JPH06117405A - Valve gear - Google Patents

Abstract

PURPOSE:To offer a valve gear capable of causing an actuator to make stable composite drive even in case that the sum of the flow requested by a flow control valve exceeds the maximum discharge flow of an oil pressure pump and the pressure differential DELTAPLs between a supply pressure Ps and a maximum load pressure PLmax changes. CONSTITUTION:With regard to a valve gear consisting of a plural number of directional control valves 38 including supply passages 42, 43, load passages 46, 47, 48, 49, first passages 44, 45, second passages 50, 51, flow control valves 36, 39 and pressure controllers 37, 40, and also including a third passage 62 and a pressure control means 70 to control the pressure oil flowing through the third passage 62, provision is made for a first induction means 76, which induces the maximum load pressure to oneside drive portion of the pressure control means 70, and a second induction means 75 which induces the supply pressure of an oil pressure supply source 33 to the other side drive portion of the pressure control means 70 in the manner opposing to the maximum load pressure.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a valve device provided in a civil engineering / construction machine having a plurality of actuators such as a hydraulic excavator.

[0002]

2. Description of the Related Art FIG. 4 is a circuit diagram showing a hydraulic circuit provided with a conventional valve device. The conventional valve device 3 shown in FIG.
Reference numeral 0 denotes a hydraulic circuit for civil engineering / construction machinery such as a hydraulic excavator, for example, a variable displacement hydraulic pump 31 and the hydraulic pump 3
1 to a hydraulic circuit including a pressure oil supply source 33 including a regulator 32 that controls the flow rate discharged from the hydraulic pump 1, and a plurality of actuators driven by the pressure oil supplied from the hydraulic pump 31, for example, hydraulic cylinders 34 and 35. It is provided. Then, the valve device 30 includes a flow control valve 36 that controls the flow of pressure oil supplied to the hydraulic cylinder 34, a direction switching valve 38 that includes a pressure controller 37, and the hydraulic cylinder 3
Flow control valve 39 for controlling the flow of pressure oil supplied to
And a directional control valve 41 including a pressure controller 40. The direction switching valves 38 and 41 described above include the first passages 44 and 45 connected to the hydraulic pump 31 and the load passage 4 connected to the hydraulic cylinders 34 and 35, respectively.
6, 47, 48, 49, and second passages 50, 51 connectable to the first passages 44, 45 and the load passages 46, 47, 48, 49, respectively. The flow rate control valves 36 and 39 described above control the flow rate of the pressure oil passing between the supply passages 42 and 43 and the first passages 44 and 45, respectively, and the second passages 50 and 51 and the load passages. 46, 47, 48, 49 are closed or communicated with each other. In addition, the pressure controller 37 described above
Is disposed between the first passage 44 and the second passage 50, and the pressure controller 40 includes a first passage 45 and a second passage 51.
And pressure control devices 37, 40, which are arranged between the control pressures guided by the control line 56 and the springs 66, 6
It can be driven by the spring force of 7 and the line pressure of the first passages 44 and 45.

Further, the valve device 30 includes a second passage 5
A transmission passage 57 connecting the 0 and the control pipe 56, a transmission passage 58 connecting the second passage 51 and the control pipe 56,
A check valve 59 that is provided in the transmission passage 57 and blocks the flow of the pressure oil from the control pipeline 56 toward the second passage 50, and a check valve 59 that is provided in the transmission passage 58 and that extends from the control pipeline 56 to the second passage 51.
The check valve 60 for blocking the flow of pressure oil in the direction, the control conduit 56, the third passage 62 that can communicate with the tank 61, and the pressure compensation valve 65 provided in the third passage 62. I have it.

The regulator 32 constituting the above-mentioned pressure oil supply source 33 has a difference between the supply pressure Ps and the control pressure of the control line 56, that is, the maximum load pressure PLmax of the load pressures of the hydraulic cylinders 34, 35. Pressure ΔPLs (= Ps
The flow rate control means for controlling the flow rate of the hydraulic pump 31 is configured such that −PLmax) becomes a set value set by the spring 64, that is, the force due to the differential pressure ΔPLs and the force of the spring 64 are balanced. .

In this prior art, the directional control valve 3
When the flow control valves 36 and 39 of 8 and 41 are switched by operating the operation levers 68 and 69, the pressure oil of the hydraulic pump 31 is correspondingly supplied to the supply passage 4 respectively.
2, 43, the variable throttle portion 52 (53) or the variable throttle portion 54 (55) and the first passages 44, 45, whereby the pressure controllers 37, 40 are pushed upward in FIG. The pressure oil is supplied to the second passages 50 and 51 and the load passage 4
6 (47) or 48 (49) to supply to the hydraulic cylinders 34 and 35.
4, 35 composite operations are performed. During this combined operation, the load pressure of the hydraulic cylinder 34 changes the load passages 46, (4
7) to the second passage 50, and further to the control pipe 56 via the transmission passage 57 and the check valve 59, while the load pressure of the hydraulic cylinder 35 is changed to the load passage 48 (4).
9) to the second passage 51, and further to the control pipe 56 via the transmission passage 58 and the check valve 60, so that the larger one of the load pressures of the hydraulic cylinders 34, 35 is eventually determined. The pressure, that is, the maximum load pressure PLmax is taken out as the control pressure in the control line 56. This maximum load pressure PLmax is the control chamber 37a, 4 of the pressure controller 37, 40.
0a of the pressure receiving portions 37b and 40b, whereby the pressure controllers 37 and 40 are lowered from the above-mentioned raised state against the supply pressure Ps, and the pressures Pa1 and Pa2 in the first passages 44 and 45 are increased. Are increased, and the pressures Pa1 and Pa2 in the first passages 44 and 45 are controlled by the maximum load pressure PLmax and the spring forces of the springs 66 and 67. The spring forces of the springs 66 and 67 are set to be substantially equal to each other, whereby the pressures Pa1 and Pa2 are controlled to be equal to each other. Then, the control pressure of the control line 56, that is, the maximum load pressure P of the hydraulic cylinders 34, 35.
Lmax is guided to one drive section of the regulator 32,
Differential pressure ΔPLs between supply pressure Ps and maximum load pressure PLmax
The hydraulic pump 31 supplies a flow rate that balances the force generated by the force with the force of the spring 64 that biases the regulator 32.

When the relation between the force acting on the pressure controllers 37 and 40 and the force acting on the regulator 32 is considered by formulating, the spring forces of the springs 66 and 67 are Fk, and the pressure controllers 37 and 40. Assuming that the pressure receiving areas of the pressure receiving portions 37b and 40b are A, from the balance of the forces acting on the pressure controllers 37 and 40, Pa1-PLmax = Fk / A (1) Pa2-PLmax = Fk / A (2) To establish. On the other hand, in the regulator 32, Ps-PLmax = ΔPLs (3) holds. From the formulas (3) and (1), and the formulas (3) and (2), the variable throttle parts 5 of the flow control valves 36 and 39 are obtained.
The differential pressures across 2 (53) and 54 (55) are Ps-Pa1 = [Delta] PLs-Fk / A (4) Ps-Pa2 = [Delta] PLs-Fk / A (5), respectively. From these equations (4) and (5), Ps-Pa1 = Ps-Pa2 = constant and Pa1 = Pa2 = Pa (6).

As described above, the pressures on the inflow sides of the variable throttle portions 52 (53) and 54 (55) of the flow rate control valves 36 and 39 are equal in the pressures of the supply passages 42 and 43, that is, the supply pressure Ps, and The pressure on the outflow side, ie the first passage 44,
The pressures Pa1 and Pa2 in the valve 45 are also equal to each other as described above, so that the variable throttle portions 5 of the flow control valves 36 and 39 can be controlled.
The differential pressures across 2 (53) and 54 (55) are always equal. Therefore, the respective throttle amounts of the variable throttle portions 52 (53) or 54 (55) with respect to the respective stroke amounts of the flow rate control valves 36 and 39, that is, the flow rates corresponding to the opening amounts, are applied to the respective load of the hydraulic cylinders 34 and 35. It is possible to realize stable combined driving of the hydraulic cylinders 34 and 35 without affecting the other hydraulic cylinders due to fluctuations.

Since the control line 56 communicates with the tank 61 via the pressure compensating valve 65, the control pressure of the control line 56 is released to the tank 61 when the flow control valves 36 and 39 are in the neutral position. The controllers 37 and 40 can be put into an unloaded state. Further, when the flow rate control valves 36, 39 are driven to be switched, the flow rate of the fluid flowing from the control line 56 to the tank 61 can be controlled by the pressure compensation valve 65 to suppress energy loss.

[0009]

As described above, in the prior art, when the hydraulic cylinders 34 and 35 are driven in combination, the variable throttle portions 52 (5) of the flow control valves 36 and 39, respectively.
3), or the differential pressure across the variable throttle 54 (55) is made equal. If the hydraulic cylinder 35 side is set to the maximum load pressure PLmax side, the pressure oil flows to the check valve 6
0, the control line 56, the passage 62, and the pressure compensating valve 65 to the tank 61, and the maximum load pressure PLmax to the respective control chambers 37a and 40a of the pressure controllers 37 and 40.
Tell. At this time, the pressure oil? PLoss is generated between the pressure controllers 37 and 40 due to the pressure oil flowing through the pipe 56. As a result, the variable throttle portion 52 of the flow control valve 36
The differential pressure across (53) increases by ΔPLoss. This is the differential pressure ΔPL between the supply pressure Ps and the maximum load pressure PLmax.
When s maintains the set value, the magnitude is not a problem, but when the sum of the required flow rates of the flow rate control valves 36 and 39 exceeds the maximum discharge flow rate of the hydraulic pump 31, the supply pressure Ps and the control pipe are Control pressure in the road 56, that is, maximum load pressure PLm
The pressure difference ΔPLs with respect to ax cannot be maintained at the set value and becomes small. The above-mentioned ΔPLoss becomes a size that cannot be ignored with respect to the differential pressure ΔPLs, and the variable throttle portion 52 of the flow control valve 36.
The differential pressure across (53) is equal to ΔPLoss for the flow control valve 3
It becomes larger than the differential pressure across the variable throttle unit 54 (55) of No. 9 and the operability during composite driving deteriorates.

It is an object of the present invention that even if the sum of the required flow rates of the flow control valves exceeds the maximum discharge flow rate of the hydraulic pump and the differential pressure ΔPLs between the supply pressure Ps and the maximum load pressure PLmax changes, An object of the present invention is to provide a valve device that can obtain a flow rate according to the opening degree and can stably drive the actuator in combination.

[0011]

To achieve this object, the present invention provides a supply passageway (42, 43) connected to a pressure oil supply source (33) and a load passageway (46) connected to an actuator. , 47, 48, 49) and the supply passage (4
A second passageway (44,45) that can be reached to
The first passage (44, 45) and the load passage (4
6, 47, 48, 49) second passage (5
0, 51), the supply passages (42, 43) and the first
Of the variable throttle portion (52, 53, 54, 55) is communicated through the variable throttle portion (52, 53, 54, 55) that is closed or built in. The flow rate of the pressure oil passing between the supply passages (42, 43) and the first passages (44, 45) is controlled in accordance with the change of the load passage, and the second passages (50, 51) and the load are controlled. A flow control valve (36, 39) that closes or communicates with the passage (46, 47, 48, 49), the first passage (44, 45) and the second passage (50, 51). And a pressure controller (37, 40) for controlling the pressure in the first passages (44, 45), the directional control valve (3
8, 41), and the pressure controller (37, 40)
And a control line (5) for transmitting control pressure to the pressure controller (37, 40).
6), which guides the maximum load pressure among the load pressures of the actuators to the control line (56) as the control pressure, and can connect the control line (56) to the tank (61). In the valve device having the passage (62), the third passage (62) is provided with the pressure control means (70) for controlling the pressure of the pressure oil flowing through the third passage (62). The load pressure is controlled by the pressure control means (70).
The first guide means (76) for guiding the first drive means to one of the driving parts, and the pressure control means (7) for controlling the supply pressure of the pressure oil supply source (33).
The other drive unit of (0) is provided with the second guide means (75) for guiding so as to oppose the maximum load pressure.

[0012]

According to the present invention, the supply pressure in the supply pipeline and the maximum load pressure in the control pipeline are guided to the pressure control means provided in the passage connecting the control pipeline to the tank, and this differential pressure ΔPLs is set. When it becomes smaller than the value, the flow rate of the pressure oil flowing from the control pipeline to the tank is limited, and the pressure loss ΔPLoss generated between the pressure controllers in the control pipeline is reduced according to the differential pressure ΔPLs. As a result, the control pressure transmitted to each pressure controller, that is, the maximum load pressure, becomes almost the same value in each pressure controller, and the differential pressure across the variable throttle of each flow control valve,
That is, the pressure difference between the supply pressure and the pressure in the first passage controlled by the pressure controller is the same. Therefore, it is possible to obtain a flow rate proportional to the throttle amount of the variable throttle portion of each flow control valve.

[0013]

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

FIG. 1 is a circuit diagram showing a hydraulic circuit provided with a first embodiment of a valve device of the present invention. This FIG. 1 is drawn corresponding to FIG. 4 described above. The embodiment shown in FIG. 1 is also equipped with the same one as shown in FIG. That is, a variable displacement hydraulic pump 31, a pressure oil supply source 33 including a regulator 32 that controls the flow rate discharged from the hydraulic pump 31, and a plurality of actuators driven by the pressure oil supplied from the hydraulic pump 31, for example, It is provided in a hydraulic circuit including hydraulic cylinders 34 and 35. The valve device 30 a includes a flow control valve 36 that controls the flow of pressure oil supplied to the hydraulic cylinder 34, a direction switching valve 38 including a pressure controller 37, and a flow of pressure oil supplied to the hydraulic cylinder 35. A flow control valve 39 for controlling the pressure control valve 40 and a directional control valve 41 including a pressure controller 40 are provided.
The direction switching valves 38 and 41 described above are respectively used for the hydraulic pump 3.
1 and the first passages 44 and 55, and the load passages 46 and 4 that communicate with the hydraulic cylinders 34 and 35, respectively.
7, 48, 49, and second passages 50, 51 that are respectively in communication with the first passages 44, 45 and the load passages 46, 47, 48, 49. Each of the flow control valves 36 and 39 described above is connected to the supply passage 4
2, 43 and the first passages 44, 45 to control the flow rate of the pressure oil, and the second passages 50, 51 and the load passages 46, 47, 48, 49 are closed or contact. Further, the pressure controller 37 is arranged between the first passage 44 and the second passage 50, and the pressure controller 40 is arranged between the first passage 45 and the second passage 51. These pressure controllers 37 and 40 can be driven by the control pressure guided by the control line 56, the spring force of the springs 66 and 67, and the line pressure of the first passages 44 and 45.

Further, in the valve device 30a, a transmission passage 57 for connecting the second passage 50 and the control pipe 56, a transmission passage 58 for connecting the second passage 51 and the control pipe 56, and a transmission passage 57. Is provided in the control pipe 56 to the second passage 5
A check valve 59 for blocking the flow of pressure oil in the 0 direction and a check valve 60 provided in the transmission passage 58 for blocking the flow of pressure oil in the direction of the second passage 51 from the control pipe 56 are provided. ing.

In this embodiment, the control line 56 and the tank 6 are
The configuration of the pressure control means provided in the third passage 62 communicating with No. 1 is different from that of the conventional example shown in FIG. That is, the pilot line 75 guides the supply pressure Ps in the supply line 42 (43) to one drive unit of the flow rate control unit 100 provided in the pressure compensating valve 70 serving as the pressure control unit, and the flow rate is increased. The control pressure in the control line 56, that is, the maximum load pressure PLma, is applied to the other drive unit of the control unit 100.
x through a pilot line 76 so as to oppose the supply pressure Ps, and the differential pressure across the fixed throttle 71 in the pressure compensating valve 70 is the supply pressure Ps and the maximum load pressure PLm.
It is provided so as to have a differential pressure ΔPLs (Ps-PLmax) of ax. Here, the pilot line 76 is shown in FIG.
Although it is introduced inside the pressure compensation valve 70, it may be introduced directly from the control line 56.

The regulator 32 constituting the above-mentioned pressure oil supply source 33 has a difference between the supply pressure Ps and the control pressure of the control line 56, that is, the maximum load pressure PLmax of the load pressures of the hydraulic cylinders 34, 35. So that the pressure ΔPLs becomes a set value, that is, the force due to the differential pressure ΔPLs and the spring 6
The flow rate control means for controlling the flow rate of the hydraulic pump 31 is configured so as to balance with the force of 4.

The operation in such an embodiment is as follows.

When the flow control valves 36 and 39 of the directional control valves 38 and 41 shown in FIG. 1 are switched by operating the operating levers 68 and 69, the pressure oil of the hydraulic pump 31 is supplied accordingly. The passages 42, 43, the variable throttle portion 52 (53), or the variable throttle portion 54 (55),
The pressure controllers 37 and 40 are guided to the first passages 44 and 45, and the pressure controllers 37 and 40 are pushed upward in FIG. 1. Further, the pressure oil is supplied to the second passages 50 and 51 and the load passages 46 (47) or 48 (49). ) To hydraulic cylinders 34 and 35, and the combined operation of these hydraulic cylinders 34 and 35 is performed. During this combined operation, the load pressure of the hydraulic cylinder 34 passes through the load passage 46 (47) to the second passage 50.
To the control pipe 56 via the transmission passage 57 and the check valve 59, while the load pressure of the hydraulic cylinder 35 is led to the second passage 51 via the load passage 48 (49). Further, the pressure is guided to the control line 56 via the transmission passage 58 and the check valve 60, and eventually, the larger one of the load pressures of the hydraulic cylinders 34 and 35, that is, the maximum load pressure P.
Lmax is taken out as control pressure. This maximum load pressure PLmax is the control chamber 37a of the pressure controller 37, 40,
Is applied to each of the pressure receiving portions 37b and 40b of 40a,
As a result, the pressure controllers 37 and 40 are lowered from the above-mentioned raised state against the supply pressure Ps, the pressures Pa1 and Pa2 in the first passages 44 and 45 are respectively increased, and the maximum load pressure PLmax and the spring 66, The spring force of 67 controls the pressures Pa1 and Pa2 in the first passages 44 and 45.
The spring forces of the springs 66 and 67 are set to be substantially equal to each other, whereby the pressures Pa1 and Pa2 are controlled to be equal to each other. Then, the control pressure of the control line 56, that is, the maximum load pressure PLmax of the hydraulic cylinders 34 and 35 is guided to one drive portion of the regulator 32, and the differential pressure ΔP between the supply pressure Ps and the maximum load pressure PLmax.
The force of Ls and the spring 64 for urging the regulator 32
The hydraulic pump 31 supplies a flow rate that balances with the force of.

The control line 56 is connected to the third passage 62.
Since it communicates with the tank 61 via the pressure compensating valve 70 provided in, the control pressure of the control line 56 is released to the tank 61 when the flow rate control valves 36, 39 are in the neutral state, and the pressure controllers 37, 40 are connected. Can be unloaded. Further, when the flow rate control valves 36, 39 are driven to be switched, a pressure compensating valve that controls the flow rate flowing from the control line 56 to the tank 61 by the differential pressure ΔPLs between the supply pressure Ps and the control pressure PLmax of the control line 56. 70 to limit the energy loss.

Here, the flow control valve 36, which has been switched,
When the sum of the required flow rates of 39 exceeds the maximum discharge flow rate of the hydraulic pump 31, the differential pressure between the supply pressure Ps and the control pressure of the control line 56, that is, the maximum load pressure PLmax of the load pressures of the hydraulic cylinders 34, 35. ΔPLs becomes small. At this time, from the control line 56 to the third passage 62, the pressure compensation valve 70
The flow rate of the pressure oil flowing through the tank to the tank 61 is the differential pressure ΔP between the pump pressure Ps and the control pressure PLmax of the control line 56.
It is significantly limited by the pressure compensating valve 70 controlled by Ls and can be closed further. As a result, each pressure controller 3 generated by the pressure oil flowing through the control line 56
The pressure loss ΔPLoss between 7 and 40 is remarkably reduced and becomes substantially equal to the maximum load pressure PLmax transmitted to the control chambers 37a and 40a of the pressure controllers 37 and 40. Therefore, assuming that the pressure receiving areas of the pressure receiving portions 37b and 40b of the pressure controllers 37 and 40 are A and the spring force of the springs 66 and 67 is Fk, from the balance of the forces acting on the pressure controllers 37 and 40, Pa1 -PLmax = Fk / A (7) Pa2-PLmax = Fk / A (8) holds. On the other hand, since the sum of the required flow rates of the flow rate control valves 36 and 39 exceeds the maximum discharge flow rate of the hydraulic pump 31, Ps−PLmax = ΔPLs ′ (9) holds regardless of the regulator 32. From the equations (9) and (7), and the equations (9) and (8), the differential pressure across the flow control valves 36 and 39, that is, the differential pressure across the variable throttle portions 52, 53 and 54 and 55, Ps-Pa1 = [Delta] PLs'-Fk / A (10) Ps-Pa2 = [Delta] PLs'-Fk / A (11). From these equations (10) and (11), Ps-Pa1 = Ps-Pa2 = constant and Pa1 = Pa2 = Pa (12).

As described above, the sum of the required flow rates of the flow rate control valves 36 and 39 exceeds the maximum discharge amount of the hydraulic pump 31, and the supply pressure Ps and the control pressure of the control line 56, that is, the maximum load pressure PL.
Even when the pressure difference ΔPLs with respect to max becomes smaller than a predetermined value, the variable throttle portions 52 of the flow rate control valves 36 and 39.
The pressure on the inflow side of (53), 54 (54) is
The pressures of 2 and 43, that is, the supply pressure Ps, are equal, and the pressures on the outflow side, that is, the pressures Pa1 and Pa2 in the first passages 44 and 45 are equal to each other as described above. Each variable diaphragm unit 52 (5
3), the differential pressure across 54 (55) is always the same. Therefore, the variable throttle portions 52 (53) or 54 (55) have their respective throttle amounts with respect to the stroke amounts of the flow control valves 36 and 39, that is, the flow rates corresponding to the opening amounts cause the load fluctuations of the hydraulic cylinders 34 and 35, respectively. It is possible to realize stable combined driving of the hydraulic cylinders 34 and 35 without affecting the other hydraulic cylinders.

FIG. 2 is a circuit diagram including a second embodiment of the present invention.

In the second embodiment, the control conduit 56 serves as a pressure control means for the third conduit 62 communicating with the tank 61, and the pilot conduit 75 for guiding the supply pressure Ps and the control conduit 5 are provided.
A pressure control valve 72 is provided so as to oppose a pilot line 76 that guides the control pressure PLmax of 6. The pressure control valve 72 controls the pressure of the pressure oil flowing through the control line 56,
The pressure is controlled by the differential pressure ΔPLs between the supply pressure Ps and the control pressure of the control line 56, that is, the maximum load pressure PLmax.
When ΔPLs becomes smaller than the predetermined value of the regulator 32, the flow rate of the pressure oil flowing from the control line 56 to the tank 61 is significantly limited, and the pressure loss between the pressure controllers 37 and 40 generated in the control line 56. ΔP Loss is reduced.
If the maximum load pressure PLmax varies little, this second
According to the embodiment, the pressure control means provided in the third passage 62 can be simplified in proportion to the first embodiment.

FIG. 3 is a circuit diagram including a third embodiment of the present invention.

In the third embodiment, a pressure compensating valve 111 to which a signal line 112 is connected is provided in the third passage 62 as a pressure control means, and the signal pipe 112 is provided at the other end thereof with the signal pipe 112. In order to control the pressure in the passage 112, an electromagnetic proportional pressure reducing valve 113 having a constant pressure source 114 connected to the primary side is connected,
The command signal 119 to the electromagnetic proportional pressure reducing valve 113 is output from a control device (not shown). The signal line 112 connected to the pressure compensating valve 111 is connected to one driving unit of the flow rate control unit 110 inside the pressure compensating valve 111, and the differential pressure across the fixed throttle 71 is changed to the signal line. 11
It is controlled according to the pressure of 2. On the other hand, the pilot conduit 115 connected to the control conduit 56 and the supply conduit 42 (4
The maximum load pressure PLmax and the supply pressure Ps are determined by the pressure detector 117 connected to the pilot line 116 connected to 3).
The differential pressure ΔPLs is detected and is transmitted to the control device (not shown) by the signal line 118. In the third embodiment configured in this way, the differential pressure ΔPLs is detected by the pressure detector 117 and transmitted to the control device by the signal line 118.
The pressure is calculated in the control device and transmitted to the electromagnetic proportional pressure reducing valve 113 by the command signal 119 to control the pressure in the signal line 112 to a pressure corresponding to the differential pressure ΔPLs.
To the variable flow rate of the pressure oil flowing in the third passage 62 according to the differential pressure ΔPLs.

According to the third embodiment, the same effect as that of the first embodiment can be obtained, and the flow rate of the pressure oil flowing through the third passage 62 is arbitrary regardless of the differential pressure ΔPLs. Can be controlled to, for example, the flow can be blocked even when the above-mentioned differential pressure ΔPLs maintains a set value.
It is also possible to eliminate the pressure loss ΔPLoss generated by the pressure oil flowing through 6.

[0028]

According to the present invention, since the pressure of the pressure oil flowing through the control line can be controlled by the pressure difference ΔPLs between the supply pressure Ps and the maximum load pressure PLmax, the sum of the required flow rates of the respective flow rate control valves can be controlled by the pump. Even when the above ΔPLs becomes smaller than the maximum discharge amount of, the flow rate corresponding to the opening area of each flow rate control valve can be obtained, and the actuator can be stably driven in combination.

[Brief description of drawings]

FIG. 1 is a circuit diagram showing a hydraulic circuit provided with a first embodiment of a valve device of the present invention.

FIG. 2 is a circuit diagram showing a hydraulic circuit provided with a second embodiment of the valve device of the present invention.

FIG. 3 is a circuit diagram showing a hydraulic circuit provided with a third embodiment of the valve device of the present invention.

FIG. 4 is a circuit diagram showing a hydraulic circuit provided with a conventional valve device.

[Explanation of symbols]

 30a Valve device 31 Variable displacement hydraulic pump 32 Regulator 33 Pressure oil supply source 34, 35 Hydraulic cylinder 36, 39 Flow control valve 37, 40 Pressure controller 38, 41 Directional switching valve 42, 43 Supply passage 44, 45 First passage 46, 47, 48, 49 Load passage 50, 51 Second passage 52, 53, 54, 55 Variable throttle portion 56 Control pipe 57, 58 Transmission passage 59, 60 Check valve 61 Tank 62 Third passage 70 Pressure Compensation valve 71 Throttle 72 Pressure control valve 75,76 Pilot line 111 Pressure compensation valve 113 Electromagnetic proportional pressure reducing valve 117 Pressure detector

Claims (3)

[Claims] 1. A supply passage (42, 43) connected to a pressure oil supply source (33), a load passage (46, 47, 48, 49) connected to an actuator, and the supply passage (4).
A second passageway (44,45) that can be reached to
The first passage (44, 45) and the load passage (4
6, 47, 48, 49) second passage (5
0, 51), the supply passages (42, 43) and the first
Of the variable throttle portion (52, 53, 54, 55) is communicated through the variable throttle portion (52, 53, 54, 55) that is closed or built in. The flow rate of the pressure oil passing between the supply passages (42, 43) and the first passages (44, 45) is controlled in accordance with the change of the load passage, and the second passages (50, 51) and the load are controlled. A flow control valve (36, 39) that closes or communicates with the passage (46, 47, 48, 49), the first passage (44, 45) and the second passage (50, 51). And a pressure controller (37, 40) for controlling the pressure in the first passages (44, 45), the directional control valve (3
8, 41), and the pressure controller (37, 40)
And a control line (5) for transmitting control pressure to the pressure controller (37, 40).
6), which guides the maximum load pressure among the load pressures of the actuators to the control line (56) as the control pressure, and can connect the control line (56) to the tank (61). In the valve device having the passage (62), the third passage (62) is provided with the pressure control means (70) for controlling the pressure of the pressure oil flowing through the third passage (62). The load pressure is controlled by the pressure control means (70).
The first guide means (76) for guiding the first drive means to one of the driving parts, and the pressure control means (7) for controlling the supply pressure of the pressure oil supply source (33).
0) The other drive unit is provided with a second guide means (75) for guiding the other drive section so as to oppose the maximum load pressure. 2. The first guiding means (76) and the second guiding means (75) are conduits.
The valve device described. 3. Valve arrangement according to claim 1, characterized in that a first guiding means (76) is connected to the control line (56).