JP2615207B2 - Hydraulic drive - Google Patents

Description

Description: BACKGROUND OF THE INVENTION The present invention relates to a hydraulic drive device for construction equipment such as a hydraulic excavator, and more particularly, to a hydraulic drive device for controlling a differential pressure across a flow control valve. A hydraulic pressure for applying a control force based on a differential pressure between a discharge pressure of a hydraulic pump controlled by load sensing and a maximum load pressure of a plurality of actuators to the compensating valve to set a target value of a differential pressure across the flow control valve. It relates to a driving device.

[Conventional technology]

2. Description of the Related Art In recent years, in a hydraulic drive device of a construction machine including a plurality of hydraulic actuators for driving a plurality of driven bodies, such as a hydraulic shovel and a hydraulic crane, a discharge pressure of a hydraulic pump is controlled in conjunction with a load pressure or a required flow rate. At the same time, a pressure compensating valve is arranged in relation to the flow control valve, and the pressure compensating valve controls the differential pressure across the flow control valve to stably control the supply flow rate during combined driving. I have. Among them, load sensing control is a typical example of controlling the discharge pressure of the hydraulic pump in conjunction with the load pressure.

Load sensing control is to control the discharge amount of the hydraulic pump so that the discharge pressure of the hydraulic pump is higher than the maximum load pressure of the plurality of hydraulic actuators by a certain value. By increasing or decreasing the discharge amount of the hydraulic pump, economical operation becomes possible.

By the way, since the discharge amount of the hydraulic pump has an upper limit, that is, the maximum possible discharge amount, when the hydraulic pump reaches the maximum possible discharge amount during the combined driving of a plurality of actuators, an insufficient pump discharge state occurs. This is commonly known as hydraulic pump saturation. When the saturation occurs, the pressure oil discharged from the hydraulic pump flows preferentially to the actuator on the low pressure side, and sufficient pressure oil is not supplied to the actuator on the high pressure side, so that the combined driving of a plurality of actuators cannot be performed.

In order to solve such a problem, DE-A1-3422165
In the hydraulic drive device described in Japanese Patent Application Laid-Open No. 60-11706, each pressure compensating valve for controlling the differential pressure across the flow control valve is
Instead of a spring for setting the target value of the differential pressure, two driving units acting in the valve opening direction and the valve closing direction are provided, and the discharge pressure of the hydraulic pump is guided to the driving unit acting in the valve opening direction, and the valve closing direction is set. The maximum load pressure of the plurality of actuators is guided to the drive unit acting on the actuator, and a control force based on the differential pressure between the pump discharge pressure and the maximum load pressure is applied in the valve opening direction. It is determined. With this configuration, when saturation of the hydraulic pump occurs, the differential pressure between the pump discharge pressure and the maximum load pressure correspondingly decreases, so that the target value of the differential pressure across the flow control valve in each pressure compensating valve also decreases. Accordingly, the pressure compensating valve relating to the low pressure side actuator is further throttled, and the pressure oil from the hydraulic pump is prevented from flowing preferentially to the low pressure side actuator.
Thereby, the pressure oil from the hydraulic pump is diverted in accordance with the ratio of the required flow rate (valve opening) of the flow control valve and supplied to a plurality of actuators, thereby enabling appropriate combined driving. In this configuration, as a result, the pressure compensating valve has a function of reliably shunting the pressure oil from the hydraulic pump and supplying it to the plurality of actuators regardless of the discharge state of the hydraulic pump. In this document, this function is called “shunt compensation” for convenience, and the pressure compensation valve is called “shunt compensation valve”.

[Problems to be solved by the invention]

By the way, in this conventional hydraulic drive device, each of the branch flow compensating valves sets a difference between the discharge pressure of the hydraulic pump to be subjected to load sensing control and the maximum load pressure of the plurality of actuators as a target value of the differential pressure across the flow control valve. The control force based on the pressure is applied, so that if the pressure receiving areas of all the drive units are the same, the control force applied to each of the shunt compensating valves is the same, and the pressure compensation characteristics of all the shunt compensating valves Is the same. For this reason, for example, when performing a composite operation of simultaneously driving two or more actuators, the distribution ratio of the flow rate supplied to the actuators, that is, the distribution ratio, of the flow control valve does not depend on the combination of the actuators of the composite operation. It is uniquely determined according to the opening ratio, and depending on the type of compound operation, the flow rate distribution to one actuator is too large or too small, and there is a problem that operability and / or work efficiency is reduced. Was.

SUMMARY OF THE INVENTION It is an object of the present invention to provide a hydraulic drive device for a construction machine that can provide individual pressure compensation characteristics to a flow compensating valve and improve operability and / or work efficiency.

[Means for solving the problem]

In order to achieve the above object, according to the present invention, a hydraulic pump, at least first and second hydraulic actuators driven by pressure oil supplied from the hydraulic pump, and a first and a second actuator are provided. First and second flow control valves for controlling the flow of the supplied pressure oil, respectively, and a first differential pressure for controlling a first differential pressure generated between an inlet and an outlet of the first and second flow control valves, respectively. 1st and 2nd
And the discharge pressure of the hydraulic pump and the first
And discharge amount control means for controlling a flow rate of hydraulic oil discharged from the hydraulic pump in response to a second differential pressure between the maximum load pressure of the second actuator and the first and second branch flow compensation. The valve is a hydraulic drive device for a construction machine having a drive unit configured to apply a control force based on the second differential pressure to a corresponding shunt compensation valve and set a target value of the first differential pressure. A first means for obtaining the second differential pressure from a discharge pressure of a hydraulic pump and a maximum load pressure of the first and second actuators, and at least a second differential pressure obtained by the first means. A second means for calculating an individual value as a value of the control force to be applied by each drive means of the first and second flow dividing compensation valves; and And second control pressures provided corresponding to A raw unit, respectively, a control pressure corresponding to the individual values obtained in the second means occurs,
There is provided a hydraulic drive device having the first and second control pressure generating means for outputting the control pressure to the drive means for the first and second flow compensating valves, respectively.

In one aspect of the present invention, the second means comprises a first and a second preset pressure corresponding to the second differential pressure determined by the first means and the first and second flow compensating valves. And a first calculating means for obtaining values of the first and second control forces corresponding to the second differential pressure from the following function:

At this time, when the first actuator is an actuator that drives an inertial load and the second actuator is an actuator that drives a normal load, preferably, the first and second functions are The relationship between the second differential pressure and the values of the first and second control forces is such that the target value of the first differential pressure decreases as the differential pressure of No. 2 decreases and the rate of decrease differs between the two. Stipulated.

When the first actuator is an actuator for driving an inertial load and the second actuator is an actuator for driving a normal load, preferably, at least the first function related to the first actuator is: The relationship between the second differential pressure and the value of the first control force is determined so that when the second differential pressure exceeds a predetermined value, an increase in the target value of the first differential pressure is suppressed. ing.

When the first and second actuators are actuators for traveling, preferably, both the first and second functions are such that the target value of the first differential pressure is greater than the second differential pressure. The relationship between the second differential pressure and the values of the first and second control forces is determined so that the pressure difference also increases.

The first actuator is one of the actuators for traveling.
In the case where the second actuator is an actuator for excavation work, preferably, the second means is configured to detect a change in the value of the first control force obtained from the first function. Further comprises a second calculating means for giving a relatively large time delay and giving a relatively small time delay to a change in the value of the second control force obtained from the second function.

In the case where the first actuator is a hydraulic motor and the second actuator is a hydraulic cylinder, the hydraulic drive device of the present invention preferably has a third function of detecting the temperature of the hydraulic oil discharged from the hydraulic pump. Means, wherein the second means calculates a temperature correction coefficient from the pressure oil temperature detected by the third means and a third function set in advance; and And a fourth calculating means for calculating the value of the second control force obtained from the function of 2 and the temperature correction coefficient to correct the value of the second control force.

In another aspect of the present invention, the hydraulic drive device of the present invention is operated from the outside and outputs a selection command signal according to the type of work or the content of work performed by driving the first and second actuators. Further comprising a fourth means for performing
The second means includes: a second differential pressure obtained by the first means; fourth and fifth functions respectively set in advance corresponding to the first and second branch flow compensating valves; There may be provided a fifth calculating means for obtaining values of the third and fourth control forces from the selection command signal outputted from the fourth means.

In this case, preferably, the fifth arithmetic means includes a plurality of functions having different characteristics as the fourth and fifth functions, respectively, and a plurality of functions are provided in accordance with the selection command signal output from the fourth means. Select one of the functions
From the second differential pressure obtained by the first means and the selected function, third and fourth control force values corresponding to the second differential pressure are obtained.

In still another aspect of the present invention, when the first actuator is an actuator for driving an inertial load and the second actuator is an actuator for driving a normal load, the hydraulic drive device according to the present invention includes the hydraulic pump 5th means for detecting the discharge pressure of
Means for obtaining a fifth pressure corresponding to the second differential pressure from the second differential pressure obtained by the first means and a preset sixth function.
A sixth calculating means for determining the value of the control force of the above, and using this as a value of the control force to be applied by the driving means of the first branching compensation valve; and setting the discharge pressure detected by the fifth means in advance. A value of a sixth control force for holding the discharge pressure at a predetermined value is obtained from the obtained seventh function, and the target value of the first differential pressure is different between the fifth control force and the sixth control force. The larger one is the second
And a seventh calculating unit that sets the value of the control force to be applied by the driving unit of the shunt compensation valve.

In this case, the hydraulic drive device of the present invention further includes sixth means which is externally operated and outputs a selection command signal relating to the predetermined value of the discharge pressure, and wherein the seventh calculation means is configured to output the selection command Changing the characteristics of the seventh function by a signal;
The predetermined value of the discharge pressure may be changeable.

Further, in another aspect of the present invention, the first actuator is an actuator for driving an inertial load, and the second actuator is
When the actuator of the first type is an actuator for driving a normal load, the hydraulic drive device of the present invention comprises: a seventh unit for detecting the drive of the first actuator;
Eighth means for setting a rate of increase in the flow rate of the pressure oil supplied through the diversion compensating valve of the above, wherein the second means comprises:
The second differential pressure obtained by the first means and a preset eighth pressure
The value of the seventh control force corresponding to the second differential pressure is obtained from the function of the above equation, and this value is used as the value of the control force to be applied by the drive means of the second branching compensation valve. And the seventh
When the start of the driving of the first actuator is detected by the means, the eighth control which changes at a speed equal to or less than a change amount corresponding to the flow rate increasing speed with the value of the seventh control force as a target value. The value of the force is obtained, and this eighth control force is
And a ninth calculating unit that sets the value of the control force to be applied by the driving unit of the shunt compensation valve.

In this case, the hydraulic drive device of the present invention further includes ninth means for detecting the drive of the second actuator, and the ninth arithmetic means uses the seventh and ninth means to perform the first operation. The value of the eighth control force may be obtained when the start of driving of the second actuator is detected.

In still another aspect of the present invention, the hydraulic drive device of the present invention further includes a tenth means for detecting a discharge pressure of the hydraulic pump, wherein the second means is determined by the first means. A tenth calculating means for calculating, from the second differential pressure, a target differential pressure discharge amount of the hydraulic pump for maintaining the differential pressure constant;
Eleventh calculating means for calculating an input restriction target discharge amount of the hydraulic pump from the discharge pressure detected by the means of 10 and a preset input restriction function of the hydraulic pump, and calculating the differential pressure target discharge amount and the input restriction target discharge amount. A thirteenth calculating means for calculating a deviation, wherein when the input restricted target discharge amount is selected as the discharge amount target value of the hydraulic pump among the differential pressure target discharge amount and the input restricted target discharge amount, the deviation of the target discharge amount And a thirteenth calculating means for calculating an individual value as a value of the control force to be applied by each of the driving means of the first and second flow compensating valves based on the above.

In still another aspect of the present invention, preferably, the hydraulic drive device of the present invention is provided on the first and second flow compensating valves, and urges each of the flow compensating valves in the valve opening direction. The driving means further includes a driving means different from the driving means described above, and a pilot pressure supply means for introducing a substantially constant common pilot pressure to the other driving means. It is arranged on the side that urges the first and second branch flow compensating valves in the valve closing direction.

[Action]

In the present invention thus configured, the second means separates individual values as control force values to be applied by the respective driving means of the first and second branching compensation valves based on the second differential pressure. Is calculated in the first and second control pressure generation means, and the control pressure is generated in accordance with these individual values, and the control pressure is output to the drive means of the first and second flow compensating valves, respectively. Thereby, the first and second shunt compensation valves are provided with individual pressure compensation characteristics, and in the combined operation of simultaneously driving the first and second actuators, an optimum shunt ratio according to the type of the actuator is obtained. Operability and / or
Or work efficiency can be improved.

〔Example〕

Hereinafter, a case where a preferred embodiment of the present invention is applied to a hydraulic excavator will be described with reference to the drawings.

First Embodiment First, a first embodiment of the present invention will be described with reference to FIGS.

In FIG. 1, a hydraulic drive device applied to the hydraulic excavator of the present embodiment includes a prime mover 21 and one variable displacement hydraulic pump driven by the prime mover 21, that is, a main pump 22.
And a plurality of actuators driven by pressurized oil discharged from the main pump 22, namely, a swing motor 23, a left running motor 24, a right running motor 25, a boom cylinder 26, an arm cylinder 27, and a bucket cylinder 28, Flow control valves for controlling the flow of the pressure oil supplied to each of the actuators, that is, the turning direction switching valve 29, the left traveling direction switching valve 30, the right traveling direction switching valve 31, the boom direction switching valve
32, arm directional switching valve 33, bucket directional switching valve 34
And are arranged upstream thereof corresponding to these flow control valves,
The differential pressure generated between the inlet and the outlet of the flow control valve, that is, the differential pressure ΔPv1, ΔPv2, ΔPv3, ΔPv4, ΔPv5, ΔPv6 of the flow control valve.
Pressure compensating valves, that is, shunt compensating valves 35, 36,
37,38,39,40.

Further, the hydraulic drive device of the present embodiment responds to the differential pressure ΔPLS between the discharge pressure Ps of the main pump 22 and the maximum load pressure Pamax of the actuators 23 to 28 within a range until the main pump 22 reaches the maximum possible discharge amount. In addition, a load sensing control type discharge amount control device 41 is provided which controls the discharge amount of the main pump 22 so that the discharge pressure Ps becomes higher than the differential pressure ΔPLS by a constant value.

Each of the flow control valves 29 to 34 has an actuator 23
Load lines 43 with check valves 42a, 42b, 42c, 42d, 42e, 42f for taking out their load pressure when driving ~ 28
a, 43b, 43c, 43d, 43e, 43f are connected and these load lines 4
3a to 43f are further connected to a common maximum load line 44.

Each of the branch flow compensating valves 35 to 40 is configured as follows. The branch pressure compensating valve 35 is guided by the outlet pressure of the turning direction switching valve 29, and is driven by the driving unit 35a that urges the valve body of the branch flow compensating valve 35 in the valve opening direction and the inlet pressure of the turning direction switching valve 29. A driving unit 35b for urging the valve body of the branch flow compensating valve 35 in the valve closing direction;
A spring 45 for urging the valve body of the branch flow compensating valve 35 in the valve opening direction with a force f.
And a control pressure Pc1 described later via the pilot line 51a.
Is introduced, and the control force Fc1 is moved in the valve closing direction of the shunt compensation valve 35 in the valve closing direction.
And a driving unit 35c biased by the driving unit 35a.
1 is applied in the valve closing direction, and a spring 45
The second control force f-Fc1 is applied to the valve body of the branch flow compensating valve 35 in the valve opening direction by the drive unit 35c and the first control force and the second control force.
The throttle amount of the diversion compensating valve 35 is determined by the balance of the control forces, and the differential pressure ΔPv1 across the turning direction switching valve 23 is controlled. Here, the second control force f-Fc1 is the turning direction switching valve 2
This is a value for setting the target value of the differential pressure ΔPv1 of 3 before and after.

The other diversion compensating valves 36 to 40 are similarly configured.
In other words, the branch flow compensating valves 36 to 40
Opposing driving units 36a, 36b; 37a, 37b; 38a, each of which is energized by the first control force based on the front-back differential pressure ΔPv2 to ΔPv6 of 30 to 34,
38b; 39a, 39b; 40a, 40b, springs 46, 47, 58, 59, 50 for urging the valve body in the valve opening direction with a force f, and pilot lines 51b, 51c,
Control pressures Pc2, Pc3, P similarly described later via 51d, 51e, 51f
c4, Pc5, Pc6 are guided, and control force Fc2, Fc3,
Drive units 36c, 37c, 38c, 3 that urge in the valve closing direction at Fc4, Fc5, Fc6
9c and 40c.

The discharge amount control device 41 includes a hydraulic cylinder device 52 that drives the swash plate 22a of the main pump 22 to control the displacement, and a control valve 53 that controls the displacement of the hydraulic cylinder device 52. A spring 54 for setting a pressure difference ΔPLS between the discharge pressure Ps of the main pump 22 and the maximum load pressure Pamax of the actuators 23 to 28, and a maximum load pressure Pa of the actuators 23 to 28
a drive 56 in which max is guided via line 55 and a main pump 22
And a drive unit 58 through which the discharge pressure Ps is guided through a conduit 57. When the maximum load pressure Pamax increases, the control valve 53 is driven to the left in the drawing in response thereto, and the hydraulic cylinder device 52 is driven to the left in the drawing to increase the displacement of the main pump 22 to increase the discharge amount. . This allows the main pump
The discharge pressure Ps of 22 is maintained at a higher pressure by a constant value determined by the spring 54.

The hydraulic drive device according to the present embodiment further includes a main pump
22 discharge pressure Ps and maximum load pressure of actuators 23 to 28
And a differential pressure detector 59 that detects the pressure difference ΔPLS between the two and outputs a corresponding electric signal X1, and detects the temperature Th of the pressure oil discharged from the main pump 22, and detects the corresponding electric signal.
A temperature detector 60 that outputs X2 and electric signals X1 and X2 from the differential pressure detector 60 and the temperature detector 61 are input, and the detected differential pressure Δ
The controller 61 calculates the values of the above-described control forces Fc1 to Fc6 based on the PLS and the oil temperature Th, and outputs the corresponding electric signals a, b, c, d, e, f, and the shunt compensation valves 35 to 40. Electromagnetic proportional pressure-reducing valves 62a, 62b, 62c, 62d, 62e, 62 respectively provided correspondingly and inputting electric signals a, b, c, d, e, f from the controller 61.
f, a pilot pump 63 that supplies a pilot pressure to the electromagnetic proportional pressure reducing valves 62a to 62f, and a relief valve that regulates the magnitude of the pilot pressure output from the pilot pump 63
And a control pressure generation circuit 65 including a control pressure generation circuit 64. The electromagnetic proportional pressure reducing valves 62a to 62f are operated by the electric signals a to f to generate control pressures Pc1 to Pc6 corresponding to the values of the control forces Fc1 to Fc6 calculated by the controller 61, and to generate the control pressures Pc1 to Pc6 by using the pilot lines 51a to 51f.
To the drive units 35c to 40c of the branch flow compensation valves 35 to 40, respectively.

The electromagnetic proportional pressure reducing valves 62a to 62f and the relief valve 64 are preferably configured as an aggregate in one block as indicated by a two-dot chain line 66.

The controller 61, as shown in FIG.
1, an input unit 70 for inputting X2, a storage unit 71, and an operation for calculating the values of the control forces Fc1 to Fc6 according to the control program stored in the storage unit using the function data stored in the storage unit 71. It includes a unit 72 and an output unit 73 that outputs the value of the control force obtained by the calculation unit 72 as electric signals a to f.

FIG. 3 is a functional block diagram showing the contents of the calculation performed by the calculation unit 72 of the controller 61. In the figure, blocks 80 to 85
Are provided corresponding to the branch flow compensating valves 35 to 40, and the differential pressure ΔPLS
Is a function block in which function data including a functional relationship between the control force Fc1 and the control force Fc6 are stored in advance, and the control force values Fc1 to Fc6 corresponding to the differential pressure ΔPLS based on the electric signal X1 at that time are obtained from these function blocks. . A block 86 is a function block for temperature correction in which function data including a functional relationship between the oil temperature Th and the correction coefficient K is stored in advance, and a correction coefficient K corresponding to the oil temperature Th based on the electric signal X2 is obtained from the function block 86. Ask for. The correction coefficient K obtained in the function block 86 is multiplied by the values of the control forces Fc4 to Fc6 obtained in the function blocks 83, 84, 85 in multiplication blocks 87, 88, 89, and the values of these control forces are temperature corrected. Control force value Fc found in function blocks 80, 81, and 82
1, Fc2, Fc3 and the control force values Fc4, Fc5, Fc6 temperature-corrected in the multiplication blocks 87, 88, 89 are delay blocks 90 to 9, respectively.
After being filtered by a first-order lag element at 5, the signals are output as electric signals a to f.

Differential pressure ΔPLS and control force Fc stored in function blocks 80 to 85
The functional relationships of 1 to Fc6 are shown in FIGS. 4A to 4D and FIG.

FIG. 4A shows a functional relationship between the differential pressure ΔPLS and the value of the control force Fc1 to be applied to the shunt compensation valve 35 related to the swing motor 23. Here, ΔPLS0 is the main pump held by the discharge amount control device 41 of the load sensing control method.
The pressure difference between the discharge pressure of 22 and the maximum load pressure, that is, the load sensing compensation differential pressure set by the spring 54 of the control valve 53,
f0 is the value of the control force Fc1 corresponding to the load sensing compensation differential pressure ΔPLS0. A is the minimum differential pressure that determines the maximum speed of the swing motor 23, that is, the maximum flow compensation differential pressure related to the swing motor 23, and fc is the maximum flow compensation control force corresponding to the maximum flow compensation differential pressure A. f is the force of the spring 45. Note that f−f0 corresponds to the second control force applied to the shunt compensation valve 35 when the load sensing compensation differential pressure ΔPLSO is secured, and this value is determined by the turning direction set thereby. The target value of the differential pressure ΔPv1 of the switching valve 23 is determined so as to substantially coincide with the load sensing compensation differential pressure ΔPLSO.

In FIG. 4A, the two-dot chain line gives a control force equal to the force f of the spring 45 when the differential pressure ΔPLS is zero, and the differential pressure ΔPLS
Shows the characteristics of the basic function that gradually reduces the control force as increases. Then, the functional relationship between the differential pressure ΔPLS and the control force Fc1 is such that when the differential pressure ΔPLS is smaller than the maximum flow compensation differential pressure A, the value of the control force Fc1 increases according to the increase of the differential pressure ΔPLS according to the characteristics of the basic function. The function is such that the control force fc is output steadily regardless of the increase in the pressure difference ΔPLS when the pressure difference ΔPLS gradually decreases and becomes equal to or higher than the maximum flow rate compensation pressure difference A. Also,
When the pressure difference ΔPLS falls below the minimum flow compensation pressure difference B, the pressure difference Δ
The maximum value fmax below the force f of the spring 45 regardless of the decrease in PLS
The relationship is restricted to.

FIG. 4B shows a functional relationship between the differential pressure ΔPLS and the values of the control forces Fc2 and Fc3 to be applied to the flow dividing valves 36 and 37 relating to the traveling motors 24 and 25. Here, the two-dot chain line shows the characteristics of the basic function as in FIG. 4A, and the differential pressure ΔPLS and the control forces Fc2, Fc
The functional relationship of 3 is that the values of the control forces Fc2 and Fc3 gradually decrease in accordance with the increase in the differential pressure ΔPLS with a gradient smaller than the gradient of the basic function, and the corrected flow rate Δ
Q is obtained.

FIG. 4C shows a functional relationship between the differential pressure ΔPLS and the value of the control force Fc4 to be applied to the flow dividing control valve 38 relating to the boom cylinder 26. The functional relationship is the control force Fc2, Fc3
The control force Fc gradually increases in accordance with the increase in the differential pressure ΔPLS with a slope smaller than the slope of the basic function as compared with the slope of the functional relationship of
The value of 4 decreases.

FIG. 4D shows a functional relationship between the differential pressure ΔPLS and the values of the control forces Fc5 and Fc6 to be applied to the branch flow compensating valves 39 and 40 relating to the arm cylinder 27 and the bucket cylinder 28.
The functional relationship is that the values of the control forces Fc5 and Fc6 gradually decrease as the differential pressure ΔPLS increases along the characteristics of the basic function as a whole, and the differential pressure ΔPLS becomes equal to or less than the minimum flow compensation differential pressure B.
Similar to the functional relationship shown in FIG. 4A, the relationship is limited to the maximum value fmax which is equal to or less than the force f of the springs 49 and 50 regardless of the decrease in the differential pressure ΔPLS.

FIG. 5 shows these functions collectively in order to make the mutual relations of the functions easier to understand.

FIG. 6 shows a functional relationship between the oil temperature Th stored in the function block 86 and the correction coefficient K. When the oil temperature Th is higher than the predetermined temperature ThO, the correction coefficient is 1;
The correction coefficient K gradually becomes smaller than 1 as the oil temperature Th becomes lower than the predetermined temperature ThO.
Here, the predetermined temperature ThO is such that the pressure oil flowing through the circuit is the main pump 2
This temperature is considered to have a viscosity that does not significantly affect the flow rate discharged from 2.

In the delay element blocks 90 to 95, the actuator
Time constants T1 to T6 that give an optimal time delay to their operation are set for each of 23 to 28. Of these, 2
Blocks 91 and 92 corresponding to shunt compensation valves 36 and 37 related to 4, 25
Time constants T2 and T3 are extremely large compared to the other time constants T1 and T4 to T6, and the control force Fc to be applied to the shunt compensating valves 36 and 37.
2. A large time delay is given to a change in the value of Fc3.

FIGS. 7 and 8 show the configuration of a working member of a hydraulic shovel driven by the hydraulic drive device of the present embodiment. The swing motor 23 drives the swing body 100, the left travel motor 24, the right travel motor 25 drives the bandage or travel bodies 101, 102, and the boom cylinder 26, the arm cylinder 27, and the bucket cylinder 28 respectively have the boom 103, the arm 104, The bucket 105 is driven.

Next, the operation of the present embodiment configured as described above will be described.

When any one or more of the flow control valves 29-34 are operated, pressure oil from the main pump 22 is supplied to the corresponding actuator through the diversion compensation valve and the flow control valve. At this time, the main pump 22 is load-sensing controlled by the discharge amount control device 41, the differential pressure ΔPLS between the discharge pressure of the main pump 22 and the maximum load pressure is detected by the differential pressure detector 59, and the corresponding electric signal X1 is Entered in 21. At the same time, the oil temperature is detected by the oil temperature detector 60, and the corresponding electric signal X2 is input to the controller 61.

In the calculation unit 72 of the controller 61, the values of the control forces Fc1 to Fc6 are calculated as described above, and electric signals a to f corresponding to these control forces are given to the electromagnetic proportional pressure reducing valves 62a to 62f. The valves 62a to 62f are driven, and the control forces Fc1 to Fc6
Are guided to the drive units 35c to 40c of the flow compensating valves 35 to 40, respectively. Therefore, the control parts Fc1 to Fc6 in the valve closing direction are applied to the branch flow compensating valves 35 to 40 by the driving units 35c to 40c. As a result, the second control force f is applied to the branch flow compensating valves 35 to 40.
−Fc1, f−Fc2, f−Fc3, f−Fc4, f−Fc5, f−Fc6 are provided in the valve opening direction. That is, at least one of the flow control valves 29 to 34
When one of them is operated, the control forces Fc1 to Fc6 are constantly applied to all the branch flow compensating valves 35 to 40. At this time, since the first control force based on the pressure difference before and after the flow control valve is not acting on the diversion compensating valve in which the flow control valve is not operated, it is maintained at the fully open position.

Next, assuming that the oil temperature is equal to or higher than Tho shown in FIG. 6, when the swing body 100, the traveling bodies 101 and 102, the boom 103, the arm 104, and the bucket 105 are operated alone, and a combined operation thereof is performed. In each case, a shunt compensation valve 35
The operations of to 40 and the accompanying operations of the actuators 23 to 28 will be described. Operate one of the flow control valves 29 to 34 to
100, running bodies 101, 102, boom 103, arm 104, bucket
When a single operation of the flow control valve 105 is performed, the flow-dividing compensating valve related to the corresponding flow control valve has the first
Is applied in the valve closing direction. The differential pressure across the flow control valve cannot exceed the differential pressure ΔPLS between the discharge pressure of the main pump 22 and the maximum load pressure under load sensing control. In the case of single operation, the differential pressure ΔPLS is generally equal to the load sensing compensation. It is maintained at the pressure difference ΔPLSO or a value close thereto.

At this time, if the operated flow control valve is related to one of the swing motor 23, the arm 27, and the bucket 28, the flow dividing control valve 35,
Control force Fc applied to the drive unit 35c, 39c or 40c of 39 or 40
1, Fc5 or Fc6 is obtained from the functional relationship shown in FIG. 4A or FIG. 4D, and the control force corresponding to the load sensing compensation differential pressure ΔPLSO is fo. Therefore, for example, f-fo is applied to the branch flow compensating valve 35 as the second control force. f-
As described above, fo is a value for controlling the differential pressure ΔPv1 across the turning direction switching valve 23 to substantially match the load sensing compensation differential pressure ΔPLSO. Therefore, the second control force f−
fo is always substantially equal to or greater than the first control force, so that the shunt compensating valve 35 is also kept at the fully open position.

When the operated flow control valve is related to one of the traveling motors 24, 25 and the boom cylinder 26, the diversion compensating valve 36, 37 or 38
The control force Fc2, Fc3, or Fc4 applied to the driving units 36c, 37c, or 38c is obtained from the functional relationship shown in FIG. 4B or 4C, and the control force corresponding to the load sensing compensation differential pressure ΔPLSO is It is smaller than fo. Therefore, for example, a force greater than f-fo is applied to the branch flow compensating valve 38 as the second control force. Therefore, also in this case, the second control force is larger than the first control force, and the shunt compensation valve 38 is held at the fully open position.

As described above, in a single operation of operating one of the flow control valves 29 to 34, the corresponding branch flow compensating valve does not basically operate, and the pressure difference before and after the flow control valve is mainly loaded by the main pump 22. The flow is controlled by the sensing control, and the flow rate according to the opening degree of the flow control valve is supplied to the actuator.

Next, by operating any two or more of the flow control valves 29 to 34, the swing body 100, the traveling bodies 101 and 102, the boom 103, the arm 10
4. A case where a combined operation of the actuator of the bucket 105 is performed will be described.

When simultaneously operating the flow control valves 29 and 32 to perform a combined operation of the revolving unit 100 and the boom 103, for example, a combined operation of swivel and boom raising, the pressure oil from the main pump 22 is supplied to the diversion compensation valves 35 and 3
The rotation motor 23 and the boom cylinder 26 are supplied to the swing motor 23 and the boom cylinder 26 through the control valve 8 and the flow control valves 29 and 32. At this time, the differential pressure ΔPLS is usually equal to or less than the maximum flow compensation differential pressure A for the swing motor 23, and the control force Fc1 applied to the drive unit 35c of the branch flow compensation valve 35 is a basic function based on the functional relationship shown in FIG. Is calculated as the control force Fc4 applied to the drive unit 38c of the shunt compensating valve 38 from the functional relationship shown in FIG. 4C.
A value smaller than c1 is calculated. Therefore, the shunt compensation valve
The second control force f-Fc1, f- in the valve opening direction applied to 35,38
Fc4 has a relationship of f−Fc1 <f−Fc4. That is, the control force f-Fc4 in the valve opening direction of the flow dividing compensation valve 38 becomes larger than the control force f-Fc1 in the valve opening direction of the flow dividing compensation valve 35. as a result,
At the start of the combined operation of turning and boom raising, the degree to which the shunt compensation valve 38 related to the boom cylinder 3 on the low load pressure side is throttled by the control force f-Fc4 becomes small, and the shunt compensation valve 38 becomes the shunt compensation valve 35. It tends to open compared to the case where the same control force f-Fc1 is applied. For this reason, the differential pressure across the flow control valve 32 is controlled to be greater than the differential pressure across the flow control valve 29, and the boom cylinder 26 controls the discharge amount of the main pump 22 to the opening ratio of the flow control valves 29 and 32. Is supplied to the swing motor 23, while a smaller flow rate is supplied to the swing motor 23.As a result, the combined operation of swing and boom raising can be reliably performed, and the boom raising speed is increased. , A complex operation in which the turning is relatively gentle is performed.

And, as described above, the swing motor 23 and the boom cylinder 26
When the flow control valve 32 is returned to the neutral position in order to stop the boom cylinder from a state in which
When the pressure oil discharged from the main pump 22 is throttled by the flow control valve 32, the pump pressure temporarily increases and the differential pressure ΔPLS
Is greater than the maximum flow compensation differential pressure A, which is the limit differential pressure during normal combined operation. For this reason, as shown in FIG. 4A, the calculation unit 72 of the controller 61 obtains the value of the constant control force Fc4 regardless of the increase in the differential pressure ΔPLS, that is, the maximum flow compensation control force fc. Therefore, the second control force in the valve opening direction applied to the branch flow compensating valve 35 relating to the swing motor 23 is f-Fc
, And the branching compensating valve 35 is regulated so as not to open too much when it tries to open proportionally as the differential pressure ΔPLS increases.

As a result of this control, the flow control valve is used to stop the boom cylinder 26 when turning and boom raising are combined.
Even if 26 is operated in the neutral direction, the shunt compensating valve 35
Is the maximum flow compensation control force fc corresponding to the maximum flow compensation differential pressure A
Is controlled so that it does not open too much, so that a relatively small flow rate is supplied to the swing motor 23 as compared with the flow rate previously supplied to the swing motor 23, and therefore, the swing motor not intended by the operator. 23 speed increase can be prevented,
Excellent operability and safety are obtained.

When operating the flow control valves 30 and 31 with the same stroke to perform straight traveling, the pressure oil from the main pump 22 is supplied to the branching compensation valve.
Left and right traveling motors 24,2 through 36,37 and flow control valves 30,31
Supplied to 5. At this time, the drive unit 36 of the branch flow compensating valves 36 and 37
As the control forces Fc2 and Fc3 applied to c and 37c, values smaller than the control forces obtained by the characteristics of the basic function are both calculated from the functional relationships shown in FIG. 4B. Therefore, the shunt compensating valves 36 and 37
The second control force f-Fc2, f-Fc3 in the valve opening direction applied to
Is f−F, where Fcr is the control force obtained from the basic function.
c2> f-Fcr and f-Fc3> f-Fcr. Here, the second control force f-Fcr based on the basic function is a value set so that the target value of the differential pressure across the flow control valve is equal to the differential pressure ΔPLS. Therefore, the flow compensating valves 36 and 37 are
Compared to the normal case in which the pressure difference between the front and rear of the valve 31 is controlled to be substantially equal to the pressure difference ΔPLS, the pressure difference between the flow control valves 30 and 31 is increased by the larger second control force in the valve opening direction. The throttle is not stopped until the pressure increases by a predetermined value ΔPo corresponding to Fc2-Fcr or Fc3-Fcr further than the differential pressure ΔPLS. Therefore, when a pressure difference occurs between the load pressures of the traveling motors 24 and 25, none of the pressure control valves is throttled in a range where the pressure difference is smaller than a predetermined value ΔPo, and the traveling motors 24 and 25 are connected in parallel. It will be in the same state as above. Even when the differential pressure exceeds the predetermined value ΔPo, the shunt compensating valve on the low load pressure side is opened larger than usual, so it is considered that the traveling motors 24 and 25 are partially connected in parallel. be able to.

As a result of the diversion compensating valve functioning as described above,
Even if the resistance applied to the left and right covers is different and the load pressure of the traveling motors 24 and 25 is different, the traveling motors 24 and 25 are at least partially in the same state as being connected in parallel. As in the case of a general hydraulic circuit in which traveling motors are connected in parallel, the flow rate of the pressure oil supplied to the left and right traveling motors 24 and 25 is forcibly made equal by the straight traveling maintenance force of the band itself, and the straight traveling is performed. Travel can be continued. Therefore, the labor for manual adjustment by the operator can be reduced, and the fatigue of the operator can be reduced.

In addition, since the functions of the branch flow compensating valves 24 and 25 are partially disabled and the straight traveling is forcibly performed by the straight traveling maintaining force of the band itself, the flow control valves 30 and 31 and the branch flow compensating valves 36 and 37
Even if there are variations in the performance of hydraulic equipment such as due to manufacturing errors, it is possible to perform intended straight running, and furthermore, it is possible to continue straight running even if there is a slight change in the operation lever position. Similarly, the labor for manual adjustment by the operator can be reduced, and the fatigue of the operator can be reduced.

Next, while operating the flow control valves 30 and 31 and driving the traveling motors 24 and 25 to perform the traveling operation, the flow control valve 32 is further operated to shift to the combined operation of traveling and boom raising. Consider the case.

From the state where only the traveling operation is being performed, the flow control valve is further increased.
By operating 32, the pressure oil from the main pump 22 has been supplied only to the left and right traveling motors 24 and 25, but is supplied to the boom cylinder 26 through the branch flow compensation valve 38 and the flow control valve 32. Become like

By the way, in the case of the combined operation of running and boom raising, the boom cylinder 26 is usually on the high load pressure side. Therefore, at the moment when the operation is shifted from the state in which only the traveling operation is performed to the combined operation of traveling and boom raising, a situation in which the differential pressure ΔPLS is extremely reduced occurs, and the arithmetic unit 72 of the controller 61 causes the functional relationship shown in FIG. Control force Fc2, Fc required from
The value of 3 increases instantaneously and greatly. For this reason, when the control forces Fc2 and Fc3 are output directly from the output unit 73 as the electric signals b and c, the second control forces f-Fc2 and f-Fc3 in the valve opening direction are used.
Decreases sharply correspondingly. In other words, in the initial stage of shifting from the operation of traveling only to the combined operation of traveling and boom raising,
The shunt compensating valves 36 and 37 are momentarily closed extremely and then start to open again.
Fluctuations in the flow rate of the pressure oil supplied to the tanks 4 and 25 increase, and accordingly, the running speed fluctuates extremely, causing a large shock to the body of the hydraulic shovel and reducing operability.

On the other hand, in the present embodiment, as described above, the delay element blocks 90 to 95 shown in FIG. 3 are provided. Of these, the time constants T2 and T3 of the blocks 91 and 92 related to the traveling motor 24 and 25 are different. The time constants T1, T4 to T6 are extremely large, and a large time delay is given to a change in the values of the control forces Fc2, Fc3. Therefore, as described above, even if the values of the control forces Fc2 and Fc3 suddenly change,
In 92, the change is moderated, and the changes in the control forces Fc2, Fc3 applied from the driving units 36c, 37c also become gentle. Therefore, it is possible to prevent the shunt compensating valves 36 and 37 from closing suddenly, to reduce the above-mentioned fluctuation in the traveling speed, and to obtain excellent operability without causing a large shock to the body of the hydraulic shovel.

Further, while operating at least one of the flow control valves 29, 33, and 34 to drive the corresponding one of the swing motor 23, the arm cylinder 27, and the bucket cylinder 28, another load pressure higher than the other is used. When the differential pressure ΔPLS becomes momentarily zero for any reason, such as when the actuator is further driven, the functional relationship between the differential pressure and the control force related to the swing motor 23, the arm cylinder 27, and the bucket cylinder 28 is as follows:
4A and 4D, since the slope is the same as that of the basic function, when the functional relationship is completely matched to the basic function, the values of the control forces Fc1, Fc5, and Fc6 are changed by the springs 45, 49. , 50, and the flow compensating valves 35, 39, 40 are completely closed. When the flow compensating valve is completely closed, the flow rate of the pressure oil supplied to the actuators 23, 27, and 28 becomes zero, and a large shock occurs in the revolving unit 100, the arm 104, and the bucket 105, and the operability is significantly deteriorated. Not only
The hydraulic equipment may be damaged.

In the present embodiment, when the differential pressure ΔPLS becomes equal to or less than the minimum flow compensation differential pressure B with respect to such a decrease in the differential pressure ΔPLS, the springs 4 are controlled regardless of the decrease in the control forces Fc1, Fc5, and Fc6.
The relationship is limited to a maximum value fmax of 5 or less. Therefore, the flow compensating valves 35, 39, and 40 are prevented from being completely closed, so that shocks can be reduced, operability can be improved, and damage to the hydraulic equipment can be prevented.

Next, when the oil temperature changes below Tho shown in FIG. 6, the operation of the branch flow compensating valves 35 to 40 and the operation of the actuators 23 to 28 associated therewith will be described.

In the calculation unit 72 of the controller 61, the third
As shown in the figure, the control force Fc obtained in the function blocks 83 to 85
The multiplication blocks 87 to 89 multiply the values of 4 to Fc6 by the correction coefficient K of the oil temperature Th obtained in the function block 86 to correct the control forces Fc4 to Fc6 by temperature. As shown in FIG. 6, the correction coefficient K is substantially 1 when the oil temperature Th is higher than the predetermined temperature Tho, and gradually becomes smaller as 1 becomes lower when the oil temperature Th is lower than the predetermined temperature Tho. . From this, if the oil temperature Th is equal to or higher than the predetermined temperature Tho in a normal working environment such as during the daytime, K = 1.
The values of the control forces Fc4 to Fc6 obtained by the function blocks 83 to 85 are directly converted into electric signals b, e, and f, and the branch flow compensating valves 38 to 40 are driven according to the control forces Fc4 to Fc6. Thereby, for example, the flow control valves 38 and 39 are operated, and the boom 103 and the arm 10 are operated.
In the case of performing the combined operation of 4, the boom cylinder 26 and the arm cylinder 27 are not affected at all, that is, since the oil temperature Th is relatively high, the viscosity of the oil temperature is small, and there is no large flow resistance. Compensation valves 38, 39 and flow control valves 32, 33
The pressure oil from the main pump 22 is supplied to the boom cylinder 26 and the arm cylinder 27 via, and the combined drive of the arm and the bucket can be performed without lowering the operation speed of these actuators.

If the oil temperature T is lower than the predetermined temperature Tho in a work environment in a cold region, or in a work environment such as early in the morning or at night in winter, K <1 and the multiplication blocks 87 to 89
The values of the control forces Fc4 to Fc6 multiplied by the correction coefficient K become smaller than the values calculated in the function blocks 83 to 85,
Moreover, the degree increases as the oil temperature Th decreases. As a result, the control forces Fc4 to Fc6, which are smaller than those at the normal time, are reduced in accordance with the decrease in the oil temperature Th, and the drive units 38c of the branching compensation valves 38 to 40 are controlled.
-40c, and the second control forces f-Fc4, f-Fc5, f-Fc6 in the valve-opening direction applied to the branch flow compensating valves 38 to 40 are larger than usual in accordance with the decrease in the oil temperature Th. Become. That is, for example, by operating the flow control valves 38 and 39, the boom 103 and the arm 10
In the case of performing the combined operation of 4, the flow rate substantially equal to the flow rate when the oil temperature Th is high is equal to the flow dividing valves 38 and 39 and the flow control valves 32 and 33.
Is supplied to the boom cylinder 26 and the arm cylinder 27, whereby the viscosity of the pressurized oil increases due to the decrease in the oil temperature Th and the flow resistance increases, but the flow control valve is provided to the boom cylinder 26 and the arm cylinder 27. The desired flow rates required by the devices 32 and 33 can be supplied, and the combined operation can be performed without lowering the operation speed of these actuators.

The same applies to the case of performing a combined operation of another combination of the boom 103, the arm 104, and the bucket 105, or a single operation thereof.

As described above, for the branch flow compensation valves 38 to 40 corresponding to the boom cylinder 26, the arm cylinder 27, and the bucket cylinder 28, the pressure compensation is performed by correcting the values of the control forces Fc4 to Fc6 according to the change in the oil temperature Th. By adjusting the characteristics, the operation speeds of these actuators can always be kept constant irrespective of changes in the oil temperature, and a stable single operation or a composite operation can be performed.

On the other hand, the control forces Fc1 to Fc3 obtained in the function blocks 80 to 82 corresponding to the turning motor 23 and the traveling motors 24 and 25 are not subjected to the oil temperature correction, and the delay element blocks 90 to 92 are not changed.
Are output as electric signals a to c. Therefore, when the oil temperature is equal to or lower than the predetermined temperature Tho, the viscosity of the pressure oil increases, the flow resistance increases, and the flow rate supplied to the boom cylinder 26 and the arm cylinder 27 decreases. Therefore,
The swing motor 23 and the traveling motors 24 and 25 which are motor-based actuators are boom cylinders 26, arm cylinders 27 and bucket cylinders which are cylinder-based actuators.
Unlike 28, pressure oil is driven by passing through the inside,
If high-viscosity pressure oil is supplied at the same flow rate as normal low-viscosity oil, internal components may be damaged, but the flow rate will decrease, and such damage will not occur. .

As described above, according to the present embodiment, in the calculation unit 72 of the controller 61, the function blocks 80 to 85 provided corresponding to the actuators 23 to 28 are used to control the branch flow compensation valves 35 to 40 based on the differential pressure ΔPLS. The values of the control forces Fc1 to Fc6 to be applied via the driving units 35c to 40c are individually calculated, and the shunt compensation valve
Since the control pressures Pc1 to Pc6 corresponding to these control forces are individually generated from the electromagnetic proportional pressure reducing valves 62a to 62f provided corresponding to 35 to 40, and the control pressures Pc1 to Pc6 are guided to the driving units 35c to 40c.
The shunt compensation valves 35 to 40 can be provided with individual pressure compensation characteristics suitable for the associated actuators 23 to 28, so that in the combined operation of the driven bodies 100 to 105, the optimum shunt flow according to the type of the driven body is provided. Ratio can be obtained, and operability and work efficiency can be improved.

Also, the control forces Fc1 to Fc correspond to the actuators 23 to 28.
Since the values of c6 are individually calculated and the corresponding control pressures Pc1 to Pc6 are individually generated from the electromagnetic proportional pressure reducing valves 62a to 62f, the values of the control forces Fc1 to Fc6 can be individually corrected. Therefore, the optimum time constants T1 to T6 are individually given to each actuator in the element blocks 90 to 95, or a function block 86 for oil temperature correction is provided, and only the control forces Fc4 to Fc6 are corrected by the correction coefficient K. It is also possible to make the operating characteristics of the shunt compensating valve more different in consideration of various conditions, such as the operation of the actuators 23 to 28, thereby further improving the operability and work efficiency in the combined operation of the actuators 23 to 28. be able to.

In the above embodiment, various modifications can be made to the function form of the differential pressure ΔPLS and the control forces Fc1 to Fc6 stored in the function blocks 80 to 85.

For example, in the function block 80 relating to the swing motor 23, as shown in FIG. 4A, when the differential pressure ΔPLS temporarily increases and becomes larger than the maximum flow compensation differential pressure A,
Although the functional relationship is determined so as to obtain a constant control force, that is, the maximum flow compensation control force fc, the functional relationship may be determined in other forms. For example, as shown in FIG. 9, in consideration of the flow characteristics of the pressurized oil, the temperature of the pressurized oil, etc., as the pressure difference ΔPLS becomes larger than the maximum flow rate compensation differential pressure A, the maximum flow rate compensation control force
As shown in FIG. 10, a functional relationship that outputs a proportionally larger control force starting from fc or a control force that increases stepwise as the differential pressure ΔPLS becomes larger than the maximum flow compensation differential pressure A, as shown in FIG. It can be set to a functional relationship to be output or a functional relationship as shown in FIG. 11, which increases in a curve as the differential pressure ΔPLS becomes larger than the maximum flow rate compensating differential pressure A. Further, as shown in FIG. In this manner, a functional relationship can be set in which a control force is output that decreases proportionally with a relatively small gradient as the differential pressure ΔPLS becomes larger than the maximum flow compensation differential pressure A.

Further, in the above embodiment, the functional relationship is established so that a constant control force fc can be obtained only when the differential pressure ΔPLS becomes larger than the maximum flow rate compensating differential pressure A only for the diversion compensating valve 35 related to the swing motor 23. Is set, but the functional relationship between the differential pressure ΔPLS and the control force can be appropriately set similarly for the branching compensating valves related to the other actuators.

Further, in the function blocks 81 and 82 relating to the traveling motors 24 and 25, as shown in FIG. 4B, as shown in FIG. 4B, as the differential pressure ΔPLS increases, the functional relationship such that the difference in control force with respect to the characteristics of the basic function decreases. As shown in FIG. 13, as shown in FIG. 13, a functional relationship in which the difference in the control force with respect to the characteristics of the basic function is constant regardless of the change in the differential pressure ΔPLS, or as the differential pressure ΔPLS increases, The same effect can be obtained even when a functional relationship in which the difference in control force with respect to the characteristics becomes large is obtained.

Second Embodiment A second embodiment of the present invention will be described with reference to FIGS. In the figure, the same reference numerals are given to members equivalent to those shown in FIGS. 1 to 12.

In FIG. 15, the turning direction switching valve 29 and the boom direction switching valve 32 detect these operations and detect the electric signals X3 and X3.
Operation detectors 110 and 111 that output X4 are provided. In addition, the shunt compensating valves 35A to 40A have springs 45 to 40 of the first embodiment.
Instead of 50, the same reference pilot pressure Pr is led via the pilot lines 112a to 112f, respectively,
Drive units 45A to 50A are provided to urge the valve bodies A to 40A in the valve opening direction with the same force as that of the springs 45 to 50.

The electric signals X3, X4 output from the operation detectors 110, 111 are
Electric signals output from the differential pressure detector 59 and the temperature detector 60
Input to the controller 61A together with X1 and X2,
In 61A, the control signals Fc1 to Fc to be applied by the drive units 35c to 40c of the shunt compensation valves 35A to 40A using the electric signals X1, X2, X3, and X4.
Calculate the value of c6 and output the corresponding electrical signals a, b, c, d, e, f.

The control pressure generation circuit 65A also serves as a reference pilot pressure generation circuit. Therefore, based on the pilot pressure output from the pilot pump 63, the control pressure generation circuit 65A absorbs fluctuations in the pilot pressure and generates a stable and constant reference pilot pressure Pr. A pressure reducing valve 113 is further provided.
r is the pilot line 112a via the pilot line 112
~ 112f.

Electromagnetic proportional pressure reducing valves 62a to 62f, relief valve 64 and pressure reducing valve 11
3 is preferably configured as an aggregate in one block, as indicated by a two-dot chain line 66A.

The controller 61A includes an input unit, a storage unit, a calculation unit, and an output unit as in the first embodiment.

FIG. 16 is a functional block diagram showing the contents of the calculation performed by the calculation unit of the controller 61A. In the present embodiment, a second function block 83A is provided in addition to the function block 83 as a function block corresponding to the branch flow compensating valve 38, and the differential pressure ΔΔ based on the electric signal X1 at that time from these function blocks 83 and 83A is provided.
The control force values Fc4 and Fc4o corresponding to the PLS are obtained, and one of them is selected by the switch function of the selection block 114. Also, the electric signals X3, X4 from the operation detectors 110, 111
Is input to an AND block 115, and when both signals are ON, an ON signal is output from the AND block 115 to the selection block 114. Selection block 114 is ON from AND block 115
When there is no signal, the control force Fc4o is selected, and when an ON signal is given, the control force Fc4 is selected.

The relationship between the differential pressure PLS and the control force Fc4 stored in the function block 83 is as described in the first embodiment. Function block
The relationship between the differential pressure PLS stored in 83A and the control force Fc4o is as follows.
This embodiment has been described with reference to FIG. 4D. Dividing flow compensation valves 39, 40 related to arm cylinder 27 and bucket cylinder 28
Are the same as the function relationships stored in the function blocks 84 and 85 corresponding to. That is, as a whole, the value of the control force Fc4o gradually decreases in accordance with the increase of the differential pressure ΔPLS along the characteristics of the basic function, and when the differential pressure ΔPLS becomes equal to or less than the minimum flow compensation differential pressure B, the differential pressure ΔPLS decreases. Irrespective of this, the relationship is limited to the maximum value fmax which is equal to or less than the urging force f of the driving unit 48A.

In the second embodiment configured as described above, the boom
In the combined operation of 103 and a driven member other than the revolving superstructure 100, the turning direction switching valve 29 is not operated, so that the operation detector 110 does not output the electric signal X3, and the controller 61A
In, the AND block 115 does not output an ON signal, and the selection block 114 selects the control force Fc4o obtained by the function block 83A as the control force. For this reason, the drive unit 3 of the branch flow compensating valve 38A
In 8c, the control force Fc4o based on the basic function is applied, and the second control force f-Fc4o in the valve opening direction is a value at which the target value of the differential pressure ΔPv4 across the flow control valve 32 substantially matches the differential pressure ΔPLS. .
That is, the second control force f-Fc4o has a normal value smaller than the second control force f-F4 by the control force Fc4 of the function block 83. Thus, when the boom cylinder 26 is on the low load pressure side, the throttle amount of the branch flow compensating valve 38A does not become too small, and the differential pressure across the flow control valve 32 is substantially equal to the differential pressure ΔP
It is possible to control the boom cylinder 26 so that the pressure oil is controlled to match the LS and an appropriate flow rate of the hydraulic oil according to the operation amount of the flow control valve 32 is provided.

In the combined operation of the revolving unit 100 and the boom 103, both the flow control valves 29, 32 are operated, so that the operation detectors 110, 1
11 output electric signals X3 and X4,
In 1A, the AND block 115 outputs an ON signal, and the selection block 114 selects the control force Fc4 obtained by the function block 83 as the control force. For this reason, as in the case of the combined operation of turning and boom raising described in the first embodiment, the second control force f-Fc1, f in the valve opening direction applied to the branch flow compensating valves 35,38.
−Fc4 has a relationship of f−Fc1 <f−Fc4, and the boom cylinder 26 is supplied with a flow rate larger than the flow rate in which the discharge amount of the main pump 22 is distributed by the opening ratio of the flow control valves 29 and 32. A combined operation of turning and boom raising, in which the raising speed is fast and the turning is relatively gentle, is performed.

Further, in the present embodiment, one drive means relating to the second control force of the branch flow compensating valves 35A to 40A is replaced with a spring, and the same reference pilot pressure Pr is guided through the pilot pipelines 112 and 112a to 112f. The drive units are 45A to 50A. Therefore, variations due to spring manufacturing errors and aging are small, and drive errors between the shunt compensating valves 35A to 40A can be extremely reduced. As a result, the individual second control forces f-Fc1, f-Fc2, f to be applied to the respective flow dividing valves 35A to 40A, respectively.
−Fc3, f−Fc4, f−Fc5, f−Fc6 can be realized more accurately than when a spring is used, and the intended composite operation can be performed accurately.

Further, in the present embodiment, the reference pilot pressure Pr guided to the driving units 45A to 50A is output from the pressure reducing valve 113, and the pressure reducing valve 113 is set by the relief valve 64, which is the same as the electromagnetic proportional pressure reducing valves 62a to 62f. It is configured to use the pilot pressure.

By the way, in the relief valve 64 having the configuration as illustrated, when the tank pressure changes due to the return oil from the actuator or the like, the pilot pressure, which is the output of the relief valve 64, also changes according to the change. When the pilot pressure changes, the outputs of the electromagnetic proportional pressure reducing valves 62a to 62f, that is, the control pressures Pc1 to Pc6 change even if the electric signals a to f are constant. Therefore, assuming that the force f applied by the drive units 45A to 50A is constant, the second control force in the valve opening direction fluctuates even though the electric signals a to f are constant.

On the other hand, in the present embodiment, the output of the pressure reducing valve 113, that is, the reference pilot pressure Pr also changes with a change in the pilot pressure. That is, when the control pressures Pc1 to Pc6 change, the reference pilot pressure Pr also changes correspondingly. For this reason,
Both changes are canceled, and as a result, the second control force in the valve opening direction becomes constant. Therefore, in the present embodiment, the influence of the change in the tank pressure due to the return oil from the actuator is not given to the drive of the shunt compensation valves 35A to 40A, and the respective shunt compensation valves 35A to 40A regardless of the change in the tank pressure. Second control force f-Fc1, f-Fc2, f-Fc3, f
−Fc4, f−Fc5, f−Fc6 can be realized more accurately, and excellent control accuracy can be obtained.

Third Embodiment A third embodiment of the present invention will be described with reference to FIGS. In the drawings, members that are the same as the members shown in FIGS. 1 to 12 are given the same reference numerals.

In FIG. 17, the flow dividing compensating valves 35B to 40B serve as driving means related to the second control force in the valve opening direction, as the driving means of the springs 45 to 50 and the driving units 35c to 40c of the first embodiment. Instead, a single drive element for biasing the valve elements of the branch flow compensating valves 35B to 40B in the valve opening direction, that is, drive units 35d to 40d is provided, and the drive units 35d to 40d are provided through pilot lines 51a to 51f. The control pressures Pc1 to Pc6 are derived, and the second control forces f-Fc1, f-Fc
2, f-Fc3, f-Fc4, f-Fc5, and f-Fc6 act directly. Hereinafter, this second control force is referred to as Hc1 to Hc, respectively.
Expressed as c6.

In this embodiment, the actuators 23 to 28
And a selection device 120 including six selection switch elements 120a to 120f, each of which can be selectively operated by an operator at one of a plurality of positions, and the selection switch elements 120a to 120f are respectively provided. A selection command signal having contents corresponding to the selected position is output as electric signals Y1 to Y6.

The controller 61B includes an input unit, a storage unit, a calculation unit, and an output unit as in the first embodiment. An electric signal X1 output from the differential pressure detector 59 and an electric signal Y1 output from the selection device 120 are input to an input portion of the controller 61B.
YY6 is input, and in the calculation unit of the controller 61B, calculation is performed to obtain the values of the control forces Hc1 to Fc6 from the electric signals X1 and Y1 to Y6 according to the function data and the control program stored in the storage unit, and the output unit The value of the control force is the electric signal a
Ｆf.

FIG. 18 is a functional block diagram showing the contents of the calculation performed by the calculation unit of the controller 61B. In the figure, blocks 80B to 85B
Are provided corresponding to the branch flow compensating valves 35B to 40B, and the differential pressure ΔP
It is a function block in which function data including a plurality of functional relationships between the LS and the control forces Hc1 to Hc6 are stored in advance. Function block 80
In B to 85B, one functional relationship according to the content of the selection command signal is selected based on the electric signals Y1 to Y6,
Further, the electric signal at that time is obtained from these selected functional relationships.
Control force values Hc1 to Hc6 corresponding to differential pressure ΔPLS based on X1 are calculated. Thus, the function block 80B
The values Hc1 to Hc6 of the control force obtained in 〜85B are filtered as first-order lag elements in delay blocks 90 to 95, respectively, and then output as electric signals a to f.

Differential pressure ΔPLS and control force Fc1 to stored in function block 80B
FIG. 19 shows a plurality of functional relationships of Fc6. In the figure, the solid line So corresponds to the characteristic of the basic function described in the first embodiment, and the control force increases as the differential pressure ΔPLS between the discharge pressure of the main pump 22 and the maximum load pressure of the actuators 23 to 28 increases. Hc
It is a functional relationship that gradually increases 1. This functional relationship So is used for normal driving of the swing motor 23 including independent operation of the swing body 100 without the need to correct the second control force in the valve opening direction of the branch flow compensation valve 35B.

The dashed lines So + 1 and So + 2 indicate a functional relationship in which the control force Hc1 is gradually increased with a gradient larger than the function So as the differential pressure ΔPLS increases, and the dashed lines So−1 and So−2 indicate the differential pressure ΔP
This shows a function that gradually increases the control force Hc1 with a smaller gradient than the function So as LS increases.

That is, the dashed lines So + 1 and So + 2 have a greater gradient than the characteristic line So of the basic function, make the second control force Hc1 in the valve opening direction of the shunt compensating valve 35B larger than in the case of using the basic function, The functional relationship is such that the differential pressure is larger than the differential pressure ΔPLS between the main pump 22 and the maximum load pressure of the actuators 23 to 28. This functional relationship is used when it is desired to increase the flow rate supplied to the slewing motor 23 in a combined operation in which the slewing motor 23 is on the low load pressure side, as compared with a normal case.

The dashed lines So + 1 and So−2 are functions that make the second control force in the valve opening direction of the valve flow compensating valve 35B smaller than that obtained by the basic function, and make the differential pressure across the flow control valve 29 smaller than the differential pressure ΔPLS. This relationship is used when it is desired to reduce the flow rate supplied to the swing motor 23 in a combined operation in which the swing motor 23 is on the low load pressure side, as compared with a normal case.

Note that ΔPLSO is the differential pressure between the discharge pressure of the main pump 22 and the maximum load pressure held by the discharge amount control device 41 of the load sensing control system, that is, the spring of the control valve 53, as in the first embodiment. This is the load sensing compensation differential pressure set at 54.

In other function blocks 81B to 85B, the function block
A plurality of functional relationships are stored substantially as in 80B. The number and type of the plurality of function relationships stored in each of the function blocks 80B to 85B are determined so that the optimal operation characteristics are given to the related actuators 23 to 28 according to the type and content of the operation at the time of the composite operation. .

Electric signals a to f output from the controller 61B are input to a plurality of electromagnetic proportional pressure reducing valves 62a to 62f as in the first embodiment. The electromagnetic proportional pressure reducing valves 62a to 62f output the electric signals a to
f and the corresponding control pressure Pc
1 to Pc6 are output. These control pressures Pc1 to Pc6 are guided to the drive units 35d to 40d of the shunt compensation valves 35B to 40B, respectively, whereby the shunt compensation valves 35B to 40B are given control forces Hc1 to Hc6 calculated by the controller 61B, The branch flow compensating valves respond accordingly to the differential pressures ΔPv1 to ΔP
Control v6.

Next, the operation of the present embodiment configured as described above will be described.

For example, when performing a combined operation of turning and boom raising for the purpose of loading earth and sand, the operator operates the corresponding selection switch elements 120a and 120d of the operation device 120 to select a functional relationship suitable for the work content, and The corresponding selection instruction signals, that is, electric signals Y1 and Y4 are output. In the controller 61B, based on the electric signals Y1 and Y4, the shunt compensation valve 35B corresponding to the swing motor 23 is subjected to, for example, a broken line So-2 in FIG. 19 from the plurality of functional relationships stored in the function block 80B. Is selected, and for the shunt compensating valve 38B corresponding to the boom cylinder 26, for example, a functional relationship corresponding to the broken line So + 2 in FIG. 19 is selected from a plurality of functional relationships stored in the function block 83B. I do.

FIG. 20 collectively shows the functional relationships selected in the function blocks 80B and 83B. In the figure, 121 is a characteristic line corresponding to the basic function So, 122 is a characteristic line corresponding to the functional relationship of the broken line So-2 selected in the function block 80B corresponding to the swing motor 23, and 123 is the boom cylinder 26 is a characteristic line corresponding to the functional relationship of the broken line So + 2 selected in the function block 83B corresponding to 26.

Further, in the function blocks 80B and 83B, the control forces H1 and H4 based on the differential pressure ΔPLS are respectively obtained from the selected functional relationships 122 and 123, and the corresponding electric signals a and d are output to the electromagnetic proportional pressure reducing valves 62a and 62b. Is done.

Accordingly, the electromagnetic proportional pressure reducing valve 62d applies a control pressure greater than the control pressure corresponding to the control force Ho based on the differential pressure ΔPLS.
Pc4, while the electromagnetic proportional pressure reducing valve 62a outputs a control pressure Pc1 smaller than the control pressure corresponding to the control force Ho, and these control pressures Pc1 and Pc4 are used to drive the drive units 35d and 35d of the shunt compensation valves 35B and 38B.
Guided to 38d respectively. In this case, since the drive unit 38d of the shunt compensation valve 38B applies a control force H4 greater than the normal control force Ho, the shunt compensation valve 38B is controlled such that the throttle amount is forcibly reduced, and therefore the flow rate is reduced. The control valve 32 is supplied with a larger flow rate than usual, and the diverting compensation valve 35
Since the drive unit 35d of B applies a control force H1 smaller than the normal control force Ho, the shunt compensating valve 35B is controlled so that the throttle amount is forcibly increased. A smaller flow rate than is supplied.

FIG. 21 and FIG. 22 show the flow rate characteristics at this time, and FIG. 21 shows the differential pressure ΔPv across the boom flow control valve 32.
FIG. 22 shows the relationship between 4 and the supply flow rate Q4, and FIG. 22 shows the relationship between the supply / reception flow rate Q1 and the pressure difference ΔPv1 across the flow control valve 29 for turning. Here, a characteristic line 1 for the characteristic line 121 of the basic function
If the ratio of the gradient of 23 is α, the flow control valve for the boom 32
On the other hand, in the case of control by the differential pressure ΔPLS which is a normal time, the flow rate is relatively small as indicated by the characteristic line 124A in FIG.
What was Q4A can be supplied with a larger flow rate Q4B than the flow rate Q4A as shown by the characteristic line 124B in FIG. Assuming that the ratio of the gradient of the characteristic line 122 to the characteristic line 121 of the basic function is β, on the side of the turning flow control valve 29, in the case of control by the differential pressure ΔPLS which is a normal state,
What was a relatively large flow rate Q1A as shown by the characteristic line 125A in the figure, but in this sediment loading work, as shown by the characteristic line 125B in FIG. 22, according to the corrected differential pressure β · ΔPLS,
A flow rate Q1B smaller than the flow rate Q1A can be supplied.

That is, during the sediment loading operation, a relatively large flow rate can be supplied to the boom cylinder 26 and a relatively small flow rate can be supplied to the swing motor 23 as compared with the normal control, so that the boom cylinder 26 and the swing motor 23 can be supplied with this flow rate. The optimum flow rate can be distributed according to the sediment loading work, thereby reducing the flow rate to be relieved on the swing motor 23 side and reducing the throttle amount of the diversion compensating valve 38B on the boom cylinder 26 side to reduce the flow rate. It is possible to suppress the energy of the pressure oil passing through 38B from being converted into heat, thereby reducing the energy loss. Further, since a relatively large flow rate can be supplied to the boom side, the amount of rise of the boom can be sufficiently secured, and excellent workability is provided.

Next, in the case of performing a digging operation aimed at improving the work efficiency as compared with a normal digging operation, that is, a combined operation of the arm bucket for the purpose of a special digging operation, the operator needs a functional relationship suitable for the content of the operation. Is operated by operating the corresponding selection switch elements 120e and 120f of the operation device 120 to output corresponding selection instruction signals, that is, electric signals Y5 and Y6.
In the controller 61B, based on the electric signals Y5 and Y6, for the shunt compensation valve 39B corresponding to the arm cylinder 27, for example, a broken line So-1 in FIG. 19 from the plurality of functional relationships stored in the function block 84B. And a function relationship corresponding to, for example, a broken line So + 1 in FIG. 19 is selected from the plurality of function relationships stored in the function block 85B for the diversion compensating valve 40B corresponding to the bucket cylinder 28. I do.

FIG. 23 collectively shows the functional relationships selected in the function blocks 84B and 85B. In the figure, reference numeral 121 denotes a characteristic line corresponding to the basic function So, 126 denotes a characteristic line corresponding to the functional relationship of the broken line So-1 selected by the function block 84B corresponding to the arm cylinder 27, and 127 denotes a bucket. This is a characteristic line corresponding to the functional relationship of the broken line So + 1 selected in the function block 85B corresponding to the cylinder 28.

Further, in the function blocks 84B and 85B, the control forces H5 and H6 based on the differential pressure ΔPLS are respectively obtained from the selected functional relationships 126 and 127, and the corresponding electric signals e and f are output to the electromagnetic proportional pressure reducing valves 62e and 62f. Is done.

As a result, the electromagnetic proportional pressure reducing valve 62e controls the control pressure smaller than the control pressure corresponding to the control force Ho based on the differential pressure ΔPLS.
Pc5, on the other hand, the electromagnetic proportional pressure reducing valve 62f outputs a control pressure Pc6 greater than the control pressure corresponding to the control force Ho, and these control pressures Pc5, Pc6 are the drive units 39d, 39d, of the shunt compensation valves 39B, 40B.
Each led to 40d. In this case, since the drive unit 39d of the shunt compensation valve 39B applies a control force H5 smaller than the normal control force Ho, the shunt compensation valve 39B is controlled such that the throttle amount is forcibly increased, and thus the flow rate is reduced. The control valve 33 is supplied with a smaller flow rate than normal, and
Since the drive unit 40d of B applies a control force H5 larger than the normal control force Ho, the branching compensation valve 40B is controlled so that the throttle amount is forcibly reduced, and therefore the flow control valve 34 is normally A larger flow rate than is supplied.

With this, in the combined operation of the arm and the bucket,
By making the driving speed of the arm cylinder 27 relatively low and the driving speed of the bucket cylinder 28 relatively high, it is possible to realize a special excavation work which is considered to be better in work efficiency than ordinary excavation.

Next, in the combined operation of the same arm and bucket, when performing the combined operation of the arm and the bucket for the purpose of shaping work to level the ground, for example, the operator selects a functional relationship suitable for the work content. The corresponding selection switch elements 120e and 120f of the operation device 120 are operated to output corresponding selection instruction signals, that is, electric signals Y5 and Y6.
In the controller 61B, based on the electric signals Y5 and Y6, the shunt compensation valve 39B corresponding to the arm cylinder 27 corresponds to, for example, a broken line So + 1 in FIG. 19 from the plurality of functional relationships stored in the function block 84B. For the diversion compensating valve 40B corresponding to the bucket cylinder 28, a functional relationship corresponding to, for example, a broken line So-1 in FIG. 19 is selected from the plurality of functional relationships stored in the function block 85B. I do.

FIG. 24 shows the functional relationships selected in the function blocks 84B and 85B. In the figure, 121 is a characteristic line corresponding to the basic function So, 128 is a characteristic line corresponding to the functional relationship of the broken line So + 1 selected by the function block 84B corresponding to the arm cylinder 27, and 129 is the bucket cylinder 28. Is a characteristic line corresponding to the functional relationship of the broken line So-1 selected in the function block 85B corresponding to.

In the function blocks 84B and 85B, the control force H′5,
H′6 are obtained, and the corresponding electric signals e, f
Is output to the electromagnetic proportional pressure reducing valves 62e and 62f.

Accordingly, the electromagnetic proportional pressure reducing valve 62e applies a control pressure greater than the control pressure corresponding to the control force Ho based on the differential pressure ΔPLS.
Pc5, while the electromagnetic proportional pressure reducing valve 62f outputs a control pressure Pc6 smaller than the control pressure corresponding to the control force Ho, and these control pressures Pc5, Pc6 are the drive units 39d, 39d, of the shunt compensation valves 39B, 40B.
Each led to 40d. In this case, since the drive unit 39d of the shunt compensation valve 39B applies a control force H'5 larger than the normal control force Ho, the shunt compensation valve 39B is controlled such that the throttle amount is forcibly reduced. Accordingly, the flow rate control valve 33 is supplied with a larger flow rate than usual, and the drive unit 40d of the flow diverting compensation valve 40B applies a control force H'6 smaller than the normal control force Ho. 40B is controlled such that its throttle amount is forcibly increased, and therefore, a smaller flow rate than usual is supplied to the flow control valve 34.

With this, in the combined operation of the arm and the bucket,
The driving speed of the arm cylinder 27 is relatively high, while the driving speed of the bucket cylinder 28 is relatively low.

Modification of Third Embodiment A modification of the third embodiment will be described with reference to FIG. In the figure, elements that are the same as the elements shown in FIG. 18 are given the same reference numerals.

In the present embodiment, for example, five selection switch elements 13 provided in correspondence with the operation modes and selectively operated by the operator, instead of the selection device 120 described above.
A selection device 130 including 0a to 130e is provided. The selection switch elements 103a to 130e respectively
The selection command signal corresponding to the corresponding work mode is transmitted to the electrical signal Za.
~ Ze, but only one of them is operated at a time, and one of the electrical signals Za ~ Ze is output from the selection device 130 corresponding to the operated selection switch element. Is output.

The controller 61C includes an input unit, a storage unit, a calculation unit, and an output unit as in the first embodiment. The input section of the controller 61C has an electric signal X1 output from the differential pressure detector 59 and an electric signal Za output from the selection device 130.
~ Ze is input, and in the operation unit of the controller 61C, in the function selection instruction block 131, the function blocks 80B to 85B are selected according to the input electric signal, and the plurality of function blocks stored in the selected function block are stored. The function relation is selected, and the corresponding selection command signals Z1 to Z6 are output. In the function blocks 80B to 85B, an operation for obtaining the values of the control forces Hc1 to Fc6 is performed according to the function data and the control program stored in the storage unit from the electric signal X1 and the selection command signals Z1 to Z6, and the output unit performs the control. The value of the force is an electrical signal a
Ｆf.

In the present embodiment configured as described above, for example, one of the selection switch elements 130a to 130e of the selection device 130 is intended for the purpose of loading earth and sand by a combined operation of turning and boom raising.
For example, when the selection switch element 130a is operated, the selection device 130 outputs an electric signal Za. Controller 6
In the function selection instruction block 131 of 1C, the electric signal Za
In addition to selecting the function blocks 80B and 83B based on the function block 80B, the function block 80B further selects the function indicated by the broken line So-2 shown in FIG. In addition, the first of a plurality of functional relationships
An operation for selecting the function indicated by the broken line So + 2 in FIG. 9 is performed, and the corresponding selection command signals Z1 and Z4 are output. Note that the other function blocks 81B, 82B, 84B, and 85B have the first
The basic function So shown in FIG. 9 is selected, and the corresponding selection command signals Z2, Z3, Z5, Z6 are output.

As a result, in the function blocks 80B and 83B, the functional relationship indicated by the selection command signals Z1 and Z4 is selected, and as in the above-described embodiment, during the earth and sand loading operation, the operation is compared with the boom cylinder 26 as compared with the normal control. A large flow rate can be supplied, and a relatively small flow rate can be supplied to the swing motor 23. Therefore, an optimum flow rate according to the earth and sand loading operation can be distributed to the boom cylinder 26 and the swing motor 23, and workability can be improved.

In addition, the selection device is designed to excavate arms and buckets with the aim of improving work efficiency compared to normal excavation work.
When one of the 130 selection switch elements 130a to 130e, for example, the selection switch element 130b is operated, the selection device 130 outputs an electric signal Zb. In the function selection instruction block 131 of the controller 61C, the function blocks 84B and 85B are selected based on the electric signal Zb, and the function block 84B is further provided with a broken line So shown in FIG. -1 is selected, and the function block 85B further includes a broken line So + 1 in FIG.
Is performed, and the corresponding selection command signals Z5 and Z6 are output.

As a result, in the function blocks 84B and 85B, the functional relationship indicated by the selection command signals Z5 and Z6 is selected.
By making the drive speed of the arm cylinder 27 relatively slow and the drive speed of the bucket cylinder 28 relatively slow, it is possible to realize a special excavation work that is considered to be better in terms of work efficiency than ordinary excavation.

Further, for example, when one of the selection switch elements 130a to 130e of the selection device 130, for example, the selection switch element 130c is operated for the purpose of shaping work to level the ground or the like by a combined operation of the arm and the bucket, the selection device 130 Outputs an electric signal Zc. In the function selection instructing block 131 of the controller 61C, the function blocks 84B and 85B are selected based on the electric signal Zc, and the function block 84B is further provided with a broken line So + 1 shown in FIG. And a function block 85B is further subjected to an operation of selecting a function indicated by a broken line So-1 in FIG. 19 among a plurality of functional relationships, and a corresponding selection command signal
Outputs Z5 and Z6.

As a result, in the function blocks 84B and 85B, the functional relationship indicated by the selection command signals Z5 and Z6 is selected, and the driving speed of the arm cylinder 27 is made relatively high while the bucket cylinder 28 By making the driving speed relatively slow, shaping work with good work efficiency can be realized.

In the above-described embodiment, the selection switch elements 130 of the selection device 130 are each configured to output a single selection command signal Za to Ze in accordance with the operation thereof. In the mode, the operation modes in which the speed ratios of the plurality of actuators 23 to 28 are different can be instructed. In the function selection instruction block 131, different function relations of the related function blocks are selected in accordance with the selection command signal to divide the flow. Can be changed, whereby the setting of the matching of the composite operation can be changed according to the work scene, and the workability and work efficiency can be further improved.

Other Embodiments of Control Pressure Generating Circuit In the above embodiments, the control pressure generating circuit controls the control pressures Pc1 to Pc6 according to the electric signals a to f from the controller.
As a control pressure generating means for outputting
Although the configuration adopting ~ 62f is adopted, other configurations may be adopted as the control pressure generating means. This embodiment shows the possibility of this point.

That is, in the present embodiment, the control pressure generation circuit 140
Are provided between the pilot pump 63 and the tank, and are connected in parallel with each other by electromagnetic variable relief valves 141a to 141a.
1f, and throttle valves 142a to 142f interposed between the electromagnetic variable relief valves 141a to 141f and the pilot pump 63, respectively.
The electromagnetic variable relief valves 141a to 141f are supplied with electric signals a to f, for example, from the controller 61 shown in FIG. 1. The electromagnetic variable relief valves 141a to 141f operate according to the electric signals. , Pilot lines 143a to 143 between throttle valves 142a to 142f and electromagnetic variable relief valves 141a to 141f.
3f is connected to, for example, the drive units 35c to 40c of the branch flow compensating valves 35 to 40 shown in FIG. 1 via the pilot lines 51a to 51f.

Also in the control pressure generating circuit 140 configured as described above, the electromagnetic variable relief valves 141a to 141f are individually driven according to the electric signals a to f output from the controller, the throttle amounts thereof are determined, and the pilot pump 63 The magnitude of the output pilot pressure is appropriately changed, and the electric signals a to
The control pressures Pc1 to Pc6 are supplied to the drive units 35c to 40c of the shunt compensating valves 35 to 40 shown in FIG. 1 through the pilot lines 143a to 143f and 51a to 51f, respectively, as shown in FIG. The same function as the proportional pressure reducing valve can be obtained.

Fourth Embodiment A fourth embodiment of the present invention will be described with reference to FIGS.

In FIG. 27, a hydraulic drive device applied to the hydraulic excavator of this embodiment is a single variable displacement hydraulic pump driven by a prime mover (not shown), that is, a main pump 200.
And a plurality of actuators driven by the pressure oil discharged from the main pump 200, namely, the swing motor 201 and the boom cylinder 202, and a flow control valve for controlling the flow of the pressure oil supplied to each of the plurality of actuators, That is, turning direction switching valve 203 and boom direction switching valve 204,
A pressure compensating valve arranged upstream of the flow control valve and controlling a differential pressure generated between an inlet and an outlet of the flow control valve, that is, a differential pressure before and after the flow control valve, that is, a diversion compensating valves 205 and 206, and It has.

Further, a relief valve and an unload valve (not shown) are connected to the discharge pipe line 207 of the main pump 200, and when the pressure oil from the main pump 200 reaches the set pressure of the relief valve, the relief valve causes the oil to flow out to the tank 208, The pump discharge pressure is prevented from becoming higher than the set pressure, and the unload valve causes the hydraulic oil from the main pump 200 to release the load pressure on the high pressure side of the swing motor 201 and the boom cylinder 202 (hereinafter referred to as the maximum). When the pressure reaches a value obtained by adding the set pressure of the unload valve to the load pressure (Pamax), the pressure is discharged to the tank 208, and the pressure is prevented from becoming higher than the pressure.

The discharge amount of the main pump 200 is controlled by a discharge amount control device 209.
The discharge pressure Ps is controlled to be higher than the maximum load pressure Pamax by a predetermined amount ΔPLSO, and load sensing control is performed.

The flow control valves 203 and 204 are hydraulic pilot type valves operated by pilot valves 210 and 211, respectively.The pilot valves 210 and 211 generate pilot pressure a1 or a2 and pilot valves b1 or b2 by manual operation of operation levers, and the flow control valves 203 and 204 , The pilot pressure a1 or a2 and the pilot pressure b1 or b2 are applied, and the flow control valves 203 and 204 are opened to the corresponding throttle amounts.

The shunt compensating valves 205 and 206 are of the same type as the shunt compensating valves in the first embodiment shown in FIG. That is, the outlet pressure and the inlet pressure of the flow control valves 203 and 204 are respectively guided, and the driving units 205a, 205b and 206a and 206b for applying the first control force based on the differential pressure in the valve closing direction, the springs 212 and 213, and the pilots A drive unit 205c, 206c through which the control pressure output from the electromagnetic proportional pressure reducing valves 216, 217 is guided via the lines 214, 215, and a valve opening direction in which the spring 212, 213 and the drive units 205c, 206c provide a target value of the differential pressure between the front and rear. Is applied.

Pilot pressure is supplied from a common pilot pump 220 to the discharge amount control device 209, the pilot valves 210 and 211, and the electromagnetic proportional pressure reducing valves 216 and 217. Flow control valves 203 and 204 include
A shuttle valve 222 for deriving the maximum load pressure of the swing motor 201 and the boom cylinder 202 is connected to each.

The hydraulic drive device according to the present embodiment further includes a main pump
The displacement corresponding to the displacement of 200 is detected and the main pump 2
00, a displacement detector 223 for detecting the discharge amount Qθ
The discharge pressure detector 224 for detecting the discharge pressure Ps of the second pump 200, the discharge pressure Ps of the main pump 200, and the maximum load pressure Pamax of the swing motor 201 and the boom cylinder 204 are introduced.
The differential pressure detector 225 for detecting LS, the detection signal from the displacement detector 223, the discharge pressure detector 224 and the differential pressure detector 225 are input, and the operation command signal is transmitted to the discharge amount control device 209 and the electromagnetic proportional pressure reducing valves 216 and 217. Controller that outputs S11, S12 and S21, S22
229.

FIG. 28 shows the configuration of the discharge amount control device 209. This embodiment is an example in which the discharge amount control device 209 is configured as an electro-hydraulic servo hydraulic drive device.

The discharge amount control device 209 has a servo piston 230 that drives the displacement mechanism 200a of the main pump 200. The servo piston 230 is housed in a servo cylinder 231. The cylinder chamber of servo cylinder 231 has servo piston 2
It is divided into a left room 232 and a right room 233 by 30,
The sectional area D of the left chamber 232 is formed larger than the sectional area d of the right chamber 233.

The left chamber 232 of the servo cylinder 231 is connected to a pilot pump 218 via lines 234 and 235, and the right chamber 233 is connected to a pilot pump 218 via a line 235.
Lines 234 and 235 are connected to tank 208 via return line 236. An electromagnetic valve 237 is provided on the line 235, and an electromagnetic valve 238 is provided on the return line 236. These solenoid valves 237 and 238 are normally closed (return to a closed state when not energized) solenoid valves. Operation command signals S11 and S12 from the controller 229 are input to these solenoid valves.
237 and 238 are thereby excited and switched to the open position.

When the operation command signal S11 is input to the solenoid valve 237 and switched to the open position, the left chamber 232 of the servo cylinder 231 communicates with the pilot pump 218, and the servo piston 230 is illustrated by the area difference between the right chamber 232 and the right chamber 233. Move to the right. As a result, the displacement angle of the displacement mechanism 200a of the main pump 200, that is, the displacement, increases, and the discharge amount increases.
When the operation command signal S11 disappears, the solenoid valve 237 returns to the original closed position, the communication between the left chamber 232 and the right chamber 233 is cut off,
Servo piston 230 is held stationary at that position. As a result, the displacement of the main pump 200 is kept constant, and the discharge amount becomes constant. When the operation command signal S12 is input to the solenoid valve 238 and switched to the open position, the left chamber 232 and the tank 208 communicate with each other, and the pressure in the left chamber 232 decreases. It is moved to the left in the figure. As a result, the displacement of the main pump 200 decreases, and the discharge amount also decreases.

As described above, the solenoid valves 237 and 238 are controlled to be turned on and off by the operation command signals S11 and S12, and the displacement of the main pump 200 is controlled.
Control is performed so as to be equal to the target discharge amount Qo calculated in 29.

The controller 229 has an input unit, a storage unit, a calculation unit, and an output unit, as in the first embodiment.

FIG. 29 is a functional block diagram showing the contents of the calculation performed by the calculation unit of the controller 229.

In FIG. 29, blocks 240, 241 and 242 are differential pressure gauges 43.
Is a block for obtaining the differential pressure target discharge amount QΔp which holds the differential pressure from the differential pressure ΔPLS detected by the load sensing compensation differential pressure, that is, the target differential pressure ΔPLSO. In this embodiment,
The differential pressure target discharge amount QΔp is obtained based on the following equation.

QΔp = g (ΔPLS) = ΣKI (ΔPLSO−ΔPLS) = KI (ΔPLSO−ΔPLS) + Qo−1 = ΔQΔp + Qo−1 (1) where KI: integral gain ΔPLSO: target differential pressure Qo−1: in the previous control cycle Output target discharge amount ΔQΔp: increment of differential pressure target discharge amount in one cycle time of control That is, differential pressure target discharge amount QΔp is calculated by an integral control method of a deviation between target differential pressure ΔPLSO and actual differential pressure. For example, blocks 240 and 241 calculate the differential pressure ΔPLS from KI (ΔPLSO−Δ
PLS) is calculated to obtain an increment ΔQΔp of the target differential pressure discharge amount per one control cycle time.
Then, the equation (1) is obtained by adding the ΔQΔp and the discharge amount target value Qo−1 output in the previous control cycle.

In this embodiment, QΔp is determined by the integral control method, but a different method, for example, QΔp = Kp (ΔPLSO−ΔPLS) (2) where Kp is a proportional control method represented by a proportional gain, or equation (1). And (2)
The ratio may be obtained by a proportional / integral control method obtained by adding equations.

The block 243 is a function block for determining the input restriction target discharge amount QT from the discharge pressure Ps of the main pump 200 detected by the pressure detector 224 and the input torque restriction function f (Ps) stored in advance. FIG. 30 shows the input torque limiting function f (Ps). The input torque of the main pump 200 is the displacement of the main pump 200, that is, the amount of tilt of the swash plate and the discharge pressure P.
It is proportional to the product of s. Therefore, the input torque limiting function f (Ps)
Uses a hyperbola or an approximate hyperbola. That is Here, it is a function represented by the equation of TP: input limiting torque κ: proportionality constant. The input restriction target discharge amount QT can be determined from the input torque restriction function f (Ps) and the discharge pressure Ps.

The magnitudes of the differential pressure target discharge amount QΔp and the input restriction target discharge amount QT determined as described above are determined by the minimum value selection block 204. If QΔp ≦ QT, QΔp is set as the discharge amount target value Qo. And if QΔp> QT, select QT. That is, the smaller of the differential pressure target discharge amount QΔp and the input restriction target discharge amount QT is selected as the discharge amount target value Qo, and the discharge amount target value Qo is set to the input torque restriction function (Ps).
Not to exceed the input limit target discharge amount QT determined by the above.

In blocks 255, 256 and 257, operation command signals S11 and S12 for the solenoid valves 237 and 238 of the discharge amount control device 209 are created from the discharge amount target value Qo obtained in the block 244 and the discharge amount Qθ detected by the displacement detector 223.

Specifically, first, in block 255, Z = Qo-Q
is calculated to determine a deviation Z between the target discharge amount Qo and the detected discharge amount Qθ. Next, in blocks 256 and 257, the operation command signal S11 or S12 is generated when the deviation Z exceeds a preset dead zone Δ according to the sign of the deviation Z. That is, when the deviation Z is positive and is equal to or greater than the dead zone Δ,
At 256, an operation command signal S11 is generated, and the discharge amount control device 20
The 9 solenoid valve 237 is turned ON. This, as described above,
The tilt angle of the main pump 200 is increased, and the discharge amount Qθ is controlled to match the discharge amount target value Qo. When the deviation Z is negative and becomes equal to or less than the dead zone Δ, an operation command signal S12 is generated in block 257, and the solenoid valve 237 is turned off and the solenoid valve 238 is turned on. As a result, the tilt angle of the main pump 200 is reduced, and control is performed such that the detected discharge amount Qθ matches the discharge amount target value Qo.

By controlling the tilt angle of the main pump 200 in this way, when the differential pressure target discharge amount QΔp is smaller than the input restriction target discharge amount QT, the discharge amount of the main pump 200 is controlled to be the differential pressure target discharge amount QΔp. Then, the pressure difference ΔPLS between the discharge pressure of the main pump 200 and the maximum load pressure is held at the target pressure difference ΔPLSO. That is, load sensing control for keeping the differential pressure ΔPLS constant is performed. On the other hand, the differential pressure target discharge amount QΔ
When p becomes larger than the input restricted target discharge amount QT, the input restricted target discharge amount QT is selected as the discharge amount target value Qo, and the discharge amount is controlled so as not to exceed the input restricted target discharge amount QT. That is, input restriction control of the main pump 200 is performed.

On the other hand, the differential pressure target discharge amount QΔp and the input restriction target discharge amount Q
T is deviated in block 258 to obtain a target discharge amount deviation ΔQ.

Next, in blocks 259, 260, and 261, the diversion compensating valves 205, 206 (second
Basic value for total consumable flow rate correction control
That is, the basic correction amount Qns is calculated. The total consumable flow rate correction control will be described later. In this embodiment, the basic correction value Qns
Is determined by an integral control method based on the following equation.

Qns = h (ΔQ) = ΣKIns · ΔQ = KIns · ΔQ + Qns−1 = ΔQns + Qns−1 (4) where KIns: integral gain Qns−1: basic correction value output in the previous control cycle ΔQns: control cycle time That is, in block 259, the increment ΔQns of the basic correction value per one control cycle time is obtained from KIns · ΔQ from the target discharge amount deviation ΔQ obtained in block 258. This value is added to the basic correction value Qns-1 output in the previous control cycle in an addition block 260 to obtain an intermediate value Q'ns, and Q'ns <0 is obtained in a block 261 having a limiter function shown in FIG. In this case, Qns = 0, and when Q'ns≥0, when Q'ns <Q'nsc, a basic correction value Qns that increases in proportion to the increase of Q'ns is output, and Q'ns≥Q ' At the time of nsc, the basic correction value Qns is determined so that Qns = Qnsmax. Here, Qnsmax and Q'nsc are values determined by the maximum tilt angle of the swash plate of the main pump 200, that is, the discharge capacity.

The basic correction value Qns obtained in the block 261 is further corrected in function blocks 262 and 263 provided for each of the actuators 201 and 202 to obtain different operation command signals S21 and S22.

FIG. 32 shows the relationship between the basic correction value Qns stored in the function blocks 262 and 263 and the operation command signals S21 and S22. In the figure, 264
Is a characteristic for the operation command signal S21, and 265 is a characteristic for the operation command signal S22. 266 is the basic correction value Qn
A characteristic that does not change s. That is, the operation command signal S21 is corrected to a value larger than the basic correction value Qns,
22 is corrected to a value smaller than the basic correction value Qns.

The operation command signals S21 and S22 obtained in blocks 262 and 263 are
The signals are output to the electromagnetic proportional pressure reducing valves 216 and 217 shown in the figure, and the electromagnetic proportional pressure reducing valves 216 and 217 are driven by this signal to generate a corresponding level of control pressure, which is output to the drive units 205c and 206c of the branch flow compensating valves 205 and 206. As a result, the shunt compensation valve 2
The second control force in the valve opening direction described above applied to the shunt compensation valve 205 becomes smaller than that in the case where the basic correction value Qns is output as a command signal.
Are corrected so as to increase, and the flow dividing ratio by the flow dividing valves 205 and 206 is corrected correspondingly.

Next, the operation of the present embodiment configured as described above will be described.

For example, when the boom pilot valve 211 is finely operated to drive the flow control valve 204 and the boom is operated alone,
Since the required flow rate is small, the controller 229 calculates the differential pressure target discharge amount QΔp smaller than the input restricted target discharge amount QT, and selects the differential pressure target discharge amount QΔp as the discharge amount target value Qo. For this reason, the differential pressure ΔPLS between the discharge pressure of the main pump 200 and the maximum load pressure becomes the target differential pressure ΔPLSO
Is carried out. on the other hand,
The basic correction value Qns is calculated to be zero, the diversion compensating valves 205 and 206 are given the second control force in the valve opening direction only by the forces of the springs 212 and 213, and the boom cylinder 202 has a flow rate corresponding to the opening of the flow control valve 204. Is supplied.

When the pilot valves 210 and 211 are simultaneously operated to perform, for example, a combined operation of turning and boom raising, since the required flow rate is large and the load pressure of the turning motor 201 is high, the controller 229 sets the differential pressure target discharge amount QΔp to the input restriction target. A value larger than the discharge amount QT is calculated, and the input restricted target discharge amount QT is selected as the discharge amount target value Qo. For this reason, the discharge amount of the main pump 200 is controlled so as not to exceed the input limited target discharge amount QT. That is, input restriction control of the main pump 200 is performed. At this time, the basic correction value Qns is calculated at the same time. If this basic correction value Qns is output as it is as an operation command signal to the electromagnetic proportional pressure reducing valves 216, 217,
The second control force in the valve opening direction of the branch flow compensating valves 205 and 206 decreases at the same rate, and the target value of the differential pressure across the flow control valves 203 and 204 decreases at the same rate. This allows the flow control valves 203, 204
The flow supplied to the
The total flow rate of the pressurized oil consumed in 201 and 202 can be reduced without changing the distribution ratio between the two, and the speed ratio between both actuators 201 and 202 can be maintained. This is referred to as total consumable flow rate correction control in this specification.

In the present embodiment, when the total consumable flow rate correction control is performed, the basic correction value Qns is further corrected to obtain different operation command signals S21 and S22,
Output to 6,217. For this reason, the second control force in the valve opening direction applied to the shunt compensating valves 205 and 206 is smaller at the shunt compensating valve 205 and larger at the shunt compensating valve 206 than when the basic correction value Qns is output as a command signal. While performing the total consumable flow rate correction control, the flow control is further performed such that the flow rate supplied to the swing motor 201 decreases and the flow rate supplied to the boom cylinder 202 increases. As a result, as in the case of the first embodiment,
The combined operation of turning and boom raising can be reliably performed, and the combined operation in which the boom raising speed is fast and the turning is relatively gentle is performed, so that the combined operability is improved and the energy can be effectively used.

As described above, also in this embodiment, substantially the same effect as in the first embodiment can be obtained in the combined operation of turning and boom.

Fifth Embodiment A fifth embodiment of the present invention will be described with reference to FIGS. In the drawing, the same members as those in the fourth embodiment shown in FIG. 27 described above are denoted by the same reference numerals.

In FIG. 33, the hydraulic drive device of this embodiment has basically the same configuration as that of the fourth embodiment shown in FIG. Therefore, the description of the same components will be omitted. In the discharge line 207 of the main pump 200, when the pressure oil from the main pump 200 reaches the relief set pressure, it flows out to the tank, and the relief valve 300 for preventing the pump discharge pressure from becoming higher than the set pressure is used. And the hydraulic oil from the main pump 200
When the pressure reaches the sum of the load pressure on the high pressure side of the boom cylinder 202 (hereinafter referred to as the maximum load pressure Pamax) and the unload set pressure, the pressure is discharged to the tank, and the unload that prevents the pressure from exceeding the pressure is reached. Valve 301 is connected.

The discharge amount of the main pump 200 is controlled by driving the swash plate 200a of the main pump 200 to increase or decrease the displacement and controlling the supply and discharge of the pressure oil to and from the drive cylinder 302a to adjust the displacement of the drive cylinder. And a discharge amount control device 302 including an electromagnetic control valve 302b. 303 is the swing motor 20
This is a relief valve that sets the swing relief pressure of 1.

The pilot valves 210 and 211 are provided with pilot pressure detectors 304 and 305 that detect that the pilot pressures a1 and a2 and the pilot pressures b1 and b2 are output from the pilot valves 210 and 211, respectively. Further, a selection device 306 that is operated by an operator and selects and sets a target value of the discharge pressure of the main pump 200 from outside is provided.

Displacement detector 223, discharge pressure detector 224, differential pressure detector 22
5. The detection signals from the pilot pressure detectors 304 and 305 and the selection device 306 are input to the controller 307, and after performing a predetermined calculation here, the electromagnetic control valve 302 of the discharge amount control device 302
b and the operation command signals S1 and S21 and S22 are output to the drive units 216c and 217c of the electromagnetic proportional pressure reducing valves 216 and 217.

The contents of the calculations performed by the controller 307 are shown in FIG. 34 in a functional block diagram. In the figure, block 310 is a differential pressure ΔPLS
Main pump 2 which holds the differential pressure ΔPLS from the target to the target differential pressure ΔPLSO
6 is a function block for calculating a target discharge amount Qo of 00. FIG. 35 shows the functional relationship between the differential pressure ΔPLS stored in the function block 310 and the target discharge amount Qo. This functional relationship is represented by the differential pressure ΔP
The relationship is such that the target discharge amount Qo increases in proportion to the decrease in LS. The target discharge amount Qo may be calculated by the integral control method as in the blocks 240 to 242 shown in FIG. 29 in the fourth embodiment.

The target discharge amount Qo is calculated by the displacement detector 2 in the addition block 311.
The deviation ΔQ from the discharge amount Qθ of the main pump 200 detected at 23
The deviation ΔQ is converted into an operation command signal S1 in an amplification output block 312 and output to the electromagnetic control valve 302b. As a result, the electromagnetic control valve 302b is driven, and the discharge pressure Ps is set to a fixed value ΔPLS that is higher than the maximum load pressure Pamax of the actuators 201 and 202.
The discharge amount of main pump 200 is controlled so as to increase by O.

A block 313 is a function block for obtaining a control force signal i1 from the differential pressure ΔPLS.
When the differential pressure ΔPLS does not reach the target differential pressure ΔPLSO even if the discharge amount of the main pump 200 is maximized at this time, the drive units 205c and 206c of the branch flow compensation valves 205 and 206 Control force Nc1, N2
And the second control force f-Nc1, Nc2 in the valve opening direction is reduced, that is, the target value of the differential pressure across the flow control valves 203, 204 is reduced, and the flow rate of the pressure oil supplied to each of the actuators 201, 202 is reduced. Although the increase in absolute amount is suppressed, the flow control valve
The pump discharge amount is distributed in accordance with the opening degree ratio between 203 and 204, that is, the required flow rate ratio. FIG. 36 shows the functional relationship between the differential pressure ΔPLS stored in the function block 313 and the control force signal i1. This functional relationship basically corresponds to the fourth embodiment of the first embodiment.
This is the same as the functional relationship for turning shown in FIG. Note that the control force signal i1 is used as a first command value of the control force Nc2 applied by the drive unit 206a to the shunt compensation valve 206.

Block 314 is a function block for obtaining, from the discharge pressure Ps of the main pump 200 detected by the discharge pressure detector 224, a control force signal i2 for holding the discharge pressure Ps at the target discharge pressure Pso by a proportional control method. The signal i2 is the control force Nc
Used to obtain a second command value of 2. The function block 314 is configured so that the target discharge pressure Pso can be changed by a command signal r from the selection device 306. FIG. 37 shows a functional relationship among the discharge pressure Ps, the control force signal i2, and the command signal r stored in the function block 314. In FIG. 37, the target discharge pressure of the functional relationship set when the command signal r is at the minimum value is indicated by Pso.

Blocks 315 and 316 are control force signals i for holding the discharge pressure Ps at the target discharge pressure Pso by an integral control method from the discharge pressure Ps of the main pump 200 detected by the discharge pressure detector 224.
The control force signal i3 is used together with the control force signal i2 to obtain a second command value of the control force Nc2. Here, in block 315, the rate of change 3 of the control force signal i3 is determined from the discharge pressure Ps based on a functional relationship stored in advance, and the rate of change 3 is integrated in block 316 to determine the control force signal i3. In block 315, similarly to block 314, the target discharge pressure Pso is the command signal r from the selection device 306.
Is configured to be changeable. FIG. 38 shows a functional relationship between the discharge pressure Ps stored in the function block 315, the rate of change 3 of the control force signal i3, and the command signal r. In addition,
Also in FIG. 38, the target discharge pressure of the functional relationship set when the command signal r is at the minimum value is indicated by Pso.

The control force signal i2 obtained in the function block 314 and the control force signal i3 obtained in the integration block 316 are added in an addition block 317, and the second command value of the control force Nc2 applied by the drive unit 206a of the shunt compensating valve 206 becomes Desired. The first command value i1 of the control force Nc2 obtained in the function block 313 and the second command value i2 + i3 of the control force Nc2 obtained in the addition block 317 are the minimum value selection block 3.
At 118, the magnitude is determined and the minimum value is selected.

On the other hand, the detection signals from the pilot pressure detectors 304 and 305 are input to an AND block 319, and the AND block 319 sends an ON signal to the switch block 320 when there is a detection signal of both the pilot pressure a1 or a2 and the pilot pressure b1 or b2. Output, and an OFF signal at other times.
Output to 20. Switch block 320 is AND block 31
When the OFF signal is output from 9, the signal is held at the illustrated position, the first command value i1 obtained in the function block 313 is selected, and when the ON signal is output from the AND block 319, the selection is performed in the block 318. Is selected, that is, the first command value i1 or the second command value i2 + i3. Accordingly, when one of the pilot valves 210, 211 is operated, that is, when the turning or the boom is operated alone, the first command value i1 is selected, and when both the pilot valves 210, 211 are operated, that is, when the turning and the boom are operated. In the case of a compound operation, the first command value i1
And the minimum value of the second command value i2 + i3 is selected.

The control force signal i1 as the command value of the control force Nc1 for the shunt compensating valve 205 obtained in the function block 313 is the amplification block 3
An operation command signal S <b> 21 is provided through the control signal 21 and output to the electromagnetic proportional pressure reducing valve 216. Further, the first command value i1 or the second command value i2 + i3 selected by the switch block 320 passes through the amplification block 322 and becomes the operation command signal S22 as the electromagnetic proportional pressure reducing valve 217.
Is output to

Next, the operation of the present embodiment configured as described above will be described.

For example, when operating the boom pilot valve 211 to drive the flow control valve 204 and operate the boom alone, the differential pressure ΔPLS between the discharge pressure Ps of the main pump 200 and the load pressure of the boom cylinder 202 is equal to the differential pressure. Detected by the detector 225, the corresponding target discharge amount Qo is calculated by the function block 310 in the controller 307, and the operation command signal S1 is calculated as described above.
Is output to the electromagnetic control valve 302b of the discharge amount control device 302, and the discharge amount is controlled so that the differential pressure ΔPLS matches the target differential pressure ΔPLSO.

At this time, in the function block 313, the differential pressure Δ
The control force signal i1 corresponding to the PLS is the control force N of the shunt compensation valve 206.
c1 is obtained as the first command value, and only the pilot valve 211 is operated and the OFF signal is output from the AND block 320.
Is selected, and this is output to the electromagnetic proportional pressure reducing valve 217 as an operation command signal S22. Therefore, a control force Nc2 corresponding to the control force signal i1 acts on the shunt compensation valve 206 in opposition to the force f of the spring 213, and the second control force f-i1 in the valve opening direction is applied to the shunt compensation valve 206. Is given. Here, the differential pressure ΔPLS
Is at the target differential pressure ΔPLSO, i.e., i1o
Is set so that the control force Nc2 corresponding to this corresponds to fo described in the first embodiment with reference to FIG. 4A, so that the flow dividing compensating valve 206 reduces the differential pressure across the flow control valve 204. Since the boom cylinder 202 is maintained at a predetermined value,
Is supplied with a flow rate corresponding to the opening degree of the flow control valve 204.
At this time, at the same time, an operation command signal S21 corresponding to the control force signal i1 is output to the electromagnetic proportional pressure reducing valve 216, and the shunt compensation valve
205 also operates to maintain a predetermined differential pressure.

Also in the independent operation of the swing driving the swing motor 201, the operation of the shunt compensation valves 205 and 206 is substantially the same as in the case of the independent operation of the boom described above.

When simultaneously operating the pilot valves 210 and 211, for example, performing a combined operation of turning and boom raising, the operator first operates the selection device 306 to output the corresponding command signal r, and adjusts the characteristics of the function blocks 314 and 315 of the controller 307. I do. That is, the target discharge pressure Pso of the main pump 200 is set to a value suitable for the combined operation of turning and raising the boom. More specifically, in this combined operation, since the swing body driven by the swing motor 201 is an inertial load, the swing motor 201 becomes an actuator on the high load pressure side, and the load pressure is normally set to the relief pressure set by the relief valve 303. To rise. From this, the target discharge pressure Pso is
Is set so as to be lower than the pressure obtained by adding the load sensing compensation differential pressure ΔPLSO to the relief pressure and to be higher than the pressure obtained by adding the differential pressure ΔPLSO to the load pressure of the boom cylinder 202.

Next, the pilot valves 210 and 211 are operated to operate the flow control valve 2.
Open 03,204 and start the combined operation of turning and boom raising. At this time, the discharge pressure Ps of the main pump 200 is increased by the load sensing control of the discharge amount control device 302, and in the process, if the discharge pressure Ps is going to become larger than the target discharge pressure Pso, the discharge pressure A relatively small control force signal i2 corresponding to Ps is determined, at the same time function block 315 and integration block 3
Also at 16, a relatively small control force signal i3 corresponding to the discharge pressure is obtained, and at an addition block 317, a relatively small addition value i2 + i3 is obtained.

On the other hand, at this time, since the main pump 200 is under load sensing control, the differential pressure ΔPLS is near the target differential pressure ΔPLSO, and the control block 313 of the controller 307 obtains the control force signal i1 corresponding to the differential pressure ΔPLSO. Can be

Here, the functional relationship between the block 313 and the functional relationships between the blocks 314 and 315 are as follows: the added value i2 + i3 when the discharge pressure Ps is near the target discharge pressure Pso, and the differential pressure ΔPLS is the target differential pressure ΔPLSO.
The mutual relationship is determined so that the control force signal i1 when it is in the vicinity is substantially equal. As a result, the discharge pressure Ps
Value i2 + i when exceeds the target discharge pressure Pso
3 is i1> i2 + i3 with respect to the control force signal i1 when the differential pressure ΔPLS is near the target differential pressure ΔPLSO, and the minimum value selection block 318 selects the added value i2 + i3, that is, the second command value.

In this case, since both pilot valves 210 and 211 are operated, an ON signal is output from the AND block 319, and the switch block 320 is switched to a position for selecting the output of the minimum value block 318. Therefore, in the switch block 320, the second command value i2 + i3 is selected, and this is used as the operation command signal S22 as the electromagnetic proportional pressure reducing valve 217.
Is output to Further, an operation command signal S21 corresponding to the control force signal i1 is output to the electromagnetic proportional pressure reducing valve 216.

As a result of the operation command signals S21 and S22 being output,
As the second control force Nc1 in the valve opening direction, f
−i1 is applied, and f− (i2 + i3) is applied to the diversion compensation valve 206 as the second control force Nc2 in the valve opening direction. Here, f− (i2 + i3)> fi1. For this reason, at the start of the combined operation of turning and boom raising, the degree to which the shunt compensation valve 206 related to the boom cylinder 202 on the low load pressure side is throttled becomes small, and the normal control force Nc2 = i1 is applied to the boom cylinder 204. Is supplied more than in the case where is given. As a result, an increase in the discharge pressure of the main pump 200 is suppressed, and the discharge pressure is balanced around the target discharge pressure Pso. Also, at this time, the flow rate of the pressure oil supplied to the boom cylinder 202 increases, and the discharge pressure is suppressed from increasing above Pso. The load pressure is reduced as compared with the case where the load pressure rises to the relief pressure, and the swing motor 201 is driven at an appropriate speed without the relief of the pressure oil. This makes it possible to perform a combined operation of turning and boom raising, in which the boom raising speed is high and the turning speed is relatively slow, and energy loss during turning acceleration can be reduced.

In the combined operation of turning and boom raising as described above, when the turning is accelerated and reaches a steady speed, the turning motor 20 is turned on.
The load pressure of 1 decreases, and accordingly, the discharge pressure of the main pump 200 that is under load sensing control also decreases, and becomes equal to or less than the target discharge amount Pso. When the discharge pressure becomes equal to or less than the target discharge amount Pso, the value of the control force signal i2 obtained in the function block 314 and the value of the control force signal i3 obtained in the blocks 314 and 316 increase, and the second command value i2 obtained in the addition block 318 is obtained.
+ I3 also becomes relatively large, and from the above-described functional relationship of block 313 and the functional relationship of blocks 314 and 415, i1
<I2 + i3. Therefore, in the minimum value selection block 318, the first command value i1 is selected, and the operation command signal S22 corresponding to the first command value i1 is output to the electromagnetic proportional pressure reducing valve 217.

As a result, the conventional f-i1 is applied to the branching compensating valve 206 as the second control force Nc2 in the valve opening direction, and at this time, the branching compensating valve 205 also has the same valve opening direction. The second control force f-i1 is applied. As a result, the differential pressure across the flow control valves 203 and 204 is controlled so as to be equal, and the swing motor 201 and the boom cylinder 202 are supplied with the flow rates required by the pilot valves 210 and 211. That is, the flow rate of the pressure oil supplied to the turning motor 201 increases, and a desired turning speed can be obtained. In this way, after turning acceleration, a composite operation intended by an operator having a relatively high turning speed can be realized.

As described above, in the present embodiment, the boom cylinder 202 which is an actuator for driving a load having a small inertia is used.
The discharge pressure of the main pump 200 is arbitrarily controlled by controlling the flow rate supplied to the main pump 200, and the drive pressure of the swing motor 201, which is an actuator for driving a load with a large inertia, is controlled. Similarly, in the combined operation of turning and boom raising, the boom raising speed is fast and the turning speed can be relatively slow, and the operability can be improved,
Energy loss during combined operation can be reduced, and economical operation is possible.

Further, according to the present embodiment, the characteristics of the function blocks 314 and 315 are appropriately changed by operating the selection device 306, and the main pump 200
The target discharge pressure Pso can be changed, so that matching between turning and boom raising can be appropriately set.

In the above embodiment, in order to achieve both control responsiveness and stability, the controller 307 obtains a control force signal for controlling the discharge pressure Ps to be maintained at the target value Pso as a function of the proportional control method. Although both the block 314 and the function blocks 315 and 316 of the integral control method are used, it is apparent that either one may be used to obtain the control force signal.

Sixth Embodiment A sixth embodiment of the present invention will be described with reference to FIGS. In the figure, the fourth embodiment shown in FIG.
Members equivalent to those in the fifth embodiment shown in the figure are denoted by the same reference numerals.

In FIG. 39, the hydraulic drive device of this embodiment has basically the same configuration as that of the fourth embodiment shown in FIG. 27, and a description of that portion will be omitted. However, the output signal from the differential pressure detector 225 that detects the differential pressure ΔPLS between the discharge pressure Ps of the main pump 200 and the maximum load pressure Pamax is represented by Edp. Also, as in the fifth embodiment shown in FIG. 33, when the pressure oil from the main pump 200 reaches the relief set pressure, it flows into the tank via the discharge line 207 of the main pump 200, and the pump discharge pressure becomes A relief valve 300 for preventing the pressure from becoming higher than the set pressure is provided, and the pressure oil from the main pump 200 is supplied with the load pressure on the high pressure side of the swing motor 201 and the boom cylinder 202 (hereinafter referred to as the maximum load pressure Pamax When the pressure reaches the sum of the unload set pressure and
An unload valve (not shown) is provided to prevent the pressure from exceeding the pressure.

Further, the main pump 200 is provided with a displacement detector 223 for detecting the displacement, and the displacement detector 223 outputs a signal Eθ corresponding to the detected displacement. The discharge amount of the main pump 200 is controlled by the discharge amount control device 302 of the fifth embodiment.
Discharge control device of load sensing control method corresponding to
The discharge rate control device 400 is controlled by the main pump 20
The tilt drive device 400a drives the zero swash plate 200a to increase or decrease the displacement, and an electromagnetic proportional pressure reducing valve 400b outputs a control pressure to the tilt drive device and adjusts the displacement.

Then, an operation of detecting that pilot pressure is applied to the pilot lines 401a and 401b for guiding the pilot pressure from the turning pilot valve (not shown) to the drive unit of the flow control valve 203, respectively, and outputting signals E402 and E403. Detectors 402 and 403 are provided. Further, a selection device 406 which is operated by the operator and selects and sets the flow rate increase rate of the pressure oil supplied to the swing motor 201 is provided, and the selection device 406 outputs a signal Es according to the setting at that time. .

The signal Edp from the differential pressure detector 225, the signals E402 and E403 from the operation detectors 402 and 403, the signal Es from the selector 406, and the signal Eθ from the displacement detector 223 are input to the controller 407, where a predetermined calculation is performed. After performing, the electromagnetic proportional pressure reducing valve 216,
217, and output the operation command signal E400 to the electromagnetic proportional pressure reducing valve 400b of the discharge amount control device 400.
Is output.

The selection device 406 is, as shown in FIG. 40 in this embodiment,
It comprises a voltage setting device including a variable resistor 408. When the position of the movable contact is changed by an operation of an operator, a voltage of a level corresponding to this is set. This voltage value is taken into the controller 407 as a signal Es, and the controller 407 converts the signal Es from A / D and sends it to the CPU. The CPU reads the A / D conversion value of this signal Es in step S1 and sets ΔE = A / D conversion value in step S2, as shown in the flowchart of FIG. A change amount ΔE per cycle of the signal E216 is obtained. This change amount ΔE is determined by the controller 40
7 is used to determine the operation command signal E216.

FIG. 42 is a flowchart showing the content of the calculation performed by the controller 407. This flow chart shows the proportional solenoid pressure reducing valve
FIG. 9 shows a calculation procedure of operation command signals E216 and E217 for 216 and 217.
The method of obtaining the operation command signal E400 with respect to b is substantially the same as the method of obtaining the operation command signal S1 in the fifth embodiment shown in FIG. 34, and a description thereof will be omitted.

First, signals Edp, E402, E403, and Es are read in step S10. Next, in step S11, the differential pressure signal Edp
And the functional relationship stored in advance, the electromagnetic proportional pressure reducing valves 216, 21
The basic drive signal EHL of 7 is calculated. This basic drive signal EH
L is the load sensing control of the main pump 200 by the discharge amount control device 400. At this time, when the differential pressure ΔPLS does not reach the target differential pressure ΔPLSO even if the discharge amount of the main pump 200 is maximized, the shunt compensation valves 205 and 206 The control forces Nc1 and Nc2 applied by the driving units 205c and 206c are increased, and the second control forces f-Nc1 and Nc2 in the valve opening direction are reduced, that is, the target value of the differential pressure across the flow control valves 203 and 204 is reduced. The flow rate of the pressure oil supplied to each of the actuators 201 and 202 is distributed in accordance with the opening ratio of the flow control valves 203 and 204, that is, the required flow rate ratio, although the increase in the absolute amount is suppressed. FIG. 43 shows the differential pressure ΔPLS for obtaining the basic drive signal EHL and the drive signal EHL.
This shows the functional relationship with. This functional relationship is substantially the same as the relationship between the differential pressure ΔPLS and the control force signal i1 shown in FIG. 36 described above.

Next, in step S12, it is determined whether or not the operation command signal E402 or E403 has been input. If not, the process proceeds to step S13, where the drive signal EH of the electromagnetic proportional pressure reducing valve 216 is set to EH = EHMAX. Here, EHMAX is the maximum value of the drive signal EH. At this time, the control force Nc1 of the drive unit 205c becomes the maximum, and the shunt compensation valve resists the force f of the spring 212.
Hold 205 in the fully closed position. Operation command signal E402 or E40
If 3 is input, the process proceeds to step S14, where EHL <EH-
It is determined whether 1−ΔE. That is, it is determined whether or not the drive signal EHL is smaller than a value obtained by subtracting the change amount ΔE set by the selection device 406 from the drive signal EH-1 of the electromagnetic proportional pressure reducing valve 216 obtained in the previous control cycle.
If it is determined that EHL is smaller than EH-1−ΔE, the process proceeds to step S15, where EH = EH−1−ΔE, and EH−1−ΔE is set.
If it is determined that it is larger than ΔE, the process proceeds to step S16, where EH
= EHL. That is, the maximum change speed of the drive signal EH is Δ
The drive signal EH is determined so as to match E.

Subsequently, EH-1 = EH is set in step S17, the drive signal EH is output as the operation command signal E216 in step S18, and the basic drive signal EHL is output as the operation command signal E217 in step S19. As a result, the control force Nc1 applied by the drive unit 205c of the shunt compensation valve 205 is controlled to match the basic drive signal EHL, and the rate of change is limited to ΔE or less. The control force Nc2 applied by the drive unit 206c of the shunt compensation valve 206 is controlled so as to match the basic drive signal EHL as in the related art.

Next, the operation of the present embodiment configured as described above will be described.

First, without operating any of the flow control valves and during non-operation when the actuator is not driven, in the controller 407, since the operation detection signal E402 or E403 is not input, in the step S12 of the flowchart shown in FIG. 42, A determination of NO is made, and in step S13, the drive signal EH of the electromagnetic proportional valve pressure valve 216 is set to the maximum value EHMAX. Therefore, the flow compensating valve 205 is held at the fully closed position. On the other hand, for the electromagnetic proportional pressure reducing valve 217, the basic drive signal EHL is set as the operation command signal E217. At this time, the set pressure (>) is set by an unload valve (not shown).
Since the discharge pressure Ps of the main pump 200 corresponding to ΔPLSO) is secured, a relatively small basic drive signal EHL is obtained in step S11 from the functional relationship shown in FIG. Is held at the fully open position by the force f.

When operating the boom boom pilot valve (not shown) to drive the flow control valve 204 and perform independent operation of the boom,
The differential pressure ΔPLS between the discharge pressure Ps of the main pump 200 and the load pressure of the boom cylinder 202 is detected by the differential pressure detector 225, and the controller 407 calculates an operation command signal E400 for keeping the differential pressure ΔPLS constant, and Control device 400 controls the discharge amount of main pump 200 according to operation command signal E400.

At this time, the controller 407 calculates operation command signals E216 and E217 for the electromagnetic proportional pressure reducing valves 216 and 217. Here, in this case, since the turning flow control valve 203 is not driven, the operation detection signal E402 or E403 is not input, and similarly to the case of the above-described non-operation, the electromagnetic proportional valve pressure valve 216 is operated. The drive signal EH is set to the maximum value EHMAX, and the shunt compensation valve 205 is held at the fully closed position. On the other hand, the basic drive signal EHL corresponding to the differential pressure ΔPLS near the target differential pressure ΔPLSO is calculated from the functional relationship shown in FIG. EHL is output to the electromagnetic proportional pressure reducing valve 217 as an operation command signal E217. Here, the forty-third functional relationship is substantially the same as the functional relationship shown in FIG. 36 described above. Therefore, the branch flow compensating valve 206 is held at the fully open position by the second control force of f-Nc2 against the first control force in the valve closing direction based on the pressure difference between the front and rear of the flow control valve 204, and the boom cylinder 202 Is supplied with a flow rate corresponding to the opening degree of the flow control valve 204.

Single operation of the swing motor 201, or the flow control valves 203, 204
Simultaneously, for example, when performing a combined operation of turning and boom raising, the operator first operates the selection device 406 to output the flow rate increasing speed signal Es, and as described above, one cycle of the operation command signal E216 The change amount ΔE per hit is set. Specifically, the change amount ΔE is set to a small value when the turning acceleration is to be performed slowly, and is set to a large value when the turning acceleration is desired to be quick.

Next, the flow control valve 203 alone or the flow control valve 2
Both the 03 and the flow control valve 204 are simultaneously driven to start a single operation of turning or a combined operation of turning and boom raising.
At this time, the discharge pressure Ps of the main pump 200 is
Due to the load sensing control of 00, the pressure rises while maintaining the differential pressure ΔPLSO.

At the same time, the controller 407 calculates operation command signals E216 and E217 for the electromagnetic proportional pressure reducing valves 216 and 217. Here, in this case, the flow control valve for turning
Since the drive 203 is driven and the operation detection signal E402 or E403 is input, a determination of YES is made in step S12 shown in FIG. 42, and the drive signal EH is obtained by the calculations in steps S14 to S16. That is, the drive signal EH that limits the change speed to ΔE or less using the basic drive signal EHL as the target value is obtained. Then, the drive signal EH is output to the electromagnetic proportional valve 216 as an operation command signal E216, and the shunt compensation valve 205 gradually starts to open from the fully closed position at a speed corresponding to the change amount ΔE. Is supplied to the swing motor 201 at a flow rate increasing speed corresponding to the change amount ΔE. Thus, the swing motor 201 is driven at an acceleration corresponding to the change amount ΔE.

FIG. 44 shows the relationship between the time t during the turning operation, the drive signal EH, and the flow rate increasing speed signal Es. After the start of turning, the drive signal EH decreases at a gradient corresponding to the change amount ΔE. The gradient increases as the flow rate increase speed signal Es, that is, the change amount ΔE increases. This gradient also corresponds to the rate of increase in the flow rate of the pressure oil supplied to the swing motor 201, that is, the drive acceleration of the swing motor 201.

On the other hand, as in the case of the single operation of the boom, the 43rd flow compensation valve 206 for the boom is
From the functional relationship shown in the figure, the differential pressure ΔPL near the target differential pressure ΔPLSO
The basic drive signal EHL corresponding to S is calculated, and this basic drive signal EHL is used as the operation command signal E217 as the electromagnetic proportional pressure reducing valve 21.
Output to 7. That is, the control force Nc2 corresponding to the signal E217 is applied to the flow dividing compensating valve 206 in the valve opening direction in opposition to the force of the spring 213. Thus, in the case of a single operation of turning, the branch flow compensating valve 206 is held at the fully open position by the second control force of f-Nc2. In the case of a combined operation of turning and boom raising, the boom cylinder 202 serves as an actuator on the low load pressure side.
Narrow down to keep at Nc2.

Then, in the case of the combined operation of turning and balm raising, the turning operation is started as described above, and in the process of increasing the turning speed, the discharge amount of the main pump 200 reaches the maximum and the differential pressure Δ
When the PLS decreases, the value of the basic drive signal EHL calculated in step S11 in FIG.
205 and 206 limit the absolute amount of the pressure oil supplied to the actuators 201 and 202, and the distribution of the flow rate is controlled so as to be appropriate.

When the turning reaches a speed corresponding to the opening degree (required flow rate) of the flow control valve 203 after the start of the turning operation, the control force Nc1 applied by the driving unit 205c of the shunt compensation valve 205 is a drive signal calculated in step S11. The value reaches EHL, and EH = EHL is always calculated in step S16. Therefore, at this time, the second control forces f-Nc1 and f-Nc2 in the valve-opening directions of the flow dividing compensating valves 205 and 206 become equal, and in the case of the combined operation of turning and boom raising, the respective actuators
A flow rate proportional to the opening of the flow control valves 203 and 204 is supplied to 201 and 202, and a combined operation of turning and boom raising can be performed at a required speed ratio.

As described above, in the present embodiment, at the start of the turning operation, the flow rate increase rate of the pressure oil supplied to the turning motor 201 can be arbitrarily set. At the beginning of the operation, the flow ratio of the pressure oil supplied to both actuators is arbitrarily changed,
The compound operation can be performed at the optimum speed ratio for the work.

In addition, at the start of the turning operation, the flow rate of the pressure oil supplied to the turning motor 201 can be arbitrarily set, so that a sharp increase in the turning load pressure is suppressed, and the pressure that is throttled by the turning relief valve is reduced. Oil is reduced and energy loss can be reduced. Further, when the setting of the flow rate increasing speed is relatively small, the driving pressure of the swing motor can be suppressed to the relief pressure or less, so that the energy loss can be further reduced and the discharge pressure of the main pump 200 can be reduced. When the main pump 200 is under horsepower limit control (input torque limit control), the discharge amount can be increased in accordance with the reduction of the discharge pressure, and the supply amount of pressurized oil to the boom cylinder increases.
The driving speed can be increased.

Modification of Sixth Embodiment A first embodiment of the sixth embodiment will be described with reference to FIGS. 45 and 46. This embodiment shows a modification of the selection device.

In FIG. 45, the selection device 406A has four contacts A to D
And a switching device including a movable contact 409 with respect to. The contacts A to C are connected to the input terminals Di1, Di2, Di3 of the CPU in the controller 407A, and the input terminals Di1, Di2, D
i3 is connected to a power supply via resistors 410a, 410b, 410c. With such a configuration, when the movable contact 409 is at a position where it contacts the contact C as shown in the figure, for example, the input terminal Di1 is grounded, the voltage becomes 0, and the other input terminals Di2, D2
i3 is kept in a state where the power supply voltage is applied.

In the controller 407A, the input terminals Di1, Di2, Di3
46, the flow rate increasing speed is set as shown in FIG. In step S20, the voltage of the input terminal Di3 becomes 0
And if it is 0, in step S21,
A change amount ΔE per one cycle of the operation command signal E216 for the electromagnetic proportional pressure reducing valve 216 is set to a value ΔEA stored in advance. If the voltage of the input terminal Di3 is not 0, step S22
To determine whether the voltage of the input terminal Di2 is 0,
In the case of, the change amount ΔE is set to the value ΔEB stored in advance in step S23. If the voltage of the input terminal Di2 is not 0, the process proceeds to step S24, where it is determined whether the voltage of the input terminal Di1 is 0. If the voltage is 0, the change amount ΔE is set to a previously stored value ΔEC in step S25. Finally, if the voltage of the input terminal Di1 is not 0, the process proceeds to step S26,
The change amount ΔE is set to a value ΔED stored in advance.

As described above, by switching the position of the movable contact 409, the change amount ΔE corresponding to the position can be set.

Next, a second modification of the sixth embodiment will be described with reference to FIGS.
This will be described with reference to the drawings. 47, the same steps as those shown in FIG. 42 are denoted by the same reference numerals. In the present embodiment, the flow rate increasing speed control for the turning motor 201 is performed only during the combined operation of turning and boom raising.

In the hydraulic drive device of the present embodiment, as shown by imaginary lines in FIG. 39, among the pilot lines 404a and 404b for guiding pilot pressure from a boom pilot valve (not shown) to the drive unit of the flow control valve 204, An operation detector 405 for detecting that pilot pressure has been applied to the pilot line 404a corresponding to the raising and outputting a signal E405 is further provided. The signal E405 is sent to the controller 407.

In the controller 407, step S30 shown in FIG. 47
In addition to the signals Edp, E402, E403, Es, the operation detector
The detection signal E405 from 405 is read. And step S
In addition to the judgment of 12, the operation detection signal E
It is determined whether or not 405 has been input. When this is also satisfied, the process proceeds to steps S14 to S16 for the first time, and the basic drive signal EHL
Drive signal EH that limits the variation to ΔE or less with
Is calculated.

According to the present embodiment, it is possible to obtain an effect that it is possible to control the rate of increase in the flow rate of the pressure oil supplied to the turning motor only during the combined operation of turning and boom raising, thereby enabling control of the turning acceleration.

〔The invention's effect〕

In the hydraulic drive device for a construction machine according to the present invention, since it is configured as described above, individual pressure compensation characteristics are given to the first and second branch flow compensation valves, and the first and second actuators are simultaneously driven. In a combined operation, an optimum shunt ratio according to the type of the actuator can be provided, and operability and / or work efficiency can be improved.

[Brief description of the drawings]

FIG. 1 is a circuit diagram showing an entire hydraulic drive device for construction equipment according to a first embodiment of the present invention, FIG. 2 is a schematic diagram showing a configuration of a controller, and FIG. 3 is performed by the controller. FIG. 4B is a functional block diagram showing the content of the calculation,
FIG. 4B is a diagram showing the functional relationship between the differential pressure ΔPLS and the value of the control force Fc1 to be applied to the shunt compensating valve related to the swing motor. FIG. FIG. 4C is a diagram showing a functional relationship between the control force Fc2 and the value of Fc3 to be applied to the control valve. FIG. 4D is a diagram showing a functional relationship, and FIG. FIG. 5 is a diagram collectively showing the functional relationships shown in FIGS. 4A to 4D, FIG. 6 is a diagram showing a functional relationship between the oil temperature Th and the correction coefficient K, and FIG. FIG. 8 is a side view of a hydraulic shovel to which the hydraulic drive device of the present embodiment is applied, and FIG. 8 is a top view of the hydraulic shovel. , FIG. 9 -
FIG. 12 is a diagram showing four modified examples of the functional relationship between the differential pressure ΔPLS and the value of the control force Fc1 to be applied to the shunt compensating valve relating to the swing motor, and FIG. 13 and FIG. FIG. 15 is a diagram showing two modified examples of the functional relationship between the differential pressure ΔPLS and the values of the control forces Fc2 and Fc3 to be applied to the shunt compensating valve relating to the traveling motor, and FIG. FIG. 16 is a circuit diagram showing the entire hydraulic drive device according to the embodiment, FIG. 16 is a functional block diagram showing the contents of calculations performed by a controller, and FIG. 17 is a hydraulic drive device according to a third embodiment of the present invention. 18 is a functional block diagram showing the contents of the operation performed by the controller, FIG.
FIG. 19 is a diagram showing a plurality of functional relationships between the differential pressure ΔPLS and the control forces Fc1 to Fc6, and FIG. 20 collectively shows the functional relationships selected when performing a combined operation of turning and boom raising. FIG. 21 is a diagram showing the relationship between the differential pressure across the boom flow control valve and the supply flow rate when performing the combined operation, and FIG. 22 is a diagram illustrating the relationship when performing the combined operation. FIG. 9 is a diagram showing a relationship between a differential pressure across the flow control valve for turning and a supply flow rate,
FIG. 23 is a diagram collectively showing a functional relationship selected when performing a combined operation of an arm and a bucket intended for special excavation work, and FIG. 24 is an arm intended for a shaping work for leveling the ground or the like. FIG. 25 is a diagram collectively showing a functional relationship selected when performing a compound operation of a bucket and a bucket, and FIG. 25 is a functional block diagram showing contents of an operation performed by a controller in a modification of the third embodiment. FIG. 26 is a circuit diagram showing another embodiment of the control pressure generating circuit, FIG. 27 is a circuit diagram showing a hydraulic drive device according to a fourth embodiment of the present invention, and FIG. FIG. 29 is a schematic diagram showing the configuration of the amount control device, FIG. 29 is a functional block diagram showing the contents of calculations performed by the controller, and FIG. 30 is a diagram showing the relationship between the discharge pressure and the input restricted target discharge amount. FIG. 31 shows that the intermediate correction value Qns Obtaining a diagram showing a limiter function, FIG. 32, the basic correction value Qns and the operation command signal S2
FIG. 33 is a diagram showing the relationship with S1 and S22, and FIG.
FIG. 34 is a circuit diagram showing a hydraulic drive device according to an example of the thirty-fourth embodiment.
FIG. 35 is a functional block diagram showing the contents of calculations performed by the controller. FIG. 35 is a diagram showing a functional relationship between the differential pressure ΔPLS and the target discharge amount Qo. FIG. 36 is a diagram showing the functional relationship between the differential pressure ΔPLS and the control. FIG. 37 is a diagram showing a functional relationship between a force signal i1 and FIG. 37 is a diagram showing a functional relationship between a discharge pressure Ps, a control force signal i2 and a command signal r, and FIG. 38 is a diagram showing a discharge pressure Ps and a control FIG. 9 is a diagram showing a functional relationship between a change rate 3 of a force signal i3 and a command signal r.
FIG. 39 is a circuit diagram showing a hydraulic drive device according to a sixth embodiment of the present invention, FIG. 40 is a diagram showing a configuration of a selection device, and FIG. 41 is a change amount according to operation of the selection device. FIG. 42 is a flowchart showing a procedure for obtaining ΔE, FIG. 42 is a flowchart showing the contents of calculation performed by the controller, and FIG. 43 is a diagram showing a functional relationship between the differential pressure ΔPLS and the basic drive signal EHL; FIG. 44 is a diagram showing a relationship between the time t at the start of the turning operation, the drive signal EH, and the flow rate increasing speed signal Es, and FIG. 45 is a selector according to a first modification of the sixth embodiment. FIG. 46 is a flowchart showing a procedure for obtaining a change amount ΔE in accordance with an operation of the selection device. FIG. 47 is a flowchart showing a controller according to a second modification of the sixth embodiment. 5 is a flowchart showing the contents of the calculation performed in step S1. 1 to 14 22... Main pumps 23 to 28... Hydraulic actuators 29 to 34... Flow control valves 35 to 40. Spring (driving means) 35c to 40c: driving unit (driving means) 59: differential pressure detector (first means) 61: controller (second means) 62a to 62f: electromagnetic proportional pressure reducing valve (control Pressure generating means) 80-85 Function block (first operation means) 90-92 Delay element block (second operation means) 60 Oil temperature detector 86 Function block (sixth operation Means) 87 to 89 Multiplication block (fourth arithmetic means) Fig. 15 45A to 50A ... Drive unit (drive means) 63 ... Pilot pump (pilot pressure supply means) 64 ... Relief valve (pilot pressure supply) Means 113... Pressure reducing valve (pilot pressure supply means) FIGS. 17 to 25 120... Selection device (fourth means) 61B ... Controller (second means) 80B to 85B ... Function block (fifth calculation means) Figs. 27 to 32 200 ... Main pump 201 ... Swing motor (first actuator) 224 ... Discharge Pressure detector (fifth means) 225 ... Differential pressure detector (first means) 205,206 ... Shunt compensation valve 229 ... Controller (second means) 240-242 ... Block (tenth arithmetic means) ) 243 Block (eleventh arithmetic means) 258 Block (twelfth arithmetic means) 259-263 Block (thirteenth arithmetic means) FIGS. 6 means) 307 controller (second means) 313 function block (sixth operation means) 314,315 function block (seventh operation means) 316 integration block (seventh operation means) 317 addition block (seventh operation means) 318 minimum value selection block (seventh operation means) FIG. 39 47 Figure 402,403,405 ...... operation detector (seventh means) 406 ...... selection device (eighth means) 407 ...... controller (second means)

Continuing on the front page (72) Inventor Yukio Aoyagi 650 Kandate-cho, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. (72) Inventor Hajime Yasuda 650, Kandate-cho, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd., Tsuchiura Plant (72) Inventor Hiroshi Watanabe 650, Kanda-cho, Tsuchiura-shi, Ibaraki Prefecture Hitachi Construction Machinery Co., Ltd. Tatsuura Izumi, Kachidate-cho, Tsuchiura-shi, Ibaraki Pref., Within the Tsuchiura Plant of Hitachi Construction Machinery Co., Ltd. (72) Inventor Yasuo Tanaka 650, Kunitachi-cho, Tsuchiura-shi, Ibaraki Pref. 650 Tsuchiura-cho, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. Tsuchiura plant (72) Inventor Shigetaka Nakamura 650 Tsuchiura-cho, Tsuchiura-city, Ibaraki prefecture Hitachi Construction Machinery Co., Ltd. Tsuchiura plant (56) References JP-A 60-11706 (JP, A) JP-A-62-75107 (JP, A) JP-A-61-28639 (JP, A) Open Akira 57-200703 (JP, A) JP Akira 63-186003 (JP, A)

Claims (15)

(57) [Claims] 1. A hydraulic pump, at least a first hydraulic actuator and a second hydraulic actuator driven by hydraulic oil supplied from the hydraulic pump, and a flow of hydraulic oil supplied to the first and second actuators. First to control each
And a second flow control valve, first and second flow compensating valves for controlling a first differential pressure generated between an inlet and an outlet of the first and second flow control valves, respectively, and the hydraulic pump Discharge amount control means for controlling a flow rate of pressure oil discharged from a hydraulic pump in response to a second differential pressure between the discharge pressure of the first and second actuators and the maximum load pressure of the first and second actuators. A construction machine having driving means for applying a control force based on the second differential pressure to a corresponding diversion compensating valve and setting a target value of the first differential pressure, respectively; A first means for obtaining the second differential pressure from a discharge pressure of the hydraulic pump and a maximum load pressure of the first and second actuators; and at least a first means for obtaining the second differential pressure. Based on the second differential pressure, the first and second A second means for calculating an individual value as a value of the control force to be applied by each drive means of the flow compensating valve; and a first means provided corresponding to each of the first and second flow compensating valves. And second control pressure generating means for generating control pressures according to the individual values obtained by the second means, respectively, and applying the control pressures to the first and second flow dividing compensation valve driving means. And a first and a second control pressure generating means for respectively outputting the control pressure. 2. The hydraulic drive system for a construction machine according to claim 1, wherein said second means includes a second differential pressure calculated by said first means and said first and second branch flow compensation valves. A first calculating means for obtaining values of first and second control forces corresponding to the second differential pressure from first and second functions correspondingly set in advance; Drive. 3. The hydraulic drive system for a construction machine according to claim 2, wherein said first actuator is an actuator for driving an inertial load, and said second actuator is an actuator for driving a normal load. The first and second functions define the second differential pressure and the first and second differential pressures such that the target value of the first differential pressure decreases as the second differential pressure decreases and the rate of decrease is different between the two. 2. A hydraulic drive device, wherein a relationship with a value of the control force is set. 4. The hydraulic drive system for a construction machine according to claim 2, wherein said first actuator is an actuator for driving an inertial load, and said second actuator is an actuator for driving a normal load. The first function relating to the first actuator is:
The relationship between the second differential pressure and the value of the first control force is determined so that when the second differential pressure exceeds a predetermined value, an increase in the target value of the first differential pressure is suppressed. A hydraulic drive device characterized in that: 5. The hydraulic drive system for a construction machine according to claim 2, wherein said first and second actuators are actuators for traveling. The first and second functions are both equal to said first difference. A hydraulic drive device, wherein the relationship between the second differential pressure and the values of the first and second control forces is determined such that the target value of the pressure is greater than the second differential pressure. 6. The hydraulic drive device for a construction machine according to claim 2, wherein said first actuator is one of actuators for traveling, and said second actuator is an actuator for excavation work. Means gives a relatively large time delay to a change in the value of the first control force obtained from the first function, and changes the value of the second control force obtained from the second function. And a second calculating means for giving a relatively small time delay to the hydraulic drive. 7. A hydraulic drive system for a construction machine according to claim 2, wherein said first actuator is a hydraulic motor, and said second actuator is a hydraulic cylinder. And a third operation for obtaining a temperature correction coefficient from the pressure oil temperature detected by the third means and a third function set in advance. Means, and a fourth calculating means for calculating the value of the second control force obtained from the second function and the temperature correction coefficient to correct the value of the second control force. Features hydraulic drive. 8. A hydraulic drive device for a construction machine according to claim 1, wherein a selection command signal that is externally operated and that corresponds to the type of work or the content of the work performed by driving the first and second actuators is provided. There is further provided a fourth means for outputting, wherein the second means is preset in correspondence with the second differential pressure obtained by the first means and the first and second branch flow compensating valves, respectively. And a fifth operation means for obtaining third and fourth control force values from the fourth and fifth functions and the selection command signal output from the fourth means. apparatus. 9. The hydraulic drive system for a construction machine according to claim 8, wherein said fifth calculating means includes a plurality of functions having different characteristics as said fourth and fifth functions, respectively. Selects one of a plurality of functions in accordance with the selection command signal output from the second unit, and converts the second function from the second differential pressure obtained by the first means and the selected function to the second differential pressure. A hydraulic drive device for determining corresponding third and fourth control force values. 10. The hydraulic drive system for a construction machine according to claim 1, wherein said first actuator is an actuator for driving an inertial load, and said second actuator is an actuator for driving a normal load. Fifth means for detecting the discharge pressure of the pump is further provided, wherein the second means uses the second differential pressure obtained by the first means and a preset sixth function to calculate the second pressure. A sixth calculating means for obtaining a value of a fifth control force corresponding to the differential pressure and setting the value of the fifth control force to be a value of a control force to be applied by the driving means of the first branching compensation valve; A value of a sixth control force for holding the discharge pressure at a predetermined value is obtained from the detected discharge pressure and a preset seventh function, and the first control force is selected from the fifth control force and the sixth control force. The larger the target value of the differential pressure of the second Hydraulic drive system and having a seventh arithmetic means which drive means the value of the control force to be applied. 11. The hydraulic drive system for a construction machine according to claim 10, further comprising sixth means which is externally operated and outputs a selection command signal relating to a predetermined value of said discharge pressure.
The hydraulic drive device according to claim 7, wherein the seventh calculating means changes a characteristic of the seventh function in accordance with the selection command signal, thereby changing a predetermined value of the discharge pressure. 12. The hydraulic drive system for a construction machine according to claim 1, wherein said first actuator is an actuator for driving an inertial load, and said second actuator is an actuator for driving a normal load. 1
And a second means for setting a flow rate increasing rate of the pressure oil supplied through the first branch flow compensating valve. Calculates the value of the seventh control force corresponding to the second differential pressure from the second differential pressure obtained by the first means and an eighth function set in advance, Eighth arithmetic means for setting the value of the control force to be applied by the drive means of the shunt compensation valve, and the seventh control means when the start of driving of the first actuator is detected by the seventh means. A value of an eighth control force that changes at a speed equal to or less than a change amount corresponding to the flow rate increasing speed is determined using the value of the force as a target value, and the driving means of the first shunt compensation valve uses the eighth control force. Ninth calculating means for setting a value of a control force to be applied, to a hydraulic drive device. 13. The hydraulic drive system for a construction machine according to claim 12, further comprising: ninth means for detecting the drive of said second actuator, wherein said ninth arithmetic means comprises: A hydraulic drive device for a construction machine, wherein the value of the eighth control force is obtained when the start of driving of the first and second actuators is detected by the means of (9). 14. The hydraulic drive system for a construction machine according to claim 1, further comprising: tenth means for detecting a discharge pressure of said hydraulic pump, wherein said second means is determined by said first means. A tenth calculating means for calculating a differential pressure target discharge amount of the hydraulic pump that holds the differential pressure constant from the second differential pressure, and a discharge pressure detected by the tenth means and a preset hydraulic pump. Eleventh calculating means for calculating the input restriction target discharge amount of the hydraulic pump from the input restriction function, and twelfth calculation means for calculating a deviation between the differential pressure target discharge amount and the input restriction target discharge amount,
When the input restricted target discharge amount of the differential pressure target discharge amount and the input restricted target discharge amount is selected as the discharge amount target value of the hydraulic pump, the first and second values are determined based on the deviation of the target discharge amount. And a thirteenth calculating means for calculating an individual value as a value of the control force to be applied by each driving means of the branch flow compensating valve. 15. A hydraulic drive system for a construction machine according to claim 1, wherein said hydraulic drive system is provided in said first and second flow compensating valves.
Each of these flow compensating valves is urged in the valve opening direction, and further includes a driving means different from the driving means described above, and a pilot pressure supply means for introducing a substantially constant common pilot pressure to the other driving means. The hydraulic drive device according to claim 1, wherein the first-mentioned driving means is disposed on a side for urging the first and second branch flow compensating valves in a valve closing direction, respectively.