JP3148722B2 - Hydraulic drive - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic drive system for supplying discharge oil of one or more hydraulic pumps used in construction machines and the like to a plurality of actuators, the actuator having at least one particularly large inertial load. When,
And a hydraulic drive suitable for simultaneously driving at least one actuator having a relatively small load.

[0002]

2. Description of the Related Art This type of hydraulic drive device is mainly used for construction machines and agricultural machines, and one that controls the discharge amount of a variable displacement pump in accordance with the load pressure is used. Also,
When driving a plurality of actuators, a pressure compensating valve is provided in each circuit to divide the pump discharge amount so that the actuators do not interfere with each other due to a difference in load pressure or the like of each actuator to cause a speed change of the actuators. Has been. Further, when the pump discharge amount falls below a predetermined required flow rate of a plurality of drive actuators, an actuator having a function of distributing the pump discharge amount to each actuator at an appropriate ratio, that is, an anti-saturation function is used.

As such a conventional hydraulic drive device, for example, a hydraulic circuit diagram of a first conventional hydraulic drive device shown in FIG. 5 is disclosed in US Pat. No. 5,347,811, JP-A-5-172112 or JP-A-8-254201. Is disclosed. The hydraulic circuit diagram shown in FIG. 5 is a hydraulic drive device having a load sensing function commonly called an after-orifice type, and a flow rate adjustment capable of controlling pump discharge oil flowing into the first and second actuators 12 and 13 respectively. First and second directional control valves 14 and 15 having functions, and first and second pressure compensating valves 50 and 51 for compensating the pressure of the first and second directional control valves correspond to each directional control valve. The first control chambers 50a and 51a are arranged between the pressure control valves 50a and 51a and are arranged between the pressure control valves 50a and 51a. The second control chambers 50b and 51b for applying pressure in the closing direction of the pressure compensating valve have a pressure receiving area,
And the second pressure receiving area are substantially equal.

According to this configuration, when the flow rate to each of the actuators 12 and 13 is relatively small and the sum of the flow rates does not reach the maximum discharge flow rate of the variable displacement pump 2, the respective directional control valves 14 and 15 are used. The directional control valve differential pressure before and after the throttle portion becomes equal to the differential pressure between the pump discharge pressure Pp and the maximum load pressure Pm, in other words, the differential pressure set in advance by the spring 18 of the flow control valve 17. Therefore, even if there is a difference between the load pressures of the respective actuators, the load pressure is not affected, and the flow rates to the respective actuators are set in advance by the throttle openings of the respective directional control valves 14 and 15 and the springs 18. It has a so-called load sensing function that is determined by the differential pressure and enables predetermined speed control. In addition, line 5
The maximum load pressure Pm of the variable pressure pump 2 is acted on a flow regulating valve 17 for driving the displacement changing means 6 of the variable displacement pump 2,
The differential pressure between the pump discharge pressure Pp and the maximum load pressure Pm is controlled to a pressure set by the spring 18 of the flow control valve 17.

In FIG. 5, an actuator 12 is a swing motor having a large inertia load for swinging an upper swing body of a hydraulic shovel, for example.
Is a boom cylinder with a relatively small load for rotating the boom, and the actuators 12 and 13 operate the swing motor and the operation levers of the directional control valves 14 and 15 for the boom cylinder to make strokes and operate simultaneously. Shall be. The stroke amount of the operation lever is usually a considerably large full stroke or a stroke amount close thereto.
Then, the discharge oil of the variable displacement pump 2 is supplied to the directional control valves 14 and 1
5, through the pressure compensating valves 50 and 51, flow into the actuators 12 and 13 to move the actuators,
The 12 inertia loads are too large to move immediately. Therefore, the actuator port on the inflow side of the swing motor 12
Excessive pressure is generated at 12a or 12b, the load pressure rises to the set pressure of the overload relief valve (not shown) installed at the inlet side actuator port, and most of the pressure oil flowing into the actuator port 12a or 12b It flows out to the tank T from an unillustrated overload relief valve provided at the port 12a or 12b.

Further, the load pressure which has risen to the set pressure of an overload relief valve (not shown) acts on the flow control valve 17 of the pump device via the shuttle valve 4 and the line 5 so that the load sensing mechanism operates. It works to increase the discharge flow rate of the variable displacement pump 2. However, when the discharge pressure of the variable displacement pump 2 rises to a preset value, the constant horsepower control mechanism of the constant horsepower control valve 19 operates preferentially, and conversely, the discharge flow rate decreases. Even in the state where the discharge flow rate is reduced in this manner, each pressure compensating valve acts to make the pressure line 7 upstream of each of the pressure compensating valves 50 and 51 equal pressure by the anti-saturation function described above. From each pressure compensating valve 5
The opening of 0,51 is in a state of being narrowed down. Therefore, the driving speed of the boom cylinder 13 becomes extremely slower than the case where the boom cylinder 13 is operated alone, and there is a problem that the work of stacking the truck is extremely difficult, the work efficiency is deteriorated, and the fatigue of the operator is increased. . At the same time, there arises a problem that an energy loss of the prime mover becomes large due to an overload relief (not shown) of the pressure oil flowing into the swing motor 12.

Next, when the rotation of the swing motor 12 is completed and the swing is performed at a steady speed, the drive torque of the swing motor rapidly decreases, that is, the swing load pressure sharply decreases, and the load pressure of the boom cylinder 13 decreases. Is larger. As a result, the maximum load pressure Pm in the line 5 drops rapidly, and the pressure in the pump discharge line 3 also drops, and the operation of the constant horsepower control mechanism described above is relaxed. Rapidly accelerates, and the actuator 12 and 13 move jerking as a whole during the simultaneous operation.

In order to solve such a problem, for example, Japanese Patent Application Laid-Open
In the publication No. 8-254201, the downstream side of the swing pressure compensating valve and the inflow port on the raising side of the boom cylinder actuator are communicated with a merging passage, and the communication passage is connected to the communication valve and the downstream side of the swing pressure compensating valve from the downstream side. Check valves that allow the boom cylinder actuator to flow into the inflow port are sequentially provided, and the pilot pressure increases according to the amount of operation provided to operate the directional control valve. When the communication valve is opened by the pilot pressure and the communication valve is opened by the pilot pressure, the check valve is sequentially opened, and pressure oil flows from the downstream side of the swing pressure compensation valve to the inflow port of the boom cylinder actuator. It has been proposed to prevent a sudden increase in the flow swirling load pressure and at the same time prevent a decrease in the rising speed of the boom cylinder.

However, Japanese Patent Application Laid-Open No. 8-254201 discloses a communication valve, a check valve, and the like in addition to an original pressure compensating valve in order to prevent a sudden increase in a turning load pressure and a decrease in a rising speed of a boom cylinder. Set up the attached valve of
An external pilot pressure is applied to these accessory valves so that the accessory valves themselves operate under certain conditions. For this reason, naturally, since various accessory valves are provided in addition to the original pressure compensating valve,
There is a problem that the overall size of the valve is large, an external pilot pressure is required, and the cost is complicated and expensive. There is also a problem that the operation of the boom cylinder becomes discontinuous because the accessory valve itself operates under certain conditions.

The problem of the hydraulic circuit diagram of the first conventional hydraulic drive device described above is, for example, the hydraulic circuit diagram of the second conventional hydraulic drive device shown in FIG. 6 (shown in Japanese Patent Laid-Open No. 7-324355). Also applies to The hydraulic circuit diagram shown in FIG. 6 is a hydraulic drive device having a load sensing function having an anti-saturation function. In FIG. 6, a plurality of directional control valves 24, 25 are connected in parallel to the discharge oil passage 3 of the variable displacement pump 2 via check valves 26, 27. The output sides of the valves 24 and 25 are connected to the actuators 12 and 13, respectively, and the return oil from each of the actuators 12 and 13 passes through the direction control valves 24 and 25 again, and then passes through a plurality of pressure compensating valves 60 and 61, respectively. The tank T is returned to the tank T via the tank line 16.
The pressure compensating valves 60 and 61 are disposed between the corresponding directional control valves 24 and 25 and the tank line 16 and are pressures downstream of the restrictor of the directional control valves, that is, the load pressures of the actuators 12 and 13.
The first control chamber 60 in which PL acts in the opening direction of the pressure compensating valve
a, 61a pressure receiving areas, and second control chambers 60b, 61b for applying the maximum load pressures Pm of the plurality of actuators detected by the shuttle valve 4 in the closing direction of the pressure compensating valve, and The first and second pressure receiving areas are made equal. With such a configuration, the same operation as that of the hydraulic circuit diagram of the first conventional hydraulic drive device shown in FIG.

To solve this problem, Japanese Patent Laid-Open No. 7-324355
In the publication, a bypass line that communicates with the pump discharge oil and communicates with the tank T in parallel with each directional control valve is added to the variable pump discharge line with respect to the hydraulic circuit of FIG. The pressure generating means is arranged in series, the pressure on the upstream side of the pressure generating means is guided to the displacement control device of the variable pump to perform the negative pump control, and the maximum load pressure among all the actuators is bleeded. The pressure compensating valve for the swing motor and the bleed-off valve act only on the off valve and the pressure compensating valve for the swing motor in the closing direction to swing the highest load pressure among actuators with a relatively small load pressure other than the swing motor. By operating the pressure compensating valves other than those for the motor in the closing direction, the load pressure of the Pressure compensating valves except for the swing motor due to the fact that is excessively closed, that load pressure other than for the turning motor is to prevent the speed reduction of the relatively small actuator is proposed me. As a result, the circuit configuration becomes complicated, the overall size of the valve increases, and the cost increases. When the swing motor accelerates and the load pressure of the swing motor decreases, the load pressure of the boom-side actuator increases, and the swing motor pressure compensating valve is closed with the boom-side load pressure, and the speed of the swing motor rapidly decreases. Some troubles came out.

Further, the above-mentioned problem of the hydraulic circuit diagram of the first conventional hydraulic drive device is described in, for example, USP 5,622,206,
This also applies to the hydraulic circuit diagram of the third conventional hydraulic drive device shown in FIGS. 7 and 8 disclosed in JP-A-5-332310 or JP-A-5-332311. In the hydraulic circuit diagram shown in FIG. 7, each pressure compensating valve 70, 71 is disposed between the pump 2 and the direction control valve 24, 25, and the pressure compensating valve is connected to the pump hydraulic oil line 3 from a plurality of actuators. Check valve portions 76 and 78 for preventing backflow of the valve and reducing the flow rate to the actuator, and a spool for closing the check valve portion
Pressure reducing valve parts 77 and 79 which can reduce the pump discharge pressure from the pump pressure oil line 3 to the maximum load pressure of a plurality of actuators, and compensate the pressure on the downstream side of the throttle part of each directional control valve. It has a first control chamber 77a, 79a pressure receiving area for acting in the opening direction of the valve, and a second control chamber 77b, 79b pressure receiving area for applying a plurality of actuator maximum load pressures in the closing direction of the pressure compensating valve. The first and second pressure receiving areas are made equal. FIG. 8 is a block diagram showing a sectional structure of the pressure compensating valve 70 in the hydraulic circuit diagram shown in FIG. With such a configuration, the same operation as that of the hydraulic circuit diagram of the first conventional hydraulic drive device shown in FIG.

To solve this problem, Japanese Patent Laid-Open No. Hei 5-332311
In the publication, in addition to the original directional control valve and the pressure compensating valve, in addition to the original direction control valve and the pressure compensating valve, the direction of the A pilot pressure for switching the control valve 25 is guided, and a pilot check valve that opens and closes with the pilot pressure is provided with a pump discharge pressure oil before the pressure reducing valve portion 77 of the pressure compensating valve on the side of the swing motor actuator 12 having a large load. There has been proposed a device that shuts off and suppresses an increase in the maximum load pressure of a plurality of actuators. As a result, the circuit configuration becomes complicated, the overall size of the valve increases, and the cost increases.

[0014]

SUMMARY OF THE INVENTION An object of the present invention is to solve the above-mentioned problems of the prior art, and to simultaneously provide at least one actuator having an extremely large inertia load and at least one actuator having a relatively small load. Even if the actuator is operated, sufficient pressure oil can be supplied to the small load side actuators, and even if the load pressure of the actuator with the large inertial load suddenly drops, the speed of each actuator changes suddenly. Enables smooth operation without shock without using. In addition, energy loss due to wasted relief oil from the overload relief valve and the load on the prime mover can be reduced.In addition, special additional valves other than the pressure compensating valve and external pilot pressure are unnecessary. An inexpensive hydraulic drive is provided.

[0015]

According to a first aspect of the present invention, there is provided a motor driven by discharge oil.
And at least a first and a second respectively having a load pressure
A hydraulic actuator, and the first and second hydraulic actuators;
The discharge oil flowing into the eater can be controlled individually.
First and second directional control valves having flow control functions,
A first directional control valve coupled to the first directional control valve;
A first pressure compensating valve for the pressure compensation of the first
Pressure compensating valve is a throttle section of the corresponding first directional control valve
To the first control chamber of the first pressure compensating valve
Act on the pressure receiving area in the opening direction, and
The highest load pressure of the first pressure compensating valve to the second
Acting on the pressure receiving area of the control room in the closing direction,
First pressure compensation of the pressure communicating with the load pressure of the tutor
Acting in the closing direction on the third control chamber pressure receiving area of the valve,
The first control room pressure receiving area and the second control room pressure receiving area are approximately equal.
And the third control room pressure receiving area is
It is only a small area with respect to the pressure area,
Corresponds to increase in load pressure of the first actuator
So as to reduce the outlet flow rate of the first pressure compensating valve.
A first pressure compensating valve and the first and second pressure compensating valves .
A variable displacement pump for discharging to an actuator of
A constant horsepower control device provided in a variable displacement pump;
And a discharge oil flow rate change device cooperating with a force control device.
Hydraulic drive, characterized in that preferably,
The third pressure receiving area of the first pressure compensating valve is equal to the first pressure compensating valve.
0.03 to 0.07 of the first pressure receiving area of the force compensating valve
The problem is solved by providing a hydraulic drive device characterized by
(Claims 1 and 2). The first and second hydraulic actuators
When the etas have different load pressures, at least one low load side first hydraulic actuator driven by the discharge oil and having a load pressure, and at least one high load side second hydraulic actuator having a larger inertial load than the actuator. Hydraulic actuators, first and second directional control valves having flow control functions capable of controlling the discharge oil flowing into the first and second actuators, respectively, and the first and second directions A first and a second pressure compensating valve respectively connected to a control valve for compensating the pressure of the first and second directional control valves, wherein each of the pressure compensating valves is a throttle of the corresponding directional control valve; The pressure downstream of the section acts on the first control chamber pressure receiving area of the pressure compensating valve in the opening direction, and the highest load pressure of the plurality of actuators is changed to the second pressure of the pressure compensating valve. The third control chamber pressure-receiving area of the pressure compensating valve is caused to act on the third control chamber pressure-receiving area of the pressure compensating valve in the closing direction. The second control room pressure receiving area is made substantially equal, and the third control room pressure receiving area is made only a small area with respect to those control room pressure receiving areas, thereby increasing the load pressure of the corresponding actuator. A first and a second pressure compensating valve adapted to reduce an outlet flow rate of the pressure compensating valve, a variable displacement pump for discharging the discharge oil to the first and second actuators, The problem has been solved by providing a hydraulic drive device comprising: a constant horsepower control device provided in a pump; and the discharge oil flow rate changing device cooperating with the constant horsepower control device (Claim 3).

Preferably, the rate of decreasing the outlet flow rate of the pressure compensating valve communicating with the high load side hydraulic actuator is made larger than the rate of decreasing the outlet flow rate of the pressure compensating valve communicating with the low load side hydraulic actuator. I have.
(Claim 4) More preferably, the third pressure receiving area of the pressure compensating valve of the high load side hydraulic actuator is 0.03-0.07 of the first pressure receiving area of the pressure compensating valve, and the pressure compensation of the low load side hydraulic actuator is performed. The third pressure receiving area of the valve is 0 to 0.02 of the first pressure receiving area of the pressure compensating valve (claim 5).

Further, according to the second aspect of the present invention, at least one low-load-side first hydraulic actuator driven by discharge oil and having a load pressure, and at least one high-hydraulic actuator having an inertia load larger than the actuator. A load-side second hydraulic actuator, the first and second hydraulic actuators;
First and second directional control valves each having a flow control function capable of controlling the discharge oil flowing into the actuator, and the first and second directional control valves respectively connected to the first and second directional control valves. The first for compensating the pressure of the second directional control valve
And a second pressure compensating valve, which is disposed between the corresponding directional control valve and the tank, wherein each of the pressure compensating valves reduces the pressure downstream of the throttle of the corresponding directional control valve. Acting in the opening direction on the first control chamber pressure receiving area of the valve,
The highest load pressure of the plurality of actuators is caused to act on the second control chamber pressure receiving area of the pressure compensating valve in a closing direction, and the first load compensating valve of the high load side second actuator side is operated.
The pressure receiving area of the second pressure receiving area is 0.93-0.97, and the first pressure receiving area of the low load side first actuator side pressure compensating valve is the second pressure receiving area of 0.98-1.00. The corresponding high-load second actuator of the pressure-compensating valve on the high-load second actuator side
First and second in which the outlet flow over data pressure compensating valve in correspondence with the increase in the load pressure of the to decrease more than the flow rate at the outlet of the low load side first actuator side pressure compensation valve A pressure compensating valve, a variable displacement pump that discharges the discharge oil to the first and second actuators, a constant horsepower control device provided in the variable displacement pump, and the constant horsepower control device cooperating with the constant horsepower control device. The above-mentioned problem has been solved by providing a hydraulic drive device having a discharge oil flow rate changing device.

[0018]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS A first embodiment, which is an embodiment of the first invention of the present invention, will be described with reference to FIG. 1 showing a hydraulic circuit diagram of the embodiment. FIG. 1 shows a plurality of directional control valves 14 and 15 connected in parallel to a discharge oil passage 3 of a variable displacement pump 2 driven by a prime mover 1 such as an engine. After passing, they are connected to a plurality of pressure compensating valves 10 and 11 via intermediate oil passages 7 and 7 and check valves 8 and 9, respectively, and after passing through directional control valves 14 and 15 again, each directional control valve 14 Actuator output side
12 and 13 respectively, and the return oil from each actuator 12 and 13 is returned to the tank line via directional control valves 14 and 15 again.
The tank is returned from 16 to tank T. Here, the actuator 13 is a low-load-side actuator having a relatively light load (for example, an actuator for a front mechanism such as a boom cylinder that raises and lowers the boom of a hydraulic shovel, an arm cylinder, and a bucket cylinder), and the actuator 12 has a large inertial load. (For example, a turning motor of a hydraulic shovel) or the like.

The pressure compensating valves 10 and 11 have an anti-saturation function, and apply a pressure upstream of each of the pressure compensating valves 10 and 11 to the first control chambers 10a and 11a in the direction of opening the pressure compensating valves. The maximum load pressure Pm of the maximum load pressure line 5 detected by the shuttle valve 4 acts on the pressure receiving areas of the second control chambers 10b and 11b in the closing direction. Normally, these first and second pressure receiving areas are made equal. Further, springs 10d, 11d are provided in a direction in which the pressure compensating valves 10, 11 are closed. Fig. 5
In the present invention shown in FIG. 1, the pressure downstream of each of the pressure compensating valves 10 and 11, that is, the pressure communicating with the load pressure PL of each actuator is changed to a pressure compensating valve. On the third control chambers 10c and 11c in the closing direction. The first and second pressure receiving areas are substantially equal, and the third pressure receiving area is a very small area (about 0 to 0.07) with respect to the first pressure receiving area. Thereby, the outlet flow rate of the pressure compensating valve is reduced in accordance with the corresponding increase in the load pressure of the actuator.

When the discharge pressure of the variable displacement pump 2 exceeds a preset pressure, the flow regulating valve 17 operates the displacement changing means 6 in the direction of decreasing the flow rate so that the flow rate is controlled so as not to exceed the rated torque of the motor 1. A so-called constant horsepower control valve 19 for regulating is used, and the constant horsepower control valve 19 is configured to operate with a higher priority than the above-described flow sensing valve 17 for load sensing. That is, the constant horsepower control is given priority over the load sensing control. Therefore, when the maximum load pressure Pm is relatively high, constant horsepower control is often applied, and the maximum discharge flow rate of the pump 2 is regulated. Therefore, the above-described anti-saturation function is an indispensable function. I have.

Further, the ratio of the third pressure receiving area of the high load side actuator 12 to the first or second pressure receiving area is determined by the third pressure receiving area of the low load side actuator 13 and the first or second pressure receiving area. Is larger than the ratio. That is, the ratio is set to about 0.03 to 0.07 in the high load side actuator 12, and is set to about 0 to 0.02 in the low load side actuator 13.

The directional control valves 14 and 15 may be operated by the pilot pressure of a pressure control valve in which the pilot pressure is increased according to the operation amount widely used in construction machines, or may be operated by electromagnetic proportional control. It may be operated by a valve or a high-speed on / off switching solenoid valve whose pulse width is controlled. In addition, although one actuator is described for each of the low-load side and the high-load side, it is for the description of the present invention, and a plurality of actuators having a similar circuit are actually provided. Needless to say. In this embodiment, a variable displacement pump for discharging one discharge oil to each of a plurality of actuators, a constant horsepower control valve 19 provided in the variable displacement pump, and a discharge oil flow changing device cooperating with the constant horsepower control device Is used, but a plurality of variable displacement pumps, a constant horsepower control valve 19 and a flow control valve 17 are used.
In addition, a constant horsepower control device 19, 6 and a bleed-off valve 17 'are used in place of the flow control valve 17, which is a discharge oil flow changing device as shown in FIG. 1 (b). There may be. The bleed-off valve 17 'may be installed on the valve device side.

The operation of the hydraulic drive will now be described. In FIG. 1, the balance of the forces acting on the pressure compensating valves 10 and 11 is considered. First, the pressure compensating valves 10, 11
F1 acting in the direction in which the pressure compensating valve is opened, the pressure on the upstream side of the pressure compensating valve (downstream of the direction control valve throttle section) is defined as Pd, and the pressure receiving area of the first control chambers 10a and 11a of the pressure compensating valve is defined as Aa F1 = (Pd ・ Aa) ...... (1) Conversely, the force F2 acting in the direction to close the pressure compensating valve is the maximum load pressure Pm, and the control chambers 10b and 11b of the pressure compensating valve Second
The pressure receiving area is defined as Ab, and the load pressure of the actuator downstream of the pressure compensating valve is defined as PL.
If the pressure receiving area of 10c and 11c is Ac, F2 = (Pm ・ Ab + PL ・ Ac) ...... (2)

Here, when controlling the pressure compensating valve, since the forces in both directions are balanced, the equations (1) and (2) become equal, so that F1 = F2, and (Pd · Aa) = (Pm・ Ab + PL ・ Ac) ...... (3) However, the acting force of the springs 10d and 11d is ignored as being weak. On the other hand, the pressure on the upstream side of the directional control valve throttle section, that is, the pump discharge pressure Pp of the discharge oil passage 3 of the variable displacement pump 2 is changed by the action of the above-described variable displacement pump 2 and the flow rate regulating valve 17 to Psp. Assuming that the differential pressure between the pump discharge pressure and the maximum load pressure set in advance by the spring 18 is Pp = Pm + Psp ...... (4) From equation (4), Pm = Pp-Psp ... .. (5) is obtained.

By substituting equation (5) into equation (3), (Pd · Aa) = (Pp−Psp) · Ab + PL · Ac (6) where the first pressure receiving area Aa Assuming that the pressure receiving area Ab is equal to the second pressure receiving area Ab, A = Aa = Ab ... (7) and equation (6) gives Pd = Pp-Psp + PL.Ac / A ... . (8) From equation (8), the directional control valve differential pressure ΔP = Pp-Pd before and after the throttle part of the directional control valve can be calculated as ΔP = Pp-Pd = Psp-PLAc / A ...... (9) (9) Psp = Pp-Pm obtained from Eq. (5)
By substituting the relationship, ΔP = Pp-Pm-PL * Ac / A ... (10)

According to the equation (9), the directional control valve differential pressure ΔP before and after the throttle portion of the directional control valve is equal to the spring 18 of the flow control valve 17.
Thus, the pressure Psp set in advance and the load pressure PL of the actuator downstream of the pressure compensating valve become a linear expression, and decrease with an increase in the load pressure PL of the actuator. That is, a downward-sloping pressure compensation characteristic in which the flow rate of the actuator decreases as the load pressure PL of the actuator increases. Further, according to the equation (10), the directional control valve differential pressure ΔP is determined by the pump discharge pressure Pp and the maximum load pressure Pm.
Differential pressure and load pressure of the actuator downstream of the pressure compensation valve
It becomes a linear expression of PL, and similarly, the load pressure PL of the actuator
Decreases, that is, a downward-sloping pressure compensation characteristic in which the flow rate of the actuator decreases as the load pressure PL of the actuator increases.

Here, in the present invention, the third pressure receiving area Ac is several% (0 to 0.07) of the first and second pressure receiving areas.
Therefore, the value of Ac / A in the second term of equation (9) and the third term of equation (10) is a very small value. Therefore, when the load pressure PL of the actuator downstream of the pressure compensating valve is relatively low, the value of the second term of the equation (9) and the value of the third term of the equation (10) are very small. Control valve differential pressure ΔP
Obtains ΔP = Psp ...... (11) from equation (9) or ΔP = Pp-Pm ...... (12) from equation (10). Therefore, when the load pressure PL is relatively low, the directional control valve differential pressure ΔP is equal to the pressure Psp set in advance by the spring 18 of the flow regulating valve 17, as in the conventional embodiment of FIG. The pressure difference becomes equal to the pressure difference between the pump discharge pressure Pp and the maximum load pressure Pm. Therefore, predetermined speed control can be performed regardless of the load pressure, and at the same time, an anti-saturation function is provided.

Further, in the present invention, as described above, in the pressure compensating valve of the high load side actuator, the third pressure receiving area is set to 0.03 to 0.07 with respect to the first and second pressure receiving areas, and the low load side pressure compensating valve is provided. Then, the third pressure receiving area is changed to the first and second pressure receiving areas.
0 to 0.02 with respect to the pressure receiving area
Ac / A is set to be larger in the high-pressure side pressure compensating valve than in the low-load side pressure compensating valve.

Now, for ease of explanation, the high load side actuator 12 is a swing motor, and the ratio of the third control chamber pressure receiving area to the first control chamber pressure receiving area of the swing motor side pressure compensating valve 10 is determined. As Ac / A, a case is considered in which the low-load-side actuator 13 is a boom cylinder and the ratio of the pressure receiving area of the pressure compensation valve 11 on the boom cylinder side is Ac / A = 0. In this case, equations (9) and (10) hold for the swing motor, and equations (11) and (12) hold for the boom cylinder. Therefore,
In the combined operation of the boom cylinder 13 and the swing motor 12,
If the pump discharge flow rate is sufficient and the saturation state has not been reached, the directional control valve differential pressure of the directional control valve 15 of the boom cylinder 13 on the low load side is expressed by the equations (11) and (12). Thus, the value is constant regardless of the load pressure of the boom. Therefore, even if the load pressure fluctuates, the flow rate remains constant. On the other hand, the directional control valve differential pressure of the directional control valve 14 of the swing motor 12, which is on the high load side, becomes smaller depending on the load pressure of the swing motor as shown in equations (9) and (10). Therefore, the flow rate decreases as the load pressure increases.

However, in such a combined operation, since the load pressure of the swing motor is excessive, the constant horsepower control mechanism of the pump device works preferentially as described above, and the pump discharge flow rate itself decreases, and the saturation state is reduced. Will be reached. In this state, the pump discharge pressure Pp cannot be kept higher than the maximum load pressure Pm by the pressure Psp set in advance by the spring 18 of the flow control valve 17. Let the pump pressure at this time be Pp 'and Pp'-Pm = Psp '
Then, the magnitude of this Psp 'depends on the degree of flow shortage at that time and does not become a constant value, but the same pump discharge pressure P is provided upstream of each directional control valve.
p 'is acting. At this time, each directional control valve differential pressure Δ
P 'is as follows.

The directional control valve differential pressure on the boom cylinder side is Δ
If Pb ', then ΔPb' = Psp '= Pp'-Pm ....... (13) If the directional control valve differential pressure on the swing motor side is ΔPs', ΔPs' = Psp'-PLsAc / A = Pp'-Pm-PLs Ac / A .... (14) is obtained. Here, PLs is the self-load pressure of the swing motor. According to equation (14), the directional control valve differential pressure ΔPs ′ on the swing motor side is Psp ′ with respect to the maximum load pressure Pm and the maximum load pressure Pm.
Pump discharge pressure Pp 'and self-load pressure
Dependent on PLs, even after saturation, the self-load pressure still decreases with increasing PLs. On the other hand (13)
According to the formula, the directional control valve differential pressure ΔPb ′ on the boom cylinder side
Does not depend on the self-load pressure of the boom but only on the maximum load pressure Pm and the pump discharge pressure Pp 'which is increased by Psp' with respect to the maximum load pressure Pm.

Therefore, in the initial stage of the simultaneous operation of raising the boom and turning, the pump discharge flow rate decreases with the rise of the turning load pressure that has risen sharply, and even when the pump enters the saturation state, the pump is supplied to the turning side with the rise of the turning load pressure. Since the flow rate is reduced, the flow rate becomes excessive as a whole, and the pump discharge pressure Pp 'becomes higher. Therefore, the boom-side directional control valve differential pressure ΔPb ′ increases according to Expression (13), and the boom-side flow rate increases. In other words, the flow rate on the boom side increases as the flow rate on the turning side decreases. Further, since the flow rate supplied to the swing motor is reduced, the useless relief flow rate from an overload relief valve (not shown) of the swing motor is reduced, and at the same time, the rise in the load pressure itself of the swing motor is suppressed. Therefore, an increase in the pump discharge pressure is suppressed low, the flow rate regulation by the constant horsepower control mechanism is relaxed, and the discharge flow rate itself increases. Thus, the flow rate of the boom further increases. In this way, it is possible to prevent the loss of energy of the prime mover due to the reduction of the boom speed and the relief of the pressure oil from the overload relief of the swing motor in the initial stage of the combined simultaneous operation.
At the initial stage of the turning operation, the flow rate flowing into the turning motor depends on the load pressure PLs on the turning side due to the excessive inertia of the turning motor.
Is increased, and the directional control valve differential pressure ΔPs ′ is reduced as shown in the equation (14), that is, the flow rate of the directional control valve is reduced. From this state, the load pressure PLs gradually decreases as the turning speed increases and the acceleration decreases thereafter, so that the directional control valve differential pressure ΔPs' increases and the flow rate of the directional control valve gradually increases. In other words, since the flow rate gradually increases as the swing load pressure decreases, a gentle acceleration of the swing motor can be obtained.

Thereafter, when the acceleration of the swing motor is completed and the swing is performed at a steady speed, the swing load pressure sharply decreases, and the load pressure of the boom cylinder increases. At this time, in the conventional embodiment, the boom cylinder is accelerated because the function of the constant horsepower mechanism is alleviated due to a sudden decrease in the pump discharge pressure, and the pump discharge flow rate increases abruptly. In this embodiment, since the speed of the boom cylinder is secured from the beginning of the turning operation as described above, the speed of the boom cylinder is not accelerated with a shock. Note that the above operation operates continuously as the load pressure of the swing motor decreases, and is not discontinuous as in the above-described conventional example, so that the occurrence of a shock is further reduced.

The hydraulic circuit diagram shown in FIG. 2 shows a second embodiment of the first invention of the present invention, and relates to an improvement of the hydraulic circuit diagram of the second conventional hydraulic drive device shown in FIG. 6 described above. FIG.
The same members as are denoted by the same reference numerals and description thereof will be omitted. 6, a hydraulic circuit diagram shown in FIG. 2 is a hydraulic drive device having a load sensing function having an anti-saturation function. In FIG. 2, a plurality of directional control valves 24 and 25 are connected in parallel to the pump discharge oil line 3 of the variable displacement pump 2 via check valves 26 and 27, and the pump discharge oil has passed through the throttle portions of the directional control valves. Thereafter, the output side of each of the directional control valves 24, 25 is supplied to the actuators 12, 13 respectively, and the return oil from each of the actuators 12, 13 passes through the directional control valves 24, 25 again. , 21 to the tank T via the tank line 16. The pressure compensating valves 20 and 21 are directional control valves 24 and 25, respectively.
And the tank line 16 and the pressure downstream of the throttle portion of the directional control valve, that is, the load pressure PL of each of the actuators 12 and 13, is applied to the first control chambers 20a and 21 in the opening direction of the pressure compensating valve.
a is applied to the pressure receiving area, and the maximum load pressure Pm of the plurality of actuators detected by the shuttle valve 4 is applied to the second control chambers 20b and 21b in the closing direction of the pressure compensating valve to apply the load pressure of the corresponding actuator. On the third control chambers 20c, 21c in the closing direction of the pressure compensating valve. Each of the first and second pressure receiving areas is substantially equal, and each of the third pressure receiving areas is only a few percent (about 0 to 0.07) of the first pressure receiving area.

Further, the ratio between the third pressure receiving area of the high load side actuator 12 and the first or second pressure receiving area is determined by the third pressure receiving area of the low load side actuator 13 and the first or second pressure receiving area. Is larger than the ratio. That is, the ratio is set to about 0.03 to 0.07 in the high load side actuator 12, and is set to about 0 to 0.02 in the low load side actuator 13.

With such a configuration, the hydraulic drive device according to the second embodiment of the first invention of the present invention shown in FIG. 2 is the same as that of the first embodiment of the first invention of the present invention shown in FIG. The operation and effect of In FIG. 2, considering the balance of the forces acting on the pressure compensating valves 20, 21, first, the pressure compensating valves 20, 21
F1 acting in the opening direction of the pressure control valve is expressed as PL, the pressure communicating with the load pressure PL on the downstream side of the directional control valve throttle portion, and the pressure receiving area of the first control chambers 20a, 21a of the pressure compensating valve is Aa. (PL ・ Aa) ...... (21)

Conversely, the force F2 acting in the direction to close the pressure compensating valve is such that the maximum load pressure is Pm, the pressure receiving areas of the second control chambers 20b and 21b of the pressure compensating valve are Ab, The load pressure on the downstream side is PL, and the third control chamber 20 of the pressure compensating valve is
c, 21c Assuming that the pressure receiving area is Ac, F2 = (Pm ・ Ab + PL ・ Ac) ...... (22)

Here, when controlling the pressure compensating valve, since the forces in both directions are balanced, the equations (21) and (22) become equal, so that F1 = F2, and (PL · Aa) = (Pm・ Ab + PL ・ Ac) ...... (23) However, the acting force of the springs 20d and 21d is ignored as being weak. On the other hand, the pressure on the upstream side of the throttle portion of the directional control valve, that is, the pump discharge pressure Pp of the discharge oil passage 3 of the variable displacement pump 2 is
Assuming that Psp is the differential pressure between the pump discharge pressure preset by the spring 18 of the flow control valve 17 and the maximum load pressure due to the action of the flow control valve 17, Pp = Pm + Psp ...... (24) From equation (24), Pm = Pp-Psp ... (25) is obtained.

By substituting equation (25) into equation (23), (PL · Aa) = (Pp−Psp) · Ab + PL · Ac (26) where the first pressure receiving area Aa And the second pressure receiving area Ab are assumed to be equal, and A = Aa = Ab ... (27) Equation (26) gives PL = Pp-Psp + PL.Ac / A ... . (28) From equation (28), the directional control valve differential pressure ΔP = Pp-PL before and after the throttle part of the directional control valve can be calculated as ΔP = Pp-PL = Psp-PLAc / A ...... (29) or Psp = Pp-Pm obtained from Eq. (25) in Eq. (29)
By substituting the relationship, ΔP = Pp−Pm−PL · Ac / A ...... (210) is obtained.

According to the expression (29), the directional control valve differential pressure ΔP before and after the throttle portion of the directional control valve is determined by the spring 18 of the flow regulating valve 17.
Thus, the pressure Psp set in advance and the load pressure PL of the actuator downstream of the directional control valve become a linear expression, and decrease as the load pressure PL of the actuator increases. That is, a downward-sloping pressure compensation characteristic in which the flow rate decreases as the load pressure PL of the actuator increases. According to the equation (210), the directional control valve differential pressure ΔP
Is a linear expression of the differential pressure between the pump discharge pressure Pp and the maximum load pressure Pm and the load pressure PL of the actuator downstream of the directional control valve, and similarly decreases as the load pressure PL of the actuator increases, that is, the actuator As the load pressure PL increases, a downward-sloping pressure compensation characteristic in which the flow rate of the actuator decreases is obtained. Therefore Figure 1
The same operation and effect as those of the first embodiment of the first invention of the present invention shown in FIG.

The hydraulic circuit diagram shown in FIG. 3 shows a second embodiment of the present invention, and the first hydraulic circuit diagram of the present invention shown in FIG.
The hydraulic compensator of the second embodiment of the present invention has pressure compensation valves 30 and 31 which are improved with respect to the hydraulic circuit diagram. As in FIG. 2, the hydraulic circuit diagram shown in FIG. 3 is a hydraulic drive device having a load sensing function having an anti-saturation function. 1 and 2 are denoted by the same reference numerals and description thereof will be omitted. The pressure compensating valves 30 and 31 of the second invention improved in FIG. 3 are disposed between the directional control valves 24 and 25 and the tank line 16, respectively, and the pressure on the downstream side of the throttle portion of the directional control valve is provided. That is, the first control chamber in which the load pressure PL of each of the actuators 12 and 13 acts in the opening direction of the pressure compensating valve.
30a and 31a have a pressure receiving area and second control chambers 30b and 31b for applying the maximum load pressures Pm of the plurality of actuators detected by the shuttle valve 4 in the closing direction of the pressure compensating valve. And the pressure compensating valve of the high load side hydraulic actuator 12
The first pressure receiving area Ba of 30 is 0.93 to 0.97 of the second pressure receiving area Ab.
The first pressure receiving area Ca of the pressure compensating valve 31 of the low load side hydraulic actuator 13 is set to 0.98 to 1.00 of the second pressure receiving area Ab, whereby the corresponding actuator of the pressure compensating valve 30 of the high load side hydraulic actuator 12 is adjusted. In response to the increase of the load pressure, the outlet flow rate of the pressure compensating valve 30 and, consequently, the inflow flow rate to the actuator 12 are reduced to the outlet flow rate of the pressure compensating valve 31 of the low load side hydraulic actuator 13 and, consequently, the inflow rate to the actuator 13. It is designed to decrease more. In short, in FIG. 2, the third control chambers 20c, 21c for receiving the load pressure of the corresponding actuator in the closing direction of the pressure compensating valve are provided with the pressure receiving area Ac, but in FIG. 3, the first control chambers 30a, 31a are provided with the pressure receiving area. Ba and Ca are obtained by subtracting the third pressure receiving area Ac in FIG. 2 from the second pressure receiving area Ab.

With such a configuration, the hydraulic drive system according to the second embodiment of the present invention shown in FIG. 3 is the same as the first embodiment shown in FIG. 1 or the first embodiment shown in FIG. The same operational effects as those of the second embodiment of the first invention of the present invention are obtained.

In FIG. 3, considering the balance of the forces acting on the pressure compensating valve 30, first, the force F1 acting in the opening direction of the pressure compensating valve 30 is expressed by PL as the load pressure on the downstream side of the directional control valve throttle section. , The pressure receiving area of the first control chamber 30a of the pressure compensating valve
Assuming Ba, F1 = (PL · Ba) ...... (31) Conversely, the force F2 acting in the direction to close the pressure compensating valve is the maximum load pressure Pm, the second control of the pressure compensating valve Assuming that the pressure receiving area of the chamber 30b is Ab, F2 = (PmAb) ...... (32)

Here, when controlling the pressure compensating valve, since the forces in both directions are balanced, the equations (31) and (32) become equal, so that F1 = F2, and (PL · Ba) = (Pm・ Ab) ...... (33) However, the acting force of the spring 30d is ignored because it is weak. On the other hand, the pressure on the upstream side of the throttle portion of the directional control valve, that is, the pump discharge pressure Pp of the discharge oil passage 3 of the variable displacement pump 2 adjusts Psp by the action of the above-described variable displacement pump 2 and the flow control valve 17. Assuming that the pressure difference between the pump discharge pressure and the maximum load pressure set in advance by the spring 18 of the valve 17, Pp = Pm + Psp ...... (34) From equation (34), Pm = Pp-Psp ... ... (35) is obtained.

By substituting equation (35) into equation (33), (PL · Ba) = (Pp−Psp) · Ab (36) where the first pressure-receiving area Ba <the second The pressure receiving area Ab is Ba
/ Ab = k (k <1), k · PL = Pp-Psp ...... (37), where k = [1-(1-k)] (37)
The expression is PL · [1-(1-k)] = Pp-Psp PL-PL · (1-k) = Pp-Psp ...... (38) When the directional control valve differential pressure ΔP = Pp-PL before and after the throttle part of the directional control valve is obtained from equation (38), ΔP = Pp-PL = Psp-PL ・ (1-k) ...... (39) Or, Psp = Pp-Pm obtained from Eq. (35) in Eq. (39)
By substituting the relationship, ΔP = Pp−Pm−PL · (1-k) ...... (310) is obtained.

Since k <1, according to equation (39), the directional control valve differential pressure ΔP before and after the throttle portion of the directional control valve is equal to the pressure set in advance by the spring 18 of the flow control valve 17.
Psp and the load pressure PL of the actuator downstream of the directional control valve
And decreases as the load pressure PL of the actuator increases. That is, a downward-sloping pressure compensation characteristic in which the flow rate of the actuator decreases as the load pressure PL of the actuator increases.

According to the equation (310), the directional control valve differential pressure ΔP is a linear expression of the differential pressure between the pump discharge pressure Pp and the maximum load pressure Pm and the load pressure PL of the actuator downstream of the directional control valve. Similarly, the actuator load pressure PL decreases with an increase, ie, the actuator load pressure PL
As the PL increases, a downward-sloping pressure compensation characteristic in which the flow rate of the actuator decreases is obtained. Therefore, the hydraulic drive device according to the second embodiment of the present invention shown in FIG. 3 is the first embodiment of the first invention shown in FIG. 1 or the first embodiment of the present invention shown in FIG. The same operation and effect as those of the second embodiment of the first invention are obtained.

FIG. 4 shows the above-mentioned US Pat.
FIG. 8 is a block diagram showing a cross-sectional structure of improved pressure compensating valves 40 and 41 which can be used in the hydraulic circuit diagram of the conventional hydraulic drive device shown in FIG. 7 disclosed in Japanese Unexamined Patent Publication No. 5-332310 or Japanese Unexamined Patent Publication No. Hei 5-332311. Like FIG. 1, the hydraulic circuit diagram shown in FIG. 7 including the improved pressure compensating valves 40 and 41 is a hydraulic drive device having a load sensing function with an anti-saturation function, and is the same as FIG. 1 to FIG. The members are denoted by the same reference numerals and description thereof is omitted.

The pressure compensating valves 40 and 41 according to the second invention improved in FIG.
Is disposed between the pump discharge pressure oil line 3 and each of the directional control valves 24 and 25, and the pressure compensating valve 40 is capable of restricting the pump discharge oil flow rate to the actuator. Check valve part 74 for preventing backflow to 3 and spool 43 for closing check valve part 74
And the pump discharge pressure Pd from the pump discharge pressure hydraulic line 3 is set to the highest load pressure among the plurality of actuators.
The pressure load PL of the actuator is a pressure reducing valve portion 42 for reducing the pressure to Pm.
In the opening direction of the pressure compensating valve 40.
a Pressure receiving area and maximum load pressure Pm of multiple actuators
Control chamber 44b which acts in the closing direction of the pressure compensating valve
And the first and second pressure receiving areas are substantially equal. Since the second control chamber 44b of the pressure reducing valve section 42 is mutually connected to the second control chambers of the other pressure compensating valves via the maximum load pressure Pm lines 5 of the plurality of actuators, a shuttle valve is shown in FIG. Do not need. In addition, the check valve section 74 determines that the upstream pressure Pz of the direction control valve 24 is the pump discharge pressure.
If the pressure is higher than Pd, the check valve portion 74 is closed. Therefore, it is the same as the conventional pressure compensating valves 70 and 71 shown in FIG. is there.

In the pressure compensating valve 40 shown in FIG. 4, the spool 43 of the pressure reducing valve portion 42 has a small diameter portion 72h extending from the middle diameter portion 72g.
The tip of the small-diameter portion 72h is brought into contact with the check valve portion 74, and the connecting step between the middle-diameter portion 72g and the small-diameter portion 72h of the spool 43 is communicated with the tank line 16 communicating with the tank T. Now, adjust the outside diameter of the medium diameter portion 72g of the spool 43 of the pressure reducing valve portion 42.
d, assuming that the outer diameter of the small-diameter portion 72h is d ′, the pressure-receiving area of the check valve portion 74 to which the upstream pressure Pz of the direction control valve 24 for closing the check valve portion 74 acts is the middle-diameter portion outer diameter d of the spool 43. The area obtained by subtracting the outer diameter d 'area of the small diameter portion 72h from the area (π (d-
d ')), which forms the third control chamber pressure receiving area. Hereafter, the third control room pressure receiving area is increased. Check valve spool 74e of check valve section 74
The pressure Pz applied to the third control chamber pressure receiving area of the pressure receiving area is
When the pressure compensating valve 40 is operated, the pressure becomes the sum of the load pressure PL of the actuator and the differential pressure before and after the throttle portion of the directional control valve 24, and substantially becomes the load pressure PL of the actuator.

More specifically, the check valve portion 74 of the pressure compensating valve 40 shown in FIG. 4 has a check valve spool 74 inserted into the valve hole 74j.
There is provided an important notch 74b, a small notch 74c, and a radius hole 74d for forming a throttle communicating with the spool shaft hole 74k. The pump discharge pressure Pd passes through the radial hole 74d, communicates with the spool shaft hole 74k, and acts on the left side surface of the check valve spool 74e. When the upstream pressure Pz of the direction control valve 24 is higher than the pump discharge pressure Pd, the upstream pressure Pz is increased by the check valve spool 74e.
By pressing, the check valve spool 74e is closed. The pressure reducing valve portion 42 of the pressure compensating valve 40 includes a pressure reducing valve spool 43, a pin 73 inserted into a hole 72i provided in the pressure reducing valve spool 43, and a spring 77d for pressing the pressure reducing valve spool 43 toward the check valve portion 74.
Consists of Pin inserted in hole 72i of pressure reducing valve spool 43
The pump discharge pressure Pd always acts on the left side of the valve 73 through a radial hole 72a forming a throttle, whereby the right side of the pin 73 contacts the left side of the valve. The outer diameter of the pin 73 is the same as the outer diameter d of the middle diameter portion 72g. Spring 77
Although d is a weak force, when there is no load pressure PL and the maximum load pressure Pm of the actuator, the small-diameter portion 72h of the pressure reducing valve spool 43 presses the left side of the small-diameter portion 72h to close the check valve spool 74e. The connecting step between the large diameter portion 72m and the middle diameter portion 72g of the pressure reducing valve spool 43 has a load pressure of the actuator corresponding to the step area A2.
The PL is guided to push the pressure reducing valve spool 43 rightward. The maximum load pressure Pm is the radius hole 72 forming the throttle
The pressure control valve 44 enters the second control chamber 44b of the pressure reducing valve section 42 through c and acts on the right side surface of the pressure reducing valve spool 43 to press the pressure reducing valve spool 43 leftward.

With this configuration, the improved pressure compensating valve 4
The hydraulic circuit diagram shown in FIG. 7 including 0,41 has the same operation and effect as the first embodiment of the first invention of the present invention shown in FIG.

In FIG. 4, considering the balance of the forces acting on the pressure compensating valve 40, first, the force F1 acting rightward for opening the pressure compensating valve 40 (the check valve portion 74 and the pressure reducing valve portion 42) is:
Check valve spool 74e Left side area is A1, pressure reducing valve spool
The step area of the connecting step between the large diameter part 72m and the medium diameter part 72g
As A2, F1 = Pd ・ A1 + PL ・ A2 ...... (41) Conversely, the force F2 acting to the left to close the pressure compensating valve is changed from the right area of the check valve spool 74e to the pressure reducing valve spool 43. The area obtained by subtracting the cross-sectional area of the small-diameter portion 72h is A4,
As A3 F2 = PzA4 + PmA2 + PdA3 ...... (42)

Here, when controlling the pressure compensating valve, since the forces in both directions are balanced, the equations (41) and (42) become equal, so that F1 = F2, and Pd · A1 + PL · A2 = Pz-A4 + Pm-A2 + Pd-A3 (43) By rearranging equation (43), Pz · A4-PL · A2 = Pd · (A1-A3)-Pm · A2 ... (44) Here, as is clear from Fig. 4, A2 = A1-A3 and A2 = k · A4 (k <1), substitute in equation (44),
By dividing by A4, Pz-k · PL = k · (Pd-Pm) ...... (45) Set the PL coefficient k in equation (45) to k = [1-(1-k)] By rearranging and obtaining the directional control valve differential pressure ΔP, ΔP = Pz−PL = k · (Pd−Pm) −PL · (1−k) (46) is obtained. Since Psp = Pd−Pm, equation (46) is expressed as ΔP = k · Psp−PL · (1−k) (47).

According to the equations (46) and (47), it means that the flow rate to the actuator decreases as the own load pressure PL on the downstream side of the directional control valve increases. That is, by using the pressure compensating valve 40 of FIG. 4 in the hydraulic circuit of FIG. 7, similarly to the pressure compensating valve used in the hydraulic circuits of FIGS. As a result, a downward pressure compensation characteristic in which the flow rate to the actuator decreases can be obtained. Therefore, an improved pressure compensating valve 4
The hydraulic circuit diagram shown in FIG. 7 including 0,41 has the same function and effect as the first embodiment of the first invention of the present invention shown in FIG.

Preferably, the value of k (k = A2 / A4) is set to 0.93 to 0.97 for the pressure compensating valve of the high load side actuator (for the swing motor) and 0.97 for the pressure compensating valve of the low load side actuator (for the boom). ~ 1.00, and as the pressure compensating valve
It is preferable to select the outer diameter d 'of the small diameter portion 72h of the pressure reducing valve spool 43 of 40.

Although the embodiment of the present invention has been described above, various other configurations are possible, and various modifications are possible without departing from the spirit of the present invention. It goes without saying that it covers everything.

[0058]

As described above, according to the first and second aspects of the present invention, the rate at which the outlet flow rate of the pressure compensating valve communicating with the high-load hydraulic actuator is reduced is communicated with the low-load hydraulic actuator. By increasing the outlet flow rate of the pressure compensating valve from the rate of decrease, when the load pressure of the high load side actuator suddenly increases, the flow rate to the high load side actuator decreases, and the reduced flow rate decreases to the low load side. Since the speed of the low-load side actuator supplied to the actuator can be prevented from decreasing, even if the high-load side actuator and the low-load side actuator with extremely different loads are simultaneously operated, the load pressure of the high-load side actuator can be further reduced. Even if the value suddenly decreases, there is no sudden change in actuator speed or Rukoto could be. In addition, the speed reduction of the low-load-side actuator can be eliminated.
After that, even if the acceleration of the high-load-side actuator is completed and the steady-state speed is reached, the speed of the low-load-side actuator is equal to the flow rate of the decrease in the flow rate to the high-load-side actuator from the beginning of the operation of the high-load-side actuator Since the speed of the low-load-side actuator is supplied to the low-load-side actuator, it is not accelerated with a shock.

Further, when the load pressure of the high-load-side actuator suddenly increases from the beginning of the operation of the high-load-side actuator, the supply flow rate of the high-load-side actuator decreases. And a reduction in the discharge flow rate can be suppressed, and a wasteful relief flow rate from the overload relief can be reduced, so that energy loss can be reduced.

Further, as means for achieving the object, various accessory valves are provided in addition to the pressure compensating valve as in the related art, and no external pilot pressure is required, so that the overall size of the valve does not increase. At the same time, there is an excellent effect that the cost is low and the usability is good.

[Brief description of the drawings]

FIG. 1 (a) is a hydraulic circuit diagram showing a first embodiment of the first invention of the present invention, and FIG. 1 (b) is a discharge oil flow rate change device of an embodiment different from FIG. 1 (a). FIG. 4 is a partial hydraulic circuit diagram using constant horsepower control devices 19 and 6 and a bleed-off valve 17 ′ instead of a certain flow regulating valve 17.

FIG. 2 is a hydraulic circuit diagram showing a second embodiment of the first invention of the present invention.

FIG. 3 is a hydraulic circuit diagram showing a second embodiment of the present invention.

FIG. 4 shows a cross-sectional structure of an improved pressure compensating valve showing a third embodiment of the first invention of the present invention, which can be used in the hydraulic circuit diagram of the third conventional hydraulic drive device of FIG. It is a block diagram.

FIG. 5 is a hydraulic circuit diagram of a first conventional hydraulic drive device.

FIG. 6 is a hydraulic circuit diagram of a second conventional hydraulic drive device.

FIG. 7 is a hydraulic circuit diagram of a third conventional hydraulic drive device.

8 is a block diagram showing a sectional structure of a pressure compensating valve used in a hydraulic circuit diagram of a third conventional hydraulic drive device of FIG.

[Explanation of symbols]

 2 Variable displacement pump 10,11,20,21,30,31,40,41 Pressure compensating valve 12 High load side hydraulic actuator 13 Low load side hydraulic actuator 14,15,24,25 Directional control valve 19 Constant horsepower control valve ( Constant horsepower control device) 17 Flow control valve (discharge oil flow changing device) 42 Pressure reducing valve section 43 Pressure reducing valve spool 74 Check valve section 74e Check valve spool ΔP Direction control valve differential pressure PL Actuator load pressure Pm Actuator maximum load pressure Pp Variable Displacement pump discharge pressure

Claims (12)

(57) [Claims] A first hydraulic actuator driven by a discharge oil and each having a load pressure, and a flow rate adjustment capable of controlling the discharge oil flowing into the first and second hydraulic actuators, respectively. A first and a second directional control valve having a function; and a first pressure compensating valve connected to the first directional control valve for compensating the pressure of the first directional control valve, wherein Pressure compensating valve causes the pressure downstream of the throttle portion of the corresponding first directional control valve to act on the first control chamber pressure receiving area of the first pressure compensating valve in the opening direction, and The maximum load pressure of the first pressure compensating valve to the second
The first pressure compensating valve acts on the third control chamber pressure receiving area of the first pressure compensating valve in the closing direction. The control room pressure receiving area and the second control room pressure receiving area are substantially equal, and the third control room pressure receiving area is only a small area with respect to those control room pressure receiving pressure areas. A first pressure compensating valve configured to decrease the outlet flow rate of the first pressure compensating valve in response to an increase in the load pressure of the actuator; and a variable for discharging the discharge oil to the first and second actuators. A hydraulic drive device comprising: a displacement pump; a constant horsepower control device provided in the variable displacement pump; and the discharge oil flow rate changing device cooperating with the constant horsepower control device. 2. The third pressure receiving area of the first pressure compensating valve is 0.03 of the first pressure receiving area of the first pressure compensating valve.
The hydraulic drive device according to claim 1, wherein the value is from 0.07 to 0.07. 3. At least one low load side first hydraulic actuator driven by discharge oil and having a load pressure, and at least one high load side second hydraulic actuator having an inertial load greater than said actuator. A first having a flow rate adjusting function capable of controlling the discharge oil flowing into the first and second actuators, respectively;
First and second directional control valves, and first and second pressure compensating valves connected to the first and second directional control valves, respectively, for compensating the pressure of the first and second directional control valves. Each of the pressure compensating valves causes the pressure downstream of the throttle section of the corresponding directional control valve to act on the first control chamber pressure receiving area of the pressure compensating valve in an opening direction, and the highest pressure among the plurality of actuators The load pressure is applied to the second control chamber pressure receiving area of the pressure compensating valve in the closing direction, and the pressure communicating with the load pressure of the corresponding actuator is closed to the third control chamber pressure receiving area of the pressure compensating valve. The first control chamber pressure receiving area and the second control chamber pressure receiving area are made substantially equal to each other, and the third control chamber pressure receiving area is only a small area with respect to those control chamber pressure receiving areas. This increases the load pressure on the corresponding actuator. A first and a second pressure compensating valve configured to reduce an outlet flow rate of the pressure compensating valve in response to the pressure increase; a variable displacement pump that discharges the discharge oil to the first and second actuators; A hydraulic drive device comprising: a constant horsepower control device provided in a variable displacement pump; and the discharge oil flow rate changing device cooperating with the constant horsepower control device. 4. The rate at which the outlet flow rate of the pressure compensating valve communicating with the high load side hydraulic actuator is reduced, is greater than the rate at which the outlet flow rate of the pressure compensating valve communicating with the low load side hydraulic actuator is reduced. The hydraulic drive device according to claim 3, wherein: 5. The third pressure receiving area of the pressure compensating valve of the high load side hydraulic actuator is 0.03 to 0.07 of the first pressure receiving area of the pressure compensating valve, and the third pressure receiving area of the pressure compensating valve of the low load side hydraulic actuator is 5. The hydraulic drive device according to claim 4, wherein the pressure receiving area is 0 to 0.02 of the first pressure receiving area of the pressure compensating valve. 6. The first pressure compensating valve is disposed between the first directional control valve and the corresponding first actuator, and is located downstream of a throttle section of the first directional control valve. A first control chamber pressure receiving area for causing the first pressure compensating valve to act in the opening direction of the first pressure compensating valve, and a second control chamber for causing the highest load pressure of the plurality of actuators to act in the closing direction of the first pressure compensating valve. 3. A pressure receiving area of the control chamber, and a third pressure receiving area of the control chamber for applying a load pressure of the first actuator in a closing direction of the first pressure compensating valve. Hydraulic drive. 7. Each of the pressure compensating valves is disposed between the directional control valve and a corresponding actuator, and applies a pressure downstream of a throttle portion of the directional control valve in an opening direction of the pressure compensating valve. A first control chamber pressure receiving area, a second control chamber pressure receiving area for applying the highest load pressure of the plurality of actuators in the closing direction of the pressure compensating valve, and a load pressure of the corresponding actuator in the closing direction of the pressure compensating valve. 6. The hydraulic drive device according to claim 5, further comprising: a third control chamber pressure receiving area that acts on the hydraulic control device. 8. The first pressure compensating valve is disposed between the tank and the first actuator corresponding to the first directional control valve, and is provided with a throttle section of the first directional control valve. A first control chamber pressure receiving area for applying a downstream pressure in the opening direction of the first pressure compensating valve, and a maximum load pressure of the plurality of actuators acting in a closing direction of the first pressure compensating valve. 3. The pressure receiving area according to claim 2, wherein the pressure receiving area has a second control chamber pressure receiving area and a third control chamber pressure receiving area for applying a load pressure of a corresponding actuator in a closing direction of the first pressure compensating valve. Hydraulic drive. 9. Each of the pressure compensating valves is disposed between a corresponding directional control valve and a tank, and applies a pressure downstream of a throttle portion of the directional control valve in an opening direction of the pressure compensating valve. A control chamber pressure receiving area, a second control chamber pressure receiving area for applying the maximum load pressure of the plurality of actuators in the closing direction of the pressure compensating valve, and a load pressure of the corresponding actuator in the closing direction of the pressure compensating valve. 6. The hydraulic drive device according to claim 5, further comprising: a third control chamber pressure receiving area that acts on the hydraulic control device. 10. At least one low load side first hydraulic actuator driven by discharge oil and having a load pressure, and at least one high load side second hydraulic actuator having an inertial load greater than said actuator. A first having a flow rate adjusting function capable of controlling the discharge oil flowing into the first and second actuators, respectively;
First and second directional control valves, and first and second pressure compensating valves connected to the first and second directional control valves, respectively, for compensating the pressure of the first and second directional control valves. The pressure compensating valve is disposed between the corresponding directional control valve and the tank, and each of the pressure compensating valves controls the pressure downstream of the throttle portion of the corresponding directional control valve in the first control chamber pressure receiving area of the pressure compensating valve. In the opening direction, the highest load pressure of the plurality of actuators is applied to the second control chamber pressure receiving area of the pressure compensating valve in the closing direction, and the pressure compensating valve on the high load side second actuator side is actuated. The first pressure receiving area is 0.92 to 0.97 of the second pressure receiving area, and the first pressure receiving area of the pressure compensating valve on the low load side first actuator side is 0.98 to 1.00 of the second pressure receiving area. Valve and thereby the pressure on the high load side second actuator side The outlet flow rate of the pressure compensating valve corresponding to the increase in the load pressure of the high load side second actuator corresponding to the compensation valve is more than the outlet flow rate of the pressure compensating valve on the low load side first actuator side. A first and a second pressure compensating valve configured to reduce the pressure, a variable displacement pump for discharging the discharge oil to the first and second actuators, and a constant horsepower control device provided in the variable displacement pump. And a discharge oil flow rate changing device that cooperates with the constant horsepower control device. 11. The first pressure compensating valve is disposed between the pump and the first directional control valve, and the first pressure compensating valve is provided between the pump and the first directional control valve.
The pressure compensating valve has a check valve portion capable of preventing a backflow from the first actuator to the pump and reducing a flow rate to the first actuator, a spool closing the check valve portion, and discharging the pump. A first control chamber pressure receiving area for reducing a hydraulic pressure to the maximum load pressure, the first control chamber pressure receiving area for applying a pressure downstream of a throttle portion of the directional control valve in an opening direction of a pressure compensating valve, and A second control chamber pressure receiving area for applying pressure in the closing direction of the pressure compensating valve, and a third control section for applying an outlet pressure of the pressure compensating valve, which is a pressure communicating with the load pressure of the actuator, in the closing direction of the pressure compensating valve. 3. The hydraulic drive device according to claim 2, comprising: a control chamber pressure receiving area. 12. Each of the pressure compensating valves is disposed between the pump and each of the directional control valves, and the pressure compensating valve prevents a backflow from the actuator to the pump and controls a flow rate to the actuator. A pressure reducing valve having a check valve portion that can be throttled and a spool that closes the check valve portion and that can reduce the pump discharge oil pressure to the maximum load pressure, and a pressure downstream of the throttle portion of the directional control valve. A first control chamber pressure receiving area for acting in the opening direction of the pressure compensating valve, a second control chamber pressure receiving area for applying the maximum load pressure in the closing direction of the pressure compensating valve, and a load pressure of the actuator. 6. The hydraulic drive device according to claim 5, further comprising: a third control chamber pressure receiving area for causing an outlet pressure of the pressure compensating valve to act in a closing direction of the pressure compensating valve.