The limiting means includes a differential pressure detection device that detects the differential pressure between the pressure of the pressure oil discharged from the main hydraulic pump and the maximum load pressure of the actuator, and a signal output from the differential pressure detection device. Claim (1) characterized in that the drive section of the branch flow compensation valve provided corresponding to the low pressure actuator includes control force applying means for applying a control force to restrict the driving of the branch flow compensation valve in the closing direction. Hydraulic drive system. (5) the control force adding means is connected to the differential pressure detection device;
A controller has a memory section that stores in advance the functional relationship between differential pressure and control force, and a control force signal output from this controller is applied to a drive section of a shunt compensation valve provided corresponding to a low pressure actuator. 5. The hydraulic drive device according to claim 4, further comprising a control pressure generating means for generating a control pressure. (6) The control pressure generating means includes a pilot hydraulic pressure source and a solenoid valve that is disposed between the pilot hydraulic pressure source and the drive section of the branch compensation valve and operates in response to a control force signal output from the controller. Claim (5)
Hydraulic drive as described. (7) The hydraulic drive device according to claim (6), wherein only one electromagnetic valve is provided for each of the plurality of branch compensation valves. (8) The hydraulic drive device according to claim (6), wherein a plurality of electromagnetic valves are provided corresponding to each of the plurality of branch compensation valves. (9) The control pressure generating means includes a pilot hydraulic pressure source, a variable throttle member that is interposed between the pilot hydraulic pressure source and the tank, and operates in response to a control force signal output from the controller, and the variable throttle member. and the pilot oil pressure source, and a pipe line connecting the pipe line between the throttle valve and the variable throttle member to the driving part of the flow-dividing compensation valve. The hydraulic drive device according to item (5). (10) The shunt compensation valve provided corresponding to the low-pressure actuator has a spring in one driving part that biases the shunt compensation valve to operate in the opening direction, and a control force in the other driving part. 5. The hydraulic drive device according to claim 4, wherein: (11) A claim characterized in that, in the branch flow compensation valve provided corresponding to the low pressure actuator, one of the drive parts forms a receiving part that receives a control force that applies a force in the direction of opening the branch flow compensation valve. (4) Hydraulic drive device as described. (12) The shunt compensating valve provided corresponding to the low pressure actuator has a spring in one of its driving parts that biases the shunt compensating valve to operate in the opening direction, and the spring is biased in response to the control force. 5. The hydraulic drive device according to claim 4, further comprising a preset force variable means for varying the preset force of the hydraulic drive device. (13) A pressure supply means is provided which is connected to one driving part of the shunt compensating valve provided corresponding to the low pressure actuator and supplies a constant pressure so that the shunt compensating valve operates in the opening direction; 5. The hydraulic drive device according to claim 4, wherein a control force is applied to the drive section.