JPH02118203A - Hydraulic driving device - Google Patents

Abstract

PURPOSE:To improve operability and/or operation efficiency by controlling each of branch flow compensation valves according to differencial pressure between a discharge pressure of a hydraulic pump and a maximum load pressure of each actuator, and thereby giving independent pressure compensation characteristic to the brunch compensation valves. CONSTITUTION:In a controller 61, a control force value is computed according to a differencial pressure between a discharge pressure of a main pump 22 and a maximum load pressure of actuators 23 to 28 detected by a differencial pressure detector 59, and an oil temperature Th detected by a temperature detector 66, and corresponding electric signals (a) to (f) are output of solenoid proportional reducing valves 62a to 62f. Each control force is output through lines 51a to 51f to driving parts 35c to 40c of branch flow compensation valves 35 to 40. Independent pressure compensation characteristic is given to each of the branch flow compensation valves 35 to 40. Optimal branch flow ratio is obtained according to the kinds of the actuators 23 to 28 at the time of complex operation, so that operability and/or operation efficiency is improved.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to a hydraulic drive system for a submersible construction machine such as a hydraulic excavator, and in particular, the present invention relates to a hydraulic drive system for a submersible construction machine such as a hydraulic excavator, and particularly includes a flow compensation valve that controls a differential pressure across a flow control valve. A control force based on the differential pressure between the discharge pressure of the hydraulic pump controlled by load sensing and the maximum load pressure of the plurality of actuators is applied to each of these branch compensation valves, and the target value of the differential pressure across the flow control valve is applied. The present invention relates to a hydraulic drive device for setting.

[Conventional technology]

In recent years, in hydraulic drive systems for construction machinery such as hydraulic excavators and hydraulic cranes, which are equipped with multiple hydraulic actuators that drive multiple driven objects, the discharge pressure of the hydraulic pump is controlled in conjunction with the load pressure or required flow rate. At the same time, a pressure compensation valve is arranged in relation to the flow control valve, and the pressure difference between the front and rear of the flow control valve is controlled by this pressure compensation valve, thereby stably controlling the supply flow rate during combined drive. There is. this house,
Load sensing control is a typical example of controlling the discharge pressure of a hydraulic pump in conjunction with load pressure.

Load sensing control is to control the discharge amount of a hydraulic pump so that it is a certain value higher than the discharge pressure of the hydraulic pump or the maximum load pressure of multiple hydraulic actuators. Economical operation is possible by increasing or decreasing the discharge amount of the hydraulic pump.

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 when a plurality of actuators are driven in combination, a state in which the pump discharge amount is insufficient occurs. This is commonly known as hydraulic pump saturation. When 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, making it impossible to perform combined driving of multiple actuators.

In order to solve such problems, D E-A 1342
2165 (corresponding to JP-A No. 60-11706), in the hydraulic 15-section movement device described in JP-A No. 60-11706, each pressure compensating valve that controls the differential pressure across the flow rate control valve is provided with a spring that sets a target value for the differential pressure across the flow control valve. Two drive parts that act in the valve opening direction and the valve closing direction are provided, the discharge pressure of the hydraulic pump is guided to the drive part that acts in the valve opening direction, and the maximum load of the multiple actuators is guided to the drive part that acts in the valve closing direction. A control force based on the pressure difference between the pump discharge pressure and the maximum load pressure is applied in the valve opening direction, and this control force is used to determine the target value of the pressure difference before and after the pump. With this configuration, when saturation occurs in the hydraulic pump, the differential pressure between the pump discharge pressure and the maximum load pressure decreases, so the target value of the differential pressure across the flow control valve in each pressure compensation valve also decreases. As a result, the pressure compensation valve associated with the low-pressure side actuator is further throttled, and pressure oil from the hydraulic pump is prevented from preferentially flowing to the low-pressure side actuator. Thereby, the pressure oil from the hydraulic pump is divided in accordance with the ratio of the required flow rate (valve opening degree) of the flow rate control valve and is supplied to the plurality of actuators, making it possible to perform appropriate combined driving. In addition, in this configuration, the pressure compensation valve has the function of reliably dividing 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 book, this function is called ``shunt flow compensation'' for convenience, and the pressure compensation valve is called ``shunt flow compensation valve.''

[Problem to be solved by the invention]

By the way, in this conventional hydraulic drive device, each branch compensation valve uses the difference between the discharge pressure of the hydraulic pump controlled by load sensing and the maximum load pressure of the plurality of actuators as the target value of the differential pressure across the flow control valve. A control force is applied based on pressure. Therefore, if the pressure-receiving area of all drive parts is the same, the control force applied to each branch compensation valve will be the same, and the pressure compensation characteristics of all branch compensation valves will be the same. are the same.

For this reason, for example, when a composite operation is performed in which two or more actuators are driven simultaneously, the ratio of the distribution of the flow rate supplied to the actuators, that is, the flow rate control, is The problem is that the flow rate is uniquely determined depending on the opening ratio of the valve, and depending on the type of complex operation, the flow rate may be distributed too much or too little to one actuator, reducing operability and/or work efficiency. was there.

An object of the present invention is to provide a hydraulic drive device for construction machinery that can provide a flow-off compensation valve with a pressure compensation characteristic and improve operability and/or work efficiency.

[Failure to solve 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 a device constituted by the hydraulic pump;
between first and second flow control valves that respectively control the flow of pressure oil supplied to these first and second actuators, and the inlet and outlet of these first and second flow control valves; first and second flow compensating valves each controlling a first differential pressure generated, and responsive to a second differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second actuators. and a discharge amount control means for controlling the flow rate of pressure oil discharged from the hydraulic pump, and the first and second flow compensating valves each apply a control force based on the second differential pressure to a corresponding flow of the flow of the pressure oil. A hydraulic drive system for a construction machine having a drive means for applying a target value of the first differential pressure to a compensating valve is configured to control the discharge pressure of the hydraulic pump and the pressure of the first and second actuators. a first means for determining the second differential pressure from the maximum load pressure; and a first means for determining the second differential pressure based on at least the second differential pressure determined by the first means; a second means for calculating individual values of the control force to be applied by each of the driving means; and a first means provided corresponding to each of the first and second flow compensating valves.
and second control pressure generating means, each of which generates a control pressure according to the individual value determined by the second means, and applies this to the driving means of the first and second flow compensation valves. There is provided a hydraulic 1stK motion device characterized in that it has the first and second control pressure generating means that output respective outputs.

In one aspect of the present invention, the second means includes the first
The first and second functions corresponding to the second differential pressure are determined from the second differential pressure obtained by the method and the first and second functions preset corresponding to the first and second flow compensation valves. It may include a first calculation means for calculating the value of the second control force.

At this time, if 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 second differential pressure and the values of the first and second control forces are adjusted such that as the differential pressure between the two decreases, the target value of the first differential pressure decreases, and the rate of decrease is different between the two. The relationship is defined.

If the first actuator is an actuator that drives a =+I load, and the second actuator is an actuator that drives a normal load, preferably,
the first actuator related to at least the first actuator;
The function of the second differential pressure is such that when the second differential pressure increases beyond a predetermined value, an increase in the target value of the first differential pressure is suppressed.
A relationship between the differential pressure of and the value of the first control force is determined.

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

When the first actuator is one of the actuators for traveling and the second actuator is an actuator for excavation work, preferably, the second means includes a first actuator that is not determined from the first function. a second control force that provides a relatively large time delay for a change in the value of the first control force, and a second control force that provides a relatively small time delay for a change in the value of the second control force obtained from the second function; It further has calculation means.

When the first actuator is a hydraulic motor and the second actuator is a hydraulic shilling, the hydraulic drive device of the present invention preferably includes a third actuator that detects the temperature of the pressure oil discharged from the hydraulic pump. The second means further comprises a third calculation means for calculating a temperature correction coefficient from the temperature of the pressure oil not detected by the third means and a preset third function; The apparatus further includes a fourth calculation 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.

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 the work to be performed by driving the first and second actuators. further comprising a fourth means for determining the pressure difference, the second means being preset in accordance with the second differential pressure determined by the first means and the first and second flow compensating valves, respectively. It may also include a fifth calculation means for calculating the values of the third and fourth control forces from the fourth and fifth functions and the selection command signal output from the fourth means.

In this case, preferably, the fifth calculation means is provided with a plurality of functions having different characteristics as the fourth and fifth functions, and each of the fifth calculation means has a plurality of functions according to the selection command signal output from the fourth means. select one of the functions, and 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. 6 In a further aspect of the present invention, when the first actuator is an actuator that drives a 10-dimensional load and the second actuator is an actuator that drives a normal load, the hydraulic drive device of the present invention further comprises a fifth means for detecting the discharge pressure of the hydraulic pump, and the second means detects the second differential pressure obtained by the first means and a preset sixth function. A fifth control force value corresponding to the second differential pressure is determined, and this value is set as a control force value to be applied by the driving means of the first flow branch compensation valve.
calculates the value of a sixth control force for maintaining the discharge pressure at a predetermined value from the discharge pressure detected by the fifth means and a preset seventh function; and the sixth
and seventh calculation means for determining the control force that increases the target value of the first differential pressure as the value of the control force to be applied by the drive means of the second flow compensation valve. Good too.

In this case, the hydraulic drive device of the present invention further includes sixth means that is operated from the outside and outputs a selection command signal related to the predetermined value of the discharge pressure, and the seventh calculation means is configured to output the selection command signal related to the predetermined value of the discharge pressure. The characteristic of the seventh function may be changed by a signal, and the predetermined value of the discharge pressure may be changed.

Furthermore, in another aspect of the invention, when the first actuator is an actuator that drives an inertial load and the second actuator is an actuator that drives a normal load, the hydraulic drive device of the invention further comprising seventh means for detecting the driving of the first actuator, and eighth means for setting a rate of increase in the flow rate of the pressure oil supplied through the first flow division compensation valve, The means obtains a seventh control force value corresponding to the second pressure difference from the second pressure difference obtained by the first means and a preset eighth function, and calculates the value of the seventh control force corresponding to the second pressure difference obtained by the second means. eighth calculation means for determining the value of the control force to be applied to the driving means of the shunt compensation valve; Using the control force value as a target value, an eighth control force value that changes at a rate less than or equal to the amount of change corresponding to the flow rate increase rate is determined, and this eighth control force is applied to the driving means for the first flow division compensation valve. may have a ninth calculation means that determines the value of the control force to be applied.

In this case, the hydraulic drive device of the present invention further includes a ninth means for detecting the driving of the second actuator, and the ninth calculating means is configured to detect the driving of the first actuator by the seventh and ninth means. The value of the eighth control force may be determined when the start of driving the second actuator is detected.

In yet another aspect of the present invention, the hydraulic drive device of the present invention further includes a tenth means for detecting the discharge pressure of the hydraulic cylinder, and the second means detects the discharge pressure determined by the first means. a tenth calculating means for calculating a differential pressure target discharge amount of the hydraulic pump to maintain the second differential pressure constant; and input of the discharge pressure detected by the tenth means and a preset hydraulic pump. an eleventh calculation means for calculating an input limit target discharge amount of the hydraulic pump from a restriction function; a thirteenth calculation means for calculating a deviation between the differential pressure target discharge amount and the input limit target discharge amount; and the differential pressure target discharge amount. When the input limited target discharge amount is selected as the discharge amount target value of the hydraulic pump, based on the deviation of the target discharge I,
13. Calculating individual values as control force values to be applied by the driving means of each of the first and second flow branch compensation valves.
It may also have a calculation means.

In still yet another aspect of the present invention, preferably the hydraulic drive device of the present invention is provided in the first and second branch compensation valves, and is configured to first bias each of the branch compensation valves in the valve opening direction. further comprising drive means separate from the previously mentioned drive means and pilot pressure supply means for directing a substantially constant common pilot pressure to the further drive means, said first mentioned drive means each It is arranged on the side that urges the first and second branch compensating valves in the valve closing direction.

[Effect]

In the present invention configured in this way, the second means determines individual values as the value of the control force to be applied to each drive means of the first and second flow branch compensation valves based on the second differential pressure. is calculated, the first and second control pressure generation means generate control pressures according to these individual values, and output these to the drive means of the first and second flow compensation valves, respectively. As a result, individual pressure compensation characteristics are given to the first and second flow divider compensation valves, and an optimal flow divider ratio according to the type of actuator can be obtained during a combined operation in which the first and second actuators are simultaneously driven. It is possible to improve operability and/or work efficiency.

〔Example〕

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Preferred embodiments of the present invention as applied to a hydraulic excavator will be described below with reference to the drawings.

11 Blood Dog Dwarf 1 First, a first embodiment of the present invention will be described with reference to FIGS. 1 to 3. FIG.

In FIG. 1, the hydraulic drive system applied to the hydraulic excavator of this embodiment includes a prime mover 21, one variable displacement hydraulic pump driven by the prime mover 21, that is, a main pump 22, and a discharge pump from the main pump 22. A plurality of actuators are driven by pressurized oil, namely, a swing motor 23, a left travel motor 24, a right travel motor 25, and a boom cylinder 2.
6. Arm cylinder 27, packet cylinder 28, and flow control valves that control the flow of pressure oil supplied to each of these plurality of actuators, that is, directional switching valve 29 for turning, directional switching valve 30 for left travel, right Directional switching valve for travel 31, directional switching valve for boom 32, directional switching valve for arm 3
3. The packet directional switching valve 34 and the pressure difference generated between the inlet and outlet of the flow control valve, which is arranged upstream of the flow control valve, that is, the pressure difference ΔPv1 before and after the flow control valve.
．． ΔPv2゜ΔPv3. ΔPv4. ΔPV5. ΔPv6
pressure compensation valves, i.e., branch compensation valves 35.
36, 37.38, 39.40.

In addition, the hydraulic drive system of this embodiment has the discharge pressure ps of the main pump 22 and the maximum load pressure P of the seven cutout units 23 to 28 within a range until the main pump 22 reaches the maximum possible discharge amount.
A discharge amount control device 41 of a load sensing control type is provided, which controls the discharge amount of the main pump 22 in response to the differential pressure ΔPLS with respect to alaX so that the discharge pressure Ps is higher than the differential pressure ΔPLS by a certain value. There is.

The flow rate control valves 29-34 include check valves 42a, 42b, 42c, 42d, and 42e for taking out the load pressure of the actuators 23-28 when they are driven, respectively.
, 42f, load lines 43a, 43b, 43c,
43d, 43e, and 43f are connected, and these load lines 43a to 43f are further connected to a common maximum load line 44.

The branch compensation valves 35 to 40 are each configured as follows. The branch flow compensation valve 35 is guided by the outlet pressure of the two swing direction switching valves, and is connected to a drive unit 35a that biases the valve element of the branch flow compensation valve 35 in the valve opening direction, and an inlet pressure of the swing direction switch valve 29. The driving part 35b biases the valve body of the branch flow compensation valve 35 in the valve closing direction, the spring 45 biases the valve body of the branch flow compensation valve 35 in the valve opening direction with force f, and the pilot line 51a. A control pressure PCI, which will be described later, is derived from the flow branch compensation valve 35.
a drive unit 35 that urges the valve body in the valve closing direction with a control force Fcl;
c, and the drive parts 35a and 35b actuate the shunt compensation valve 3.
A first control force based on the differential pressure ΔPV1 across the swing direction switching valve 29 is applied to the valve element 5 in the valve closing direction. A control force f-Fcl is applied in the valve opening direction, and the balance between the first control force and the second control force determines the throttling amount of the branch compensation valve 35, and the differential pressure ΔPvt across the swing direction switching valve 23 is controlled. be done. Here, the second control force f-Fc1 is a value that sets the target value of the differential pressure ΔPv1 across the turning direction switching valve 23.

The other branch compensating valves 36 to 40 are similarly configured. That is, the shunt compensation valves 36 to 40 have opposing drive units 36 that bias their valve bodies with first control forces based on the differential pressures ΔPv2 to ΔPv6 across the flow control valves 30 to 34, respectively.
a, 36b; 37a, 37b: 38a738b;
39a, 39b; 40a, 40b and springs 46, 47, 58, 59, 50 that bias the valve body in the valve opening direction with force
and pilot lines 51b, 51c, 51d, 51e
, 51f to control pressure Pc2., which will be described later. P
c3. Pc4゜Pc5. Pc6 is guided to apply a control force Fc2. Fc3. Fc4. Fc5
．． Drive units 36c and 37c that bias in the valve closing direction with Fc6
, 38c, 39c, 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 and controls the displacement, and a control valve 53 that controls the displacement of the hydraulic cylinder device 52. Main pump 22 discharge pressure Ps
and the maximum load pressure P ama of actuators 23 to 28
The spring 54 that sets the differential pressure ΔPLS with x and the maximum load pressure P amax of the actuators 23 to 28 are
and a drive section 58 through which the discharge pressure ps of the main pump 22 is guided through a conduit 57.When the large load pressure p amax rises, the control is performed in response to The valve 53 is driven to the left in the figure, driving the hydraulic cylinder device 52 to the left in the figure, increasing the displacement of the main pump 22 and increasing the discharge amount. As a result, the discharge pressure Ps of the main pump 22 is a constant li! determined by the spring 54. ! held at high pressure.

The hydraulic drive system of this embodiment further includes a main pump 2.
A differential pressure detector 5 introduces the discharge pressure Ps of No. 2 and the maximum load pressure PalaX of the actuators 23 to 28, detects the differential pressure ΔPLS between the two, and outputs the corresponding electric signal X1.
9, and a temperature detector 60 that detects the temperature Th of the pressure oil discharged from the main pump 22 and outputs a corresponding electric signal X2.
and electric signals xi and x2 from the differential pressure detector 60 and temperature detector 61, and calculate the values of the control forces Fc1 to Fc6 described above based on the detected differential pressure ΔP[S and oil temperature Th, A controller 61 that outputs corresponding electric signals a, b, c, d, e, f, and a controller 61 that is provided corresponding to the shunt compensation valves 35 to 40, and an electric signal a from the controller 61
, b, c, d, e, f, respectively.
It is equipped with a pilot pump 63 that supplies pilot pressure to the electromagnetic proportional pressure reducing valves 62a to 62f, and a control pressure generation circuit 65 that includes a relief valve 64 that defines the magnitude of the pilot pressure output from the pilot pump 63. .

The electromagnetic ratio S pressure reducing valves 62a to 62f are operated by electric signals a to f, and control forces Fcl to F calculated by the controller 61 are operated.
Control pressures Pc1 to PC6 are generated according to the value of C6, and outputted to the drive units 35c to 40c of the branch compensation valves 35 to 40, respectively, via pilot lines 51a to 51f.

trii proportional pressure reducing valves 62a to 62f and relief valve 64
are preferably configured in one block as a collection, as shown by a two-dot chain line 66.

The controller 61, as shown in FIG.
1, by using the input section 7o for inputting
A calculation unit 72 that performs calculations to determine the value of , and an output unit 73 that outputs the control force values determined by the calculation unit 72 as electrical signals a to f are provided.

The contents of the calculations performed by the calculation unit 72 of the controller 61 are shown in block 80 to 0 in FIG. 3 as a functional block diagram.
Reference numeral 85 denotes a function block that is provided corresponding to the shunt compensation valves 35 to 40 and stores in advance function data including the functional relationship between the differential pressure ΔP[S and the control forces Fc1 to FC6. The control force values Fc1 to Fc6 corresponding to the differential pressure ΔPLS based on the electrical signal X1 are determined. The block 86 is a function programmer for temperature correction that stores in advance numerical data including the functional relationship between the oil temperature Th and the correction coefficient.
From this function block 86, a correction coefficient/factor K corresponding to the oil temperature Th based on the electrical signal X2 is determined. The correction coefficient obtained in the function block 86 is applied to the multiplication block 87.
．． At step 88.89, it is multiplied by the values of control forces Fc4 to Fc6 determined at function block 83.84.85, 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 multiplication block 87.888
Temperature corrected at 9 and letter system fg [Fc4. Fc5.
Fc6 is output as electrical signals a to f after being filtered by first-order delay elements in delay blocks 90 to 95, respectively.

The functional relationship between the differential pressure ΔPLS stored in the function blocks 80 to 85 and the control forces Fcl to FC6 is shown in FIGS. 4A to 4D and FIG. It shows the functional relationship with the value of the control force Pc1 to be applied to the compensation valve 35. Here, ΔPLSO is the differential pressure between the discharge pressure of the main pump 22 and the maximum load pressure maintained by the discharge amount control device 41 of the load sensing control method,
That is, it is the load sensing compensation differential pressure set by the spring 54 of the control valve 53, and fO is the value of the control force Fc1 corresponding to the load sensing compensation differential pressure ΔPLSO. A is the minimum differential pressure that determines the maximum speed of the swing motor 23, that is, the maximum flow rate compensation differential pressure related to the swing motor 23, and fc is the maximum flow rate compensation control force corresponding to this maximum flow rate compensation differential pressure A. f is the force of the spring 45.

Note that f-fo is the load sensing and sinking compensation differential pressure Δp t
This value corresponds to the second control force applied to the shunt compensating valve 35 when so is secured, and this value is based on the target value of the differential pressure ΔPv1 across the turning direction switching valve 23 set thereby. It is determined to substantially match the load sensing compensation differential pressure Δp tso.

In FIG. 4A, the two-dot chain line indicates that when the differential pressure ΔP[S is zero, a control force equal to the force f of the spring 45 is applied, and the differential pressure Δ
It shows the characteristics of a basic function that gradually decreases the control force as PLS increases, and the differential pressure ΔPLS and control force F
The functional relationship of cl is the differential pressure ΔP[S or the maximum flow rate compensation differential pressure A
If smaller, the differential pressure ΔPL follows the characteristics of the basic function.
The value of the control force Fc1 gradually decreases as S increases, and when the differential pressure ΔP [S becomes equal to or higher than the maximum flow compensation differential pressure A, the differential pressure Δ
The relationship is such that a constant control force fc is output regardless of the C system as PLS increases. Further, when the differential pressure ΔP[S becomes equal to or less than the minimum flow compensation differential pressure B, the relationship is such that the differential pressure ΔP[S is limited to the maximum value f IaX less than the force f of the spring 45, regardless of the decrease in the differential pressure ΔPLS.

FIG. 4B shows the differential pressure ΔPLS and the control force Fc to be applied to the shunt compensation valves 36 and 37 related to the travel motors 24 and 25.
2. It shows the functional relationship with the value of Fc3. Here, the two-point lid line shows the characteristics of the basic function as in FIG. 4A, and the differential pressure ΔPLS and control force Fc2. The functional relationship of Fc3 is such that the value of the control force Fc2°Fc3 gradually decreases as the differential pressure ΔP[S increases, with a slope smaller than the slope of the basic function, and the correction flow rate decreases compared to when controlled by the basic function. The relationship is such that ΔQ can be obtained.

FIG. 4C shows the functional relationship between the differential pressure ΔP[S and the value of the control force Fc4 to be applied to the branch compensation valve 38 related to the boom cylinder 26. The functional relationship is that the control force Fc2. Compared to the slope of the functional relationship of Fc3, the value of the control force Fc4 gradually decreases as the differential pressure ΔPLS increases, with a slope that is even smaller than the slope of the basic function.

FIG. 4D shows the differential pressure ΔP [S and the control force Fc5. It shows the functional relationship with the value of Fc6. The functional relationship generally shows that the control force Fc5. When the value of Fc6 decreases and the differential pressure ΔPLS becomes less than the minimum flow compensation differential pressure B, the spring 49.5 will increase regardless of the decrease in the differential pressure ΔPLS, similar to the functional relationship shown in FIG. 4A.
The relationship is such that the maximum value f laX is less than or equal to the force f of 0.

FIG. 5 shows the above-mentioned functions together in order to make the mutual relationship easier to understand.

FIG. 6 shows the functional relationship between the oil temperature Th and the correction coefficient stored in the function block 86.
is larger than the predetermined temperature ThO, the correction coefficient is 1, and as the oil temperature Th falls below the predetermined temperature ThO, the correction coefficient K gradually becomes smaller than 1. Here, the predetermined temperature ThO is a temperature at which the pressure oil flowing through the circuit is considered to have a viscosity that does not significantly affect the flow rate discharged from the main pump 22.

In the delay element blocks 90 to 95, time constants T1 to 1'6 are set for each of the actuators 23 to 28 to give a time delay of ifl to their operation. this house,
Shunt compensation valve related to 24.25 for traveling motor 36.37
The time constants 72 and T3 of the blocks 91 and 92 corresponding to the blocks 91 and 92 are made extremely large compared to the other time constants TI and 74 to T6, and the control force F to be applied to the shunt compensation valves 36 and 37 is
c2. A large time delay is given to changes in the value of Fc3.

The structure of the working members of the hydraulic excavator driven by the hydraulic drive system of this embodiment is shown in FIGS. 7 and 8. The swing motor 23 drives the swing structure 100, the left travel motor 24,
The right traveling motor 25 drives crawler tracks, that is, traveling bodies 101 and 102, and the boom cylinder 26, arm cylinder 27, and packet cylinder 28 drive the boom 103 and arm 10, respectively.
4. Drive packet 105.

Next, the operation of this embodiment configured as above will be explained.

Actuation of any one or more of the flow control valves 29-34 supplies pressurized oil from the main pump 22 through the shunt compensation valve and the flow control valve to the corresponding actuator. At this time, the main pump 22 is load sensing controlled by the discharge amount control device 41, and the differential pressure ΔP[S between the discharge pressure of the main pump 22 and the maximum load pressure is detected by the differential pressure detector 59,
A corresponding electrical signal X1 is input to the controller 21. At the same time, the oil temperature is detected by the oil temperature detector 60 and a corresponding electrical signal X2 is input to the controller 61.

The calculation unit 72 of the controller 61 calculates the values of the control forces Fc1 to Fc6 as described above, and electrical signals a to f corresponding to these control forces are sent to the electromagnetic proportional pressure reducing valves 62a to 6.
2f, the electromagnetic proportional pressure reducing valves 62a to 62f are driven, and the control pressure Pc1- corresponding to the control forces FC1 to FC6 is applied.
Pc6 is the drive unit 35c to 40c of the branch compensation valves 35 to 40
guided by. Therefore, the control forces FC1 to FCG in the valve closing direction are applied to the branch compensating valves 35 to 40 by the drive units 35c to 40c.
is given, and as a result, the second flow compensation valves 35 to 40 are
Control force f-Fc1. ．． f=Fc2. ,'f-Fc
3. f-Fc4. f=FC5·, 'f-Fc6 are applied in the valve opening direction. That is, if at least one of the flow rate control valves 29 to 34 is operated, all of the branch compensation valves 35
A control force Fc: 1 to FC6 is constantly applied to the flow rate control valve 40, and at this time, the flow rate control valve and the branch flow compensation valve whose control valve is not operated are set to the first control force based on the differential pressure across the flow rate control valve. Since no control force is applied, it remains at the fully open position.

Next; on the premise that the oil temperature is at least 100 cm as shown in Figure 6, if the revolving body 100, traveling bodies 101, 102, boom 103, arm 104, and packet 10"5 are operated individually, and The operation of the flow rate compensation valves 35 to 40 and the associated operation of the actuators 23 to 28 will be explained for each case where a combined operation is performed.One of the flow rate control valves 29 to 34 is operated,
01° 102, when the boom 103, arm 104, and packet 105 are operated independently, the first control force based on the differential pressure across the flow control valve is applied to the branch flow compensation valve related to the corresponding flow control valve in the valve closing direction. f1 is given.

The differential pressure across the flow rate control valve is the differential pressure ΔPLS between the discharge pressure of the main pump 22 controlled by load sensing and the maximum load pressure.
In the case of single operation, the differential pressure ΔP is generally
[S is maintained at the load sensing compensation differential pressure Δp tso or a value close to it.

At this time, the operated flow control valve is connected to the swing motor 23, arm 27. If one of the packets 28 is connected to the f system, the drive section 35c, 39c or 4 of the shunt compensation valve 35, 39 or 40
Control force Fcl applied to 0c.

Fc5 or Fc6 is determined from the functional relationship shown in FIG. 4A or 4D, where the load sensing compensation differential pressure Δ
The control force corresponding to p tso is fO.

Therefore, for example, f-fo is applied to the branch compensating valve 35 as the second control force. As described above, f-fo is a value that controls the front and rear differential pressure ΔPv1 of the turning directional control valve 23 to almost match the load sensing compensation differential pressure Δptso. Therefore, the second control force f-fo is always approximately equal to or greater than the first control force, and as a result, the branch compensating valve 35 also remains in the fully open position.

If the operated flow control valve relates to one of the travel motor 24.25 and the boom cylinder 26, the shunt compensation valve 36.3
7 or 38, the control force Fc2. applied to the drive unit 36c, 37c or 38c. Fc3 or Fc4 is obtained from the functional relationship shown in FIG. 4B or 4C, where the control force corresponding to the load sensing compensation differential pressure ΔPLSO is fO
It is a smaller value.

Therefore, for example, a force larger than f-fo is applied to the branch compensating valve 38 as the second control force.

Therefore, in this case as well, the second control force is greater than the first control force, and the branch compensation valve 38 is maintained at the fully open position.

In this way, when one of the flow control valves 29 to 34 is operated independently, the corresponding branch flow compensation valve basically does not operate, and the differential pressure across the flow control valve is mainly caused by the main pump 22 being loaded. It is controlled by sensing control, and a flow rate corresponding to the opening degree of the flow rate control valve is supplied to the actuator.

Next, by operating any two or more of the flow control valves 29 to 34, the rotating body 100, the traveling bodies 101, 102, and the boom 10
3. A case in which the actuators of the arm 104 and the packet 105 are operated in a combined manner will be described.

By simultaneously operating the flow control valves 29 and 32, the rotating body 100
and the combined operation 21 of the boom 103. For example, when performing a combined operation of swinging and raising the boom, the pressure oil from the main pump 22 passes through the branch compensation valve 35, 38 and the flow control valve 29, 32 to the swing motor 23 and the boom cylinder 26. supplied to

At this time, the differential pressure ΔP[S is normally less than or equal to the maximum flow compensation differential pressure A for the swing motor 23, and the control force Fc1 applied to the drive portion 35c of the branch compensation valve 35 is
A value in accordance with the characteristics of the basic frequency is calculated from the functional relationship shown in the figure, and a control force F is applied to the drive section 38c of the shunt compensation valve 38.
c4 is the control force Fc from the functional relationship shown in Fig. 4C.
A value smaller than l is calculated.

Therefore, the second control forces f-FC1 and f-Fc4 in the valve opening direction applied to the branch compensation valves 35 and 38 are such that fFcl<
The relationship is f-Fc4. That is, the control force f-FC4 of the branch flow compensation valve 38 in the valve opening direction becomes larger than the control force f-FC1 of the branch flow compensation valve 35 in the valve opening direction. As a result, at the start of the combined operation of swinging and boom raising, the degree to which the control force f-FC4 restricts the flow dividing compensating valve 38 associated with the boom cylinder 3, which is on the low load pressure side, becomes smaller, and the dividing flow compensating valve 38 Control force f-Fc1 same as compensation valve 35
The pressure difference across the flow control valve 32 is controlled to be greater than the pressure difference across the flow control valve 29, and the boom cylinder 26 is connected to the main pump 22. A flow rate larger than the flow rate obtained by distributing the discharge amount by the opening ratio of the flow rate control valves 29 and 32 is supplied, while a flow rate lower than the same flow rate is supplied to the swing motor 23, and as a result, the swing and boom raising The combined operation can be performed reliably, the boom raising speed is fast, and the turning is relatively smooth.

In this way, the swing motor 23 and the boom cylinder 2
If the flow control valve 32 is not returned to the neutral position in order to stop the boom cylinder from a state in which the boom cylinder is operated in combination with
As a result, the pump pressure increases temporarily, and the differential pressure ΔP[S becomes larger than the maximum flow rate compensation differential pressure A, which is the limit differential pressure during normal combined operation. In the calculating section 72 of the swivel controller 61, the value of the control force Fc4 is constant regardless of the increase in the differential pressure ΔPLS, as shown in FIG. 4A.
That is, the maximum flow rate compensation control force fc is determined.

Therefore, the second control force in the valve opening direction applied to the shunt compensation valve 35 related to the swing motor 23 is constant f-Fc, and the shunt compensation valve 35 opens proportionally as the differential pressure ΔPLS increases. If you try to do so, you will be regulated to ensure that you do not open yourself up too much.

As a result of being controlled in this manner, even if the flow rate control valve 26 is operated in the neutral direction to stop the boom cylinder 26 during a combination of swinging and boom raising, the shunt compensation valve 35 compensates for the maximum flow rate difference as described above. Since it is regulated so as not to open too much according to the maximum flow rate compensation control force fc corresponding to the pressure A, a flow rate with relatively little change is supplied to the swing motor 23 compared to the flow rate that was previously supplied to the swing motor 23. is,
Therefore, an increase in speed of the swing motor 23 that is not intended by the operator can be prevented, and excellent operability and safety can be obtained.

When operating the flow control valves 30.31 with the same stroke to perform straight travel, the pressure oil from the main pump 22 passes through the branch compensation valve 36.37 and the flow control valve 30.31 to the left and right travel motors 24.25. supplied to At this time, as the control forces FC2 and Fc3 applied to the drive parts 36c and 37c of the shunt compensation valves 36 and 37, values smaller than the control force obtained from the characteristics of the basic function are calculated from the functional relationship shown in FIG. 4B. be done. Therefore, the second control force f-Fc2. in the valve opening direction is applied to the branch compensating valves 36, 37. fFC3 is f-, where Far is the control force obtained from one basic function.
Fc2>f-Fcr-f-Fc3>f-Fcr.

Here, the second control force f-FC based on the basic function is
This value is set so that the target value of the differential pressure before and after the flow control valve is equal to the differential pressure ΔPLS.

Therefore, the flow rate control valves 30, 31
Compared to the normal case where the differential pressure across the flow rate control valve 30 and 31 is controlled to be almost equal to the differential pressure ΔPLS, the second control force is applied in the valve opening direction, which is larger, and the differential pressure across the flow rate control valve 30 and 31 becomes equal to the differential pressure ΔPLS. Pressure ΔP[Fc2-Fcr or Fc3-F further than S
If a differential pressure occurs between the load pressures of the travel motors 24 and 25, which pressure is not throttled until the pressure increases by a predetermined value ΔPo corresponding to cr, any pressure compensation is performed within the range where the differential pressure is smaller than the predetermined value ΔPO. The valve is not throttled and the travel motor 24
, 25 are in the same state as if they were connected in parallel.

Furthermore, even if the differential pressure exceeds the predetermined value ΔPO, the branch compensating valve on the low load pressure side is opened more widely than usual, so it can be assumed that the travel motors 24 and 25 are partially connected in parallel. be able to.

As a result of the shunt compensating valve functioning in this way, the resistance received by the left and right tracks is different during straight traveling, and the running motors 24, 25
Even if there is a difference in the load pressure of the travel motors 24, 2
5 are in the same state as if they were at least partially connected in parallel, so as in the case of a general hydraulic circuit in which the left and right traverse motors are connected in parallel, the left and right traverse motors are controlled by the straight-line maintaining force of the tracks themselves. Forcibly equalize the flow rates of pressure oil supplied to 24 and 25 to drive straight ahead! I can use it as a pillow. This lick, manual if by the operator
It is possible to reduce the labor required for one adjustment and reduce the operator's sense of fatigue.

In addition, since the functions of the shunt compensation valves 24 and 25 are partially disabled in this way, and the crawler track itself is forced to travel straight due to its own straight-line maintaining force, the flow rate control valves 30, 31 and the shunt compensation valves 36, 37 Even if there are variations in the performance of hydraulic equipment such as those due to manufacturing errors, it is possible to perform the intended straight-line travel, and furthermore, it is possible to continue straight-line travel even if there is a slight change in the control lever position. Similarly, the effort required for manual adjustment by the operator can be reduced, and the operator's sense of fatigue can be reduced.

Next, operate the flow rate control valves 30, 31 and drive the travel motor 24.
．． Let us consider a case in which the flow rate control valve 32 is further operated in a state in which a travel operation is performed by driving the pump 25, and a transition is made to a combined operation of traveling and raising the boom.

While only running operation is being performed, the flow control valve 3 is
2, the pressure oil from the main pump 22, which was previously supplied only to the left and right travel motors 24 and 25, will now be
The flow is supplied to the boom cylinder 26 through the flow compensation valve 38 and the flow control valve 32.

By the way, in the case of a combined operation of traveling and raising the boom, the boom cylinder 26 is normally on the high load pressure side. Therefore, at the moment when the state of only traveling operation is shifted to the combined operation of traveling and raising the boom, a situation occurs in which the differential pressure ΔPLS decreases extremely, and the calculation unit 7 of the controller 61
2, the control force Fc2.2 obtained from the functional relationship shown in FIG. The value of Fc3 also increases instantaneously and greatly. Therefore, this control force F c2. If Fc3 is directly output from the output section 73 as an electric signal S,c, the second control force f-Fc2. Correspondingly, f-Fc3 decreases rapidly. That is, at the initial stage of transition from a travel-only operation to a combined travel and boom-raising operation, a phenomenon occurs in which the shunt compensation valves 36 and 37 are momentarily extremely closed and then begin to open again, causing the diagonal travel motor 2
Fluctuations in the flow rate of the pressure oil supplied to the hydraulic excavators 4 and 25 become large, which causes extreme fluctuations in traveling speed, causing a large shock to the body of the hydraulic excavator and reducing operability.

On the other hand, in this embodiment, as described above, the delay element blocks 90 to 95 shown in FIG. is made extremely large compared to the other time constants T1, 74 to T6, and the control force Fc2. A large time delay is given to a change in the value of Fc3. Therefore, as described above, the control force Fc2. Fc3
Even if the value of Fc2. changes suddenly, the change is softened in blocks 91 and 92, and the control force Fc2. Changes in Fc3 also become gradual.

Therefore, sudden closing of the branch compensating valves 36 and 37 is avoided, the above-mentioned fluctuation in traveling speed is reduced, no large shock is caused to the body of the hydraulic excavator, and excellent operability is obtained.

Furthermore, while operating at least one of the flow control valves 29, 33, and 34 and driving the corresponding one of the swing motor 23, arm cylinder 27, and packet cylinder 28,
When the differential pressure ΔP[S momentarily becomes zero for some reason, such as when another actuator with a higher load pressure is further driven, the difference between the swing motor 23, arm cylinder 27, and packet cylinder 28 Since the functional relationship between the pressure and the control force has the same slope as the basic function as shown in FIGS. 4A and 4D, if the functional relationship is completely matched to the basic function, the control force Fc1. Fc5
．． The value of Fc6 becomes equal to the force f of the spring 45, 49, 50, and a phenomenon occurs in which the shunt compensating valves 35, 39, 40 are completely closed. When the shunt compensation valve is completely closed, the flow rate of the pressure oil supplied to the actuators 23, 27 and 28 becomes zero, and the rotating body ioo, the arm 104, and the packet 1
05, a large shock will occur, which will not only significantly deteriorate operability, but also cause damage to hydraulic equipment.

In this embodiment, in response to such a decrease in the differential pressure ΔPLS, when the differential pressure ΔPLS becomes equal to or less than the minimum flow compensation differential pressure B, the control force Fc1. Fc5. The relationship is such that Fc6 is limited to the maximum value flaX which is less than the force f of the spring 45, regardless of the decrease in the differential pressure ΔPLS. For this reason, the shunt compensation valve 35
．． 39.40 is prevented from completely closing,
It is possible to reduce shock, improve operability, and prevent damage to hydraulic equipment.

Next, the operations of the branch compensating valves 35 to 40 and the accompanying operations of the actuators 23 to 28 will be described in the case where the oil temperature changes below Tho shown in FIG.

In the calculation section 72 of the controller 61, as shown in FIG. The positive coefficient K is the multiplication block 87-8
9 to correct the temperature of the control forces FC4 to FC6. As shown in FIG. 6, the correction coefficient is approximately l when the oil temperature Th is higher than the predetermined temperature Tho;
When h is lower than the predetermined temperature Tho, it gradually becomes smaller than 1 as it becomes lower. This means that in a normal work environment such as during the daytime, the oil temperature Th is equal to the predetermined temperature Th.
o or more, since K=1, the values of the control forces Fc4 to FCG obtained in function blocks 83 to 85 are converted as they are into electrical signals S, e, and f, and
40 is driven according to the control forces FC4 to FC6.

With this, for example, when operating the flow rate control valves 38 and 39 to perform combined operation of the boom 103 and arm 104,
There are no problems with the boom cylinder 26 and the arm cylinder 27, that is, because the oil temperature Th is relatively high, the viscosity of the oil temperature is low and no large flow resistance occurs, and the flow control valves 38 and 39 and flow rate control are not affected. The main pump 2 is connected to the boom cylinder 26 and the arm cylinder 27 via valves 32, 33.
Pressure oil is supplied from 2, and the arm and packet can be driven in combination without reducing the operating speed of these actuators.

In addition, if the oil temperature T is lower than the predetermined temperature ThO due to work in a cold region or work environment such as early morning or night in winter, since K<1, multiplication blocks 87 to
8, the control force Fc4 multiplied by the correction coefficient I (
The value of Fc6 becomes smaller than the values calculated in function blocks 83 to 85, and the degree of this becomes larger as the oil temperature Th becomes lower. As a result, control forces Fc4 to Fc6 smaller than normal are applied from the driving parts 38c to 40c of the branch flow compensation valves 38 to 40 in response to a decrease in the oil temperature Tll, and the control forces Fc4 to Fc6 are applied to the branch flow compensation valves 38 to 40 to open the valves. The second control forces f-FC4, f-Fc5, and f-Fc6 in the directions become larger than normal as the oil temperature Th decreases.

That is, for example, by operating the flow control valves 38 and 39, the boom 1
When performing combined operation of 03 and arm 104, the oil temperature T
The flow rate that is almost the same as the flow rate when h is high is the flow rate at the branch compensating valve 38.
．． 39 and flow control valve 3233 to the boom cylinder 26 and arm cylinder 27, thereby,
Although the viscosity of the pressure oil increases due to the decrease in oil temperature Th and the flow resistance increases, the desired flow rate required by the flow rate control valves 32 and 33 can be supplied to the boom cylinder 26 and arm cylinder 27, and the flow rate of these actuators can be increased. Complex operations can be performed without reducing operating speed.

The same holds true when performing a combined operation of other combinations of the boom 103, arm 104, and packet 105, or a single operation of one of these.

In this way, the boom cylinder 26, the arm cylinder 27
and a shunt compensation valve 38~ corresponding to the packet cylinder 28.
40, the control force Fc4 is adjusted according to the change in oil temperature Th.
~ By correcting the value of Fc6 and adjusting the pressure compensation characteristics, the operating speed of these actuators can be kept constant regardless of changes in oil temperature, allowing stable single operation or combined operation. .

On the other hand, the control forces FC1 to F obtained by the function blocks 80 to 82 corresponding to the swing motor 23 and the traveling motor 24.25
C3 is output as electric signals a to C through delay element blocks 90 to 92 without being subjected to oil temperature correction. Therefore, when the oil temperature is below the predetermined temperature ThO, the viscosity of the pressure oil increases, flow resistance increases, and the flow rate supplied to the boom cylinder 26 and arm cylinder 27 decreases. Therefore, unlike the boom cylinder 26, arm cylinder 27, and packet cylinder 28, which are cylinder-based actuators, the swing motor 23 and travel motors 24, 25, which are motor-based actuators, are driven by pressure oil passing through the inside. If high viscosity pressure oil is supplied at the same flow rate as normal low viscosity oil, there is a risk of damaging internal parts, but since the flow rate is reduced, such damage will not occur. .

As explained above, according to this embodiment, the differential pressure ΔP
The drive unit 35c of the branch compensation valves 35 to 40 based on LS
The values of the control forces Fc1 to Fc6 to be applied through the control forces Fc1 to Fc6 are individually calculated, and the control pressures PC1 corresponding to these control forces are calculated from the electromagnetic proportional pressure reducing valves 62a to 62f provided corresponding to the branch compensation valves 35 to 40. ~Pc6 are individually generated and guided to the relevant drive units 35c to 40c, so the shunt compensation valves 35 to 40 have associated actuators 23 to 40.
28 can be provided with individual pressure compensation characteristics suitable for
When performing a combined operation of the driven bodies 100 to 105, it is possible to obtain an optimum flow division ratio according to the type of driven bodies, and it is possible to improve operability and work efficiency.

Also, control force Fc is provided corresponding to the actuators 23 to 28.
The values of 1 to Fc6 are individually calculated, and the electromagnetic proportional pressure reducing valve 62a
Since the corresponding control pressures Pc1 to Pc6 are generated individually from ~62f, it is possible to modify the values of the control forces FC1 to Fc6 individually. Optimal time constant T1
~Function block 8 for giving T6 individually and correcting oil temperature
6, and only the control forces Fc4 to Fc6 are corrected by the correction coefficient I. It is also possible to further differentiate the operating characteristics of the shunt compensating valve by considering various conditions. When performing the combined operations of 23 to 28,
Furthermore, operability and work efficiency can be improved.

In addition, in the above embodiment, the function blocks 80 to 85
The shape of the function between the differential pressure ΔPLS and the control forces Fc1 to FC6 stored in can be modified in various ways.

For example, in the function block 80 related to the swing motor 23, when the differential pressure ΔPLS temporarily increases and becomes larger than the maximum flow rate compensation differential pressure A, as shown in FIG. Although the functional relationship was determined so as to obtain the flow rate compensation control force fc, the functional relationship may be determined in other forms.9For example, as shown in Fig. 9, the flow characteristics of pressure oil, the temperature of pressure oil In consideration of the above, as the differential pressure ΔP[S becomes larger than the maximum flow rate compensation differential pressure A, the functional relationship that outputs a control force that increases proportionally starting from the maximum flow rate compensation control force fc, or the 10th As shown in the figure, as the differential pressure ΔPLS becomes larger than the maximum flow compensation differential pressure A, there is a functional relationship that outputs a control force that increases step by step, and as shown in FIG. It can be set to a functional relationship that increases in a curve as it becomes larger than the compensation differential pressure A, and as shown in FIG. It is possible to set a functional relationship that outputs a control force that decreases proportionally with a small gradient.

Further, in the above embodiment, when the differential pressure Δ·PLS;/+<M becomes larger than the large flow rate compensation differential pressure A, a constant control force fc
Although the functional relationship is set so as to obtain the following, the relationship between the differential pressure ΔPLS and the control force stop function can be similarly set as appropriate for the branch compensation valves related to other actuators.

Further, a function block 81 related to the travel motors 24 and 25
．． 82, as shown in FIG. 4B, the differential pressure ΔP[
The functional relationship was determined so that as S increases, the difference in the control force for the characteristics of the basic function becomes smaller.
As shown in Figure 3, a functional relationship in which the difference in control force with respect to the characteristics of the basic function is constant regardless of changes in the differential pressure ΔPLS,
Alternatively, a similar effect can be obtained by adopting a functional relationship in which the difference in control force with respect to the characteristic of the basic function increases as the differential pressure ΔP[S increases.

(8-F-#<v) Second Embodiment The second embodiment of the present invention will be explained with reference to FIGS. 15 and 16. In FIG. The same reference numerals are given to the members.

In FIG. 15, the two turning directional control valves and the boom directional control valve 32 detect these operations and send an electric signal X.
A manipulation detector 110,111 is provided which outputs 3 and X4. In addition, the same reference pilot pressure P' is guided to the branch flow compensation valves 35A to 40A through pilot lines 112a to 112f, respectively, instead of the springs 45 to 50 of the first embodiment, and Driving units 45A to 50A are provided which urge the valve body 40A to 40A in the valve opening direction with the same force f as the springs 45 to 50.

Electrical signal X3 output from operation detector 110.111
, X4 are connected to the controller 61 together with the electrical signals X1 and
A, the controller 61A uses the electric signals X1, X2, X3, and X4 to calculate the values of the control forces Fc1 to Fc6 to be applied by the drive units 35C to 40c of the branch compensation valves 35A to 4OA, and takes appropriate action. Electrical signals a, b, "c, d, e.

Output f.

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

The electromagnetic proportional pressure reducing valves 62a to 62f, the relief valve 64, and the pressure reducing valve 113 are preferably assembled into one block as shown by a chain double-dashed line 66A.

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

The contents of the calculations performed by the calculation section of the controller 61A are shown in FIG. 16 in a functional block diagram. In this embodiment, in addition to the function block 83, a second function block 83A is provided as a function block corresponding to the shunt compensation valve 38, and the current electric signal
Control force value F C4 corresponding to differential pressure ΔPLS based on 1
, Fe12 are determined, and one of them is selected by the switch function of the selection blower 114. In addition, the electric signals X3 and X4 from the operation detector ito and 111 are A
is input to the ND block 115, and both signals are ON.
At this time, an ON signal is output from the AND block 115 to the selection block 114. The selection program 114 is
ON from the ND block 115 (when there is no word, control force FC40 is selected; when an ON signal is given, control force FC4 is selected.

Differential pressure P[S and control force Pc4 stored in function block 83
The relationship is as explained in the first embodiment.

Differential pressure PLS and control force F stored in function block 83A
The relationship with C40 is the same as the functional relationship stored in the function block 84.85 corresponding to the shunt compensation valve 39.40 related to the arm cylinder 27 and baguette cylinder 28, which was explained in FIG. 4D in the first embodiment. be. That is, overall, the value of the control force Fc4Q gradually decreases as the differential pressure ΔP[S increases, in accordance with 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, Differential pressure ΔP
[Regardless of the decrease in S, the relationship is such that the biasing force f of the drive unit 48A is limited to the maximum value fmax which is less than or equal to the biasing force f.

In the second example of repair constructed in this way, the boom 1
03 and the driven member other than the rotating body 100. During this combined operation, the two turning direction switching valves are not operated, so the operation detector 110 does not output the electric signal x 3, and the controller 61A outputs the AND block 115. is ON
Without outputting a signal, the selection block 114 selects the control force F c4o determined by the function block 83A as the control force. For this reason, a control force FC40 based on the basic function is applied to the drive unit 38c of the shunt compensation valve 38A, and the second control force f-Fc4o in the valve opening direction is the target of the differential pressure ΔPv4 across the flow control valve 32. The value is approximately equal to the differential pressure ΔPLS. That is, the second control force f-Fc4° has a normal value smaller than the second control force f-F4 due to the control force FC4 of the function block 83. As a result, when the boom cylinder 26 is on the low load pressure side, the amount of restriction of the flow control valve 38A does not become too small.
Control the differential pressure across the front and rear of the unit so that it almost matches the differential pressure ΔPLS,
Pressure oil can be supplied to the boom cylinder 26 at an appropriate flow rate according to the amount of operation of the flow control valve 32.

When performing a combined operation of the rotating body 100 and the boom 103,
Since both flow rate control valves 29 and 32 are operated, electrical signals X3 and X4 are output from both operation detectors 110 and 111, and in the controller 61A, the A N D block 115 outputs an ON signal, and the selection block 114
is the control force Fc4 obtained in function block 83 as the control force.
Select. For this reason, as in the case of the combined operation such as turning and boom raising described in the first embodiment, the flow compensating valve 35
The second control force f-Fc1.38 in the valve opening direction is applied to f-Fc1.
f-Pc4 has a relationship of f-Fcl<f-Fc4, and the boom cylinder 26 is supplied with a flow rate larger than the flow rate obtained by distributing the discharge amount of the main pump 22 by the opening ratio of the flow rate control valve 29.32, A combined operation of turning and raising the boom is performed, in which the boom raising speed is fast and the turning is relatively slow.

In addition, in this embodiment, the second flow compensation valve 35A to 40A is
One of the driving means related to the control force of is replaced with a spring, and the same reference pilot pressure Pr is guided through the pilot pipes 112 and 112a to 112f to the driving parts 45A to 5O.
It is set as A. Therefore, there is little variation due to spring manufacturing errors or changes over time, and the flow compensation valves 35A to 40A are mutually
The driving error can be extremely reduced. As a result, individual second control forces f-Fc1. f-Fc2f-F
c3. f-Fc4. f-Fc5. f-Fc6 can be realized more accurately than when using a spring, and the intended composite operation can be performed accurately.

Furthermore, in this embodiment, the reference pilot pressure P" guided to the drive units 45A to 50A is output from the pressure reducing valve 113, and the pressure reducing valve 113 is the electromagnetic proportional pressure reducing valve 62a to 62f.
The same configuration uses the pilot pressure set by the relief valve 64.

By the way, in the relief valve 64 configured as shown in the figure, when the tank pressure changes due to return oil from the actuator, etc., the pilot pressure, which is the output of the relief valve 64, also changes in accordance with the change. When the pilot pressure changes, even if the electric signals a to f are constant, the outputs of the electromagnetic proportional pressure reducing valves 62a to 62f, that is, the control pressures Pc1 to
Pc6 changes. Therefore, whether the driving parts 45A to 50A or 1
Assuming that the applied force is constant, the second control force in the valve opening direction varies even though the electric signals a to f are constant.

On the other hand, in this embodiment, the output of the pressure reducing valve 113, that is, the standard pilot pressure P'' also changes as the pilot pressure fluctuates. That is, when the control pressures Pc1 to PcG change, the standard pilot pressure corresponds to the change. Pr also changes.

Therefore, the two changes cancel each other out, and as a result, the second control force in the valve opening direction becomes constant. Therefore, in this example,
Changes in tank pressure due to return oil from the actuator do not affect the driving of the branch compensating valves 35A to 4OA, and the operation of each branch compensating valve 35A is maintained regardless of changes in tank pressure.
The individual second control forces f-Fc1. f-Fc2. f-Fc3. fFc4
．． f-Fc5. f −Fc6 can be realized more accurately, and excellent control accuracy can be obtained.

The third embodiment of the present invention will be explained with reference to FIGS. 17 to 24. In FIGS. 1 to 24, the same members as those shown in FIGS. ing.

In FIG. 17, the shunt compensation valves 35B to 40B are driven by two driving elements, the springs 45 to 50 and the driving parts 35c to 40c of the first embodiment, as driving means related to the second control force in the valve opening direction. Instead, a single driving element, that is, driving portions 35d to 40d, is provided to force the valve bodies of the branch compensation valves 35B to 40B in the valve opening direction.
d via pilot lines 51a to 51f, and the second control forces f-Fc1. f-
Fc2. f-Pc3. f-Fc4. f-F
c5. This is a configuration in which f-Fc6 acts directly. Hereinafter, this second control force will be expressed as Hcl to Hc6, respectively.

Further, in this embodiment, the actuators 23 to 2
A selection device 120 is provided corresponding to 8 and includes six selection switch elements 120a to 120f, each of which is selectively operable to one of a plurality of positions by an operator, and each of the selection switch elements 120a to 120f is A selection command signal having contents corresponding to the selected position is output as electrical signals Y1 to Y6.

The controller 61B includes an input section, a storage section, a calculation section, and an output section as in the first embodiment. The electric signal X1 output from the differential pressure detector 59 and the electric signals ¥1.about.Y6 output from the selection device 120 are input to the input section of the controller 61B, and the electric signal x1 is input to the calculation section of the controller 61B. According to the function data and control program stored in the storage unit from Y1 to Y6, calculations are performed to obtain the values of the control forces HC1 to Fc6, and the values of the control forces are outputted from the output unit as electrical signals a to f. .

Block 80B in FIG.
-85B are function blocks provided corresponding to the branch compensating valves 35B-40B and pre-stored with function data including a plurality of relationships between the differential pressure ΔPLS and the control forces Hc1-C6. In function blocks 808-85B,
One functional relationship corresponding to the content of the selection command signal is selected based on the electric signals Y1 to Y6, and further, from these selected functional relationships, the value of the control force corresponding to the differential pressure ΔPLS based on the electric signal X1 at that time is determined. HCI to Hc6 are each calculated. In this way, the function block 80B~
The control force values Hcl to HC6 determined in step 85B are filtered by first-order delay elements in delay blocks 90 to 95, respectively, and then output as electrical signals a to f.

Differential pressure ΔP[s and control force F stored in function block 80B
In FIG. 19, which shows multiple functional relationships between C1 and Fc6, the solid line So corresponds to the characteristics of the basic function explained in the first embodiment, and is the relationship between the discharge pressure of the main pump 22 and the actuators 23 to 28. The functional relationship is such that as the differential pressure ΔPLS with respect to the maximum load pressure increases, the control force (IC1) gradually increases. This functional relationship SO is the shunt compensation valve 35B.
It is used during normal driving of the swing motor 23, including independent operation of the swing body 1°, without the need to correct the second control force in the valve opening direction.

The broken lines So+1 and So+2 indicate a functional relationship in which the control force Hcl gradually increases with a gradient greater than the numerical aperture So as the differential pressure ΔPLS increases, and the broken line 5o-1
, 5o-2 represents a function that gradually increases the control force HC1 at a slope smaller than the function So as the differential pressure ΔPLS increases.

That is, the dashed lines So+1 and So+2 have a larger slope than the characteristic line SO of the basic function, and the second control force HC1 in the valve opening direction of the branch compensating valve 35B is made larger than when using the basic function, and The differential pressure is defined as the differential pressure ΔPLs between the main pump 22 and the maximum load pressure of the actuators 23 to 28.
It is a functional relationship that makes it larger than . This functional relationship is used when the flow rate supplied to the swing motor 23 is not greater than the normal case in a complex operation in which the swing motor 23 is on the low load pressure side.

Broken lines So+1 and 5o-2 make the second control force in the opening direction of the branch compensating valve 35B smaller than when using the basic frequency, and make the differential pressure across the two flow control valves smaller than the differential pressure ΔP[S. This is a functional relationship that is used when the flow rate supplied to the swing motor 23 is not lower than the normal case in a complex operation in which the swing motor 23 is on the low load pressure side.

Note that, as in the case of the first embodiment, ΔP 1SO is the differential pressure between the discharge pressure of the main pump 22 and the maximum load pressure maintained by the discharge amount control device 41 of the load sensing control type, that is, the pressure of the control valve 53. This is the load sensing compensation differential pressure set by the spring 54.

In the other function blocks 81B to 85B, a plurality of functional relationships are stored in substantially the same way as in the function block 80B. Note that the number and type of a plurality of functional relationships stored in each of the function blocks 80B to 85B are determined by the related actuators 23 to 2 depending on the type and content of the work during the composite operation.
8 is determined to give M-suitable operating characteristics.

Electric signals a to f output from the controller 61B are transmitted to a plurality of electromagnetic proportional pressure reducing valves 62a to 62 as in the first embodiment.
It is input to f. The electromagnetic proportional pressure reducing valves 62a to 62f are driven by the electrical signals a to f, respectively, and output corresponding control pressures Pc1 to PC6. These control pressures PC1 to Pc6 are controlled by the branch compensating valves 35B to 40B, respectively.
The flow control valves 35B to 40B are guided to the drive units 35d to 40d, and the control quadrants Icl to Hc6 calculated by the controller 61B are applied to the branch flow compensation valves 35B to 40B. 34 front and rear differential pressure ΔP
Control v1 to ΔPv6.

Next, the operation of this embodiment configured as above will be explained.

For example, when performing a combined operation of turning and raising the boom for earth and sand loading work, the operator selects a functional relationship suitable for the content of the work by selecting the corresponding selection switch element 120a of the operating device 120.

120d to output corresponding selection instruction signals, that is, electrical signals Y1 and Y4. In the controller 61B, based on the electrical signals Y1 and Y4, for example, the broken line 5o-2 in FIG. Select the functional relationship corresponding to
For the branch compensating valve 38B corresponding to the boom cylinder 26, a functional relationship corresponding to, for example, the broken line SO+2 in FIG. 19 is selected from a plurality of functional relationships stored in the function block 83B.

In FIG. 20, 121 is the basic function So
122 is a characteristic line corresponding to the functional relationship of the broken line 5O-2 selected in the function block 80B corresponding to the swing motor 23, and 123 is a characteristic line corresponding to the function block 83B corresponding to the boom cylinder 26. This is a characteristic line corresponding to the functional relationship of the selected broken line SO+2.

Further, in the function blocks 80B and 83B, control forces H1 and H4 based on the differential pressure ΔP[S are obtained from the selected functional relations 122 and 123, respectively, and the corresponding electric signals a and d are sent to the electromagnetic proportional pressure reducing valve 62a. , 62d.

As a result, the electromagnetic proportional pressure reducing valve 62d outputs a control pressure Pc4 which also has a large DI control positive pressure corresponding to the control card IO based on the differential pressure ΔPLS, while the electromagnetic proportional pressure reducing valve 62a
outputs a control pressure Pc1 smaller than the control pressure corresponding to the control card IO, and these control pressures PCI and PC4 are guided to drive parts 35d and 38d of the branch compensation valves 35B and 38B, respectively. In this case, since the driving part 38d of the branch compensation valve 38B applies a control force H4 larger than the normal control force HO, the branch compensation valve 38B is controlled so that its throttle amount is forcibly reduced. Therefore, the flow control valve 3
2 is supplied with a larger flow rate than normal, and the drive section 35d of the branch compensation valve 35B is supplied with the normal control force H.

Since a smaller control angle 1 is provided, the flow rate compensating valve 35B is controlled so that its throttle amount is forcibly increased, and therefore a smaller flow rate than normal is supplied to the two flow rate control valves. .

Figures 21 and 22 show the flow rate characteristics at this time. Figure 21 shows the relationship between the pressure difference ΔPv4 across the flow control valve 32 for the boom and the supply flow rate Q4, and Figure 22 shows the relationship between the flow rate control valve 32 for the boom and the supply flow rate Q4. The relationship between the differential pressure ΔPv1 before and after the two flow rate control valves and the supply flow rate Q1 is shown.

Here, the characteristic line 123 for the characteristic line 121 of the basic function is
If the slope ratio of is α, then the boom flow control valve 32
On the side of , the flow rate Q4A, which is relatively small as shown in the characteristic line 124A in FIG. As shown by a characteristic line 124B in FIG. 21, a flow rate 04B larger than the flow rate Q4A can be supplied depending on the pressures α and ΔPLS. In addition, the characteristic line 121 of the basic function
If the ratio of the slope of the characteristic line 122 to
In the case of control by S, the relatively large flow rate QIA as shown in the characteristic line 125A in Fig. 22 is changed to the corrected differential pressure β・ΔP [in accordance with S in Fig. 22 during this earth and sand loading work. As shown in characteristic line 125B, the flow rate QI
A smaller flow rate QIB than A can be supplied.

That is, during earth and sand loading work, 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, compared to during normal control.
The optimal flow rate can be distributed to the boom cylinder 26 and the swing motor 23 according to this soil loading work, thereby reducing the flow rate relieved on the swing motor 23 side and reducing the throttle amount of the flow compensation valve 38B on the boom cylinder 26 side. By making this smaller, it is possible to suppress the energy of the pressure oil passing through this branch compensating valve 38B from being converted into heat, thereby making it possible to reduce energy loss. Also,
Since a relatively large flow rate can be supplied to the boom side, sufficient lift of the boom can be ensured, providing excellent workability.

Next, when performing combined operation of the arm and packet for excavation work aimed at improving work efficiency compared to normal excavation work, that is, for special excavation work, the operator must select a functional relationship suitable for the work content. The corresponding selection switch elements 120e and 12Of of the operating device 120 are operated to select the corresponding selection instruction signals, that is, the electrical signals Y5 and Y.
Outputs 6. In the controller 61B, based on the electric signals Y5 and Y6, a function block 84 is set for the branch compensation valve 39B corresponding to the arm cylinder 27.
For example, the functional relationship corresponding to the broken line 5O-1 in FIG. 19 is selected from the plurality of functional relationships stored in the function block 85B. For example, a functional relationship corresponding to the broken line So+1 in FIG. 19 is selected from a plurality of functional relationships.

FIG. 23 collectively shows the functional relationships selected by function blocks 84B and 85B'''C''. In the figure, 121 is a characteristic line corresponding to the basic function SO, 126 is a characteristic line corresponding to the functional relationship of the broken line 5O-1 selected in the function block 84B corresponding to the arm cylinder 27, and 127
is the function block 85B corresponding to the packet cylinder 28.
This is a characteristic line corresponding to the functional relationship of the broken line s o+1 selected in .

Furthermore, in the function blocks 84B and 85B, control quadrants I5 and H6 based on the differential pressure ΔP[S are obtained from the selected functional relationships 126 and 127, respectively, and the corresponding electric signals e and f are sent to the electromagnetic proportional pressure reducing valve 62e. , 62f.

As a result, the electromagnetic proportional pressure reducing valve 62e outputs a control pressure PC5 smaller than the control pressure corresponding to the control force Ha based on the differential pressure ΔPLS, while the electromagnetic proportional pressure reducing valve 62f outputs a control pressure PC5 smaller than the control pressure corresponding to the control force HO based on the differential pressure ΔPLS. Also large control pressure PC
6, and these control pressures Pc5. Pc'6 is guided to drive parts 39d and 40d of branch compensation valves 39B and 40B, respectively. In this case, the drive section 39 of the shunt compensation valve 39B
Since d applies a control force H5 smaller than the normal control force Ho, the branch compensation valve 39B is controlled so that its throttle amount is forcibly increased, and therefore the flow rate control valve 33 has a smaller amount than normal. The flow rate is supplied, and the driving portion 40d of the branch compensating valve 40B is supplied with the normal control force H.

Since a larger control angle 5 is provided, the flow rate compensating valve 40B is controlled so that its throttle amount is forcibly reduced, and therefore, a larger flow rate than normal is supplied to the flow rate control valve 34.

This makes the driving speed of the arm cylinder 27 relatively slow and the driving speed of the packet cylinder 28 relatively fast during the combined operation of the arm and packet, which is considered to be better in terms of work efficiency than normal excavation. Special excavation work can be realized.

Next, even in the same arm and packet combination operation, when performing a combined arm and packet operation intended for flattening the ground, etc., the operator should select a functional relationship suitable for the work content. The corresponding selection switch elements 120e and 12Of of the operating device 120 are operated to generate the corresponding selection instruction signals, that is, the electric signals Y5 and Y.
Outputs 6. In the controller 61B, based on the electric signals Y5 and Y6, a function block 84 is set for the branch compensation valve 39B corresponding to the arm cylinder 27.
Select, for example, the functional relationship corresponding to the broken line So+1 in FIG.
For example, a functional relationship corresponding to the broken line 5o-1 in FIG. 19 is selected from the plurality of functional relationships stored in the function block 85[3.

FIG. 24 shows a summary of the functional relationships selected in function blocks 84B and 85B. In the figure, 121 is the basic function So
128 is the characteristic line corresponding to the arm cylinder 27.
This is a characteristic line corresponding to the functional relationship of the broken line SO+1 selected by the function block 84B'''C'' corresponding to 129
is the function block 85B corresponding to the packet cylinder 28.
This is a characteristic line corresponding to the functional relationship of the broken line 5O-1 selected in .

In the function blocks 84B' and 85B, the control force H'5. based on the differential pressure ΔP[S is further calculated from the selected functional relationships 128 and 129. H'6 is obtained, and the corresponding electric signals e and f are output from the electromagnetic proportional pressure reducing valves 62e and 62f.
is output to.

As a result, the electromagnetic proportional pressure reducing valve 62e outputs a control pressure Pc5 larger than the control pressure corresponding to the control force HO based on the differential pressure ΔPLS, while the electromagnetic proportional pressure reducing valve 62f outputs a control pressure Pc5 larger than the control pressure corresponding to the control force HO based on the differential pressure ΔPLS. Even small control pressure PC
6, and these control pressures Pc5. Pc6 is guided to drive parts 39d and 40d of branch compensation valves 39B and 40B, respectively. In this case, the drive part 39d of the branch compensation valve 39B
applies a control force H'5 larger than the normal control force Ho, so the branch compensation valve 39B is controlled so that its throttle amount is forcibly reduced, and therefore the flow rate control valve 33 is given a control force H'5 that is larger than the normal control force Ho. Since a large flow rate is supplied and the drive unit 40d of the branch compensation valve 40B applies a control force H'6 smaller than the normal control force HO, the amount of restriction of the branch compensation valve 40B is forced to increase. Therefore, the flow rate control valve 34 is supplied with a smaller flow rate than usual.

This makes it possible to make the driving speed of the arm cylinder 27 relatively fast while making the driving speed of the packet cylinder 28 relatively slow during combined operation of the arm and the packet, thereby realizing ground grading, that is, shaping work with good work efficiency. .

Modification of Third Embodiment A modification of the third embodiment described above will be explained with reference to FIG. 25. In the figure, elements equivalent to those shown in FIG. 18 are given the same reference numerals.

In this embodiment, instead of the selection device 120 described above, for example, five selection switch elements 1 are provided corresponding to each work mode and can be selectively operated by the operator.
A selection device 130 is provided including 30a-130e. The selection switch elements 103a to 130e each output a selection command signal according to the corresponding work mode as electric signals Za to ZO in accordance with their operation, but sometimes only one of them is operated. The selection device 130 outputs one of the electrical signals Za to Ze, corresponding to the operated selection switch element.

The controller 61C includes an input section, a storage section, a calculation section, and an output section as in the first embodiment. The electric signal X1 output from the five differential pressure detectors and one of the electric signals Za to Ze output from the selection device 130 are input to the input section of the controller 61C.
In the calculation unit, in the function selection instruction block 131,
Function blocks 80B to 85 according to the input electrical signal
B and a plurality of functional relationships stored in the selected function block are selected, and corresponding selection command signals 21 to Z6 are output. Function blocks 80B to 85B
, the electrical signal X1 and selection command signals 21 to Z6
According to the function data and control program stored in the storage unit, calculations are performed to obtain the values of the control forces HC1 to Fc6, and the values of the control forces are outputted from the output unit as electrical signals a to f.

In this embodiment configured in this way, the selection switch elements 130a to 13 of the selection device 130 are designed to perform earth and sand loading work by a combined operation of turning and lifting the boom, for example.
0e, for example, the selection switch element 130a, the selection device 130 outputs an electric signal Za. Function selection instruction block 131 of controller 61C
In the function block 80B based on the electric signal Za
.

83B, for the function block 80B, the function indicated by the broken line 5O-2 shown in FIG. Among the functional relationships, the broken line S in Figure 19
An operation is performed to select the o+2 function, and corresponding selection command signals Zl and Z4 are output. Note that another function block 81B.

For 82B, 84B, and 85B, the basic function SO shown in FIG. 19 is selected, and corresponding selection command signals Z2, Z3, Z5, and Z6 are output.

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

In addition, the selection device 1 was designed for arm and packet excavation work with the aim of improving work efficiency compared to normal excavation work.
When one of the thirty selection switch elements 130a to 130e, for example, the selection switch element 130b, is operated, the selection device 130 outputs an electrical signal zb. The function selection instruction block 131 of the controller 61C selects the function blocks 84B and 85B based on the electric signal zb, and also selects the function block 84B based on the broken line 5O shown in FIG. 19 among the plurality of function relationships. -1
and for function block 85B,
An operation is performed to select the function indicated by the broken line So+1 in FIG. 19 from among a plurality of functional relationships, and corresponding selection command signals Z5 and Z6 are output.

As a result, in the function blocks 84B and 85.8, the functional relationship indicated by the selection command signals Z5 and Z6 is selected, and as in the above embodiment, the driving speed of the arm cylinder 27 is controlled during the combined operation of the arm and the packet. By making the driving speed of the packet cylinder 28 relatively slow and relatively fast, it is possible to realize a special excavation operation that is considered to be better in terms of work efficiency than ordinary excavation.

Furthermore, when one of the selection switch elements 130a to 130e, for example, the selection switch element 130C, of the selection device 130 is operated with the intention of flattening the ground or the like by a combined operation of an arm and a packet, the selection device 13
From 0, an electrical signal of 2C is output. controller 61
In the function selection instruction block 131 of C, function blocks 84B and 85B are selected based on the electric signal Zc, and for the function block 84B, the broken line SO+1 shown in FIG. After selecting the function, for the function block 85B, an operation is performed to select the function indicated by the broken line 5O-1 in FIG. 19 from among the plurality of functional relationships, and the corresponding selection command signal Z5,
Output Z6.

As a result, in function blocks 84B and 85B,
The functional relationship indicated by the selection command signals Z5 and Z6 is selected, and as in the above-described embodiment, the drive speed of the arm cylinder 27 is made relatively fast, and the drive speed of the packet cylinder 28 is made relatively fast, thereby improving work efficiency. It is possible to achieve good shaping work.

In the above embodiment, each of the selection switch elements 130 of the selection device 130 is configured to output a single selection command signal Za to ze in response to its operation, but each can be operated in a plurality of stages, and the same operation can be performed. The function selection instruction block 131 selects different functional relationships of related function blocks in accordance with this selection instruction signal, and selects a different function relationship for the shunt compensation valve in response to this selection instruction signal. This makes it possible to change the matching settings for complex operations depending on the work situation, further improving work efficiency and efficiency.

Other Examples of Control Pressure Main Circuit In the above embodiments, in the control pressure generation circuit, control pressures Pc1 to Pc1 to Pc are generated in response to electrical signals a to f from the controller.
Although the configuration employs the electromagnetic proportional pressure reducing valves 62a to 62f as the control pressure generation circuit that outputs C6, other configurations may also be adopted as the control pressure generation means. This embodiment shows the possibility of this point. It is something.

That is, in this embodiment, the control pressure generation circuit 140 includes an electromagnetic variable relief valve 141a interposed between the pilot pump 63 and the tank and connected in parallel to each other.
~141f and the electromagnetic variable relief valves 141a~14
1f and the pilot pump 63, respectively, and the electromagnetic variable leaf valves 141a to 141f receive electric signals a to f from the controller 61 shown in FIG. is supplied, and the electromagnetic variable relief valves 141a to 141f operate according to the electric signal, and the pilot lines 143a to 143f between the throttle valves 142a to 142f and the electromagnetic variable relief valves 141a to 141f are connected to the pilot lines 51a to 5.
For example, the branch compensating valves 35 to 40 shown in FIG.
6, which is configured to communicate with the drive units 35c to 40c of the
Also in the control pressure generation circuit 140 configured in this manner, the electromagnetic variable relief valves 141a to 141f are individually driven according to the electric signals a to f output from the controller.5. The amount of throttle is determined, and the pilot pump 63
Change the magnitude of the pilot pressure output from the
Control pressures Pc1 to PC at levels corresponding to electrical signals a to f
6 as pilot lines 143a to 143f and 5
1a to 51f, it can be supplied to the driving parts 35c to 40c of the branch compensating valves 35 to 40 shown in FIG. 1, for example, to obtain the same function as the electromagnetic proportional pressure reducing valve described above.

A fourth embodiment of the present invention will be described with reference to FIGS. 27 to 32.

In FIG. 27, the hydraulic drive system applied to the hydraulic excavator of this embodiment includes one variable displacement hydraulic pump, that is, a main pump 200 driven by a prime mover (not shown).
and a plurality of actuators driven by the pressure oil discharged from the main pump 200, that is, the swing motor 201 and the boom cylinder 202, and a flow control valve that controls the flow of the pressure oil supplied to each of the plurality of actuators. That is, the swing direction switching valve 203 and the boom direction switching valve 204 are arranged upstream of these flow rate control valves, and the differential pressure generated between the inlet and outlet of the flow rate control valves;
That is, it is provided with a pressure compensation valve, that is, a branch compensation valve 205206, which controls the difference between the front and rear of the flow rate control valve and the pressure, respectively.

In addition, a relief valve and an unload valve (not shown) are connected to the discharge pipe 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 pressure oil to flow out into the tank 208. The pump discharge pressure is prevented from reaching a high pressure higher than the set pressure, and the unload valve allows the pressure oil from the main pump 200 to reach the load pressure (hereinafter referred to as maximum) on the high pressure side of the swing motor 201 and boom cylinder 202. When the pressure reaches the sum of the load pressure (P alax) and the set pressure of the unload valve, it is caused to flow into the tank 208 and is prevented from exceeding the pressure.

The discharge amount of the main pump 200 is controlled by the discharge amount control device 209 so that the discharge pressure Ps is higher than the maximum load pressure P alIlax by a predetermined value ΔP LSO, and load sensing control is performed.

Meteor control valves 203 and 204 are pilot valves 21, respectively.
0.211, the pilot valves 210 and 211 generate pilot pressure a1 or a2 and pilot pressure b1 or b2 by manual operation of the operating lever, and the flow control valves 203 and 204 When the pilot pressure a1 or a2 and the pilot pressure b1 or b2 are added, the flow rate control valves 203 and 204 are opened to a corresponding throttle amount.

Diversion compensation valves 205 and 206 are not shown in FIG. 1 and are of the same type as the division compensation valve in the first embodiment.

That is, the drive units 205a, 205b and 206 are guided by the outlet pressure and inlet pressure of the flow rate control valves 203 and 204, respectively, and apply a first control force in the valve closing direction based on the pressure difference between the front and rear.
a, 206b, the spring 212213, and the electromagnetic proportional pressure reducing valve 216.21 via the pilot lines 214, 215.
a drive unit 205c to which the control pressure output from 7 is guided;
206c, the springs 212, 213 and the drive unit 205
c206C applies a second control force in the valve opening direction that provides the t3 value of the front and rear differential pressure.

Pilot pressure is supplied from a common pilot pump 220 to the discharge amount control device 209, the pilot valve 210211, and the electromagnetic proportional pressure reducing valves 216, 217. The flow control valves 203 and 204 each have a swing motor 201.
and a shuttle valve 222 for deriving the maximum load pressure of the boom cylinder 202.

The hydraulic drive system of this embodiment further includes a main pump 2.
A displacement detector 223 detects the displacement corresponding to the displacement of 00 and detects the discharge amount Qθ of the main pump 200; a discharge pressure detector 224 detects the discharge pressure PS of the main pump 200; Pressure ps and maximum load pressure P a of swing motor 201 and boom cylinder 204
la
24 and the differential pressure detector 225 are input, and an operation command signal 311. It has a controller 229 that outputs S12 and S21.322.

The configuration of the discharge amount control device 209 is shown in FIG. 2-8.

This embodiment is an example in which the discharge amount control device 209 is configured as an electro-hydraulic servo type hydraulic drive device.

The discharge amount control device 209 includes a servo piston 23 that drives the displacement variable machine M 200 a of the main pump 200.
0, and the servo piston 230 has the servo cylinder 23
It is stored in 1. The cylinder chamber of the servo cylinder 231 is divided into a left chamber 232 and a right chamber 233 by the servo piston 230, and the cross-sectional area of the left chamber 232 is larger than the cross-sectional area d of the right chamber 233.

The left chamber 232 of the servo cylinder 231 is connected to the line 234.
The right chamber 233 is connected to the pilot pump 218 through the line 235, and the right chamber 233 is connected to the pilot pump 218 through the line 235.
18, and lines 234, 235 are connected to tank 208 via return line 236. A solenoid valve 237 is interposed in the line 235, and the return line 2
36 is provided with a solenoid valve 238. These solenoid valves 237 and 238 are normally closed solenoid valves (a function that returns to the closed state when no electricity is applied), and an operation command signal 311S12 from the controller 229 is input to these solenoid valves 237 and 238. 〃J magnetic field and are respectively switched to the open position.

When the operation command signal S11 is input to the solenoid valve 237 and it is 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 moves to the right in the figure due to the difference in area between the left chamber 232 and the right chamber 233. As a result, the tilt angle of the displacement variable 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 its original closed position, and the left side chamber 2
32 and the right chamber 233 is cut off, and the servo piston 230 is held stationary at that position. This keeps the displacement of the main pump 200 constant,
The discharge amount becomes constant. Operation command signal S to solenoid valve 238
12 is input and switches to the open position, the left side chamber 232
The pressure in the left chamber 232 is reduced by communicating with the tank 208, and the servo piston 230 is moved to the left in the figure by the pressure in the right chamber 233. As a result, the displacement volume of the main pump 200 decreases, and the discharge amount also decreases.

In this way, the solenoid valves 237 and 238 are operated by the operation command signal S11.
．． The main pump 200 is controlled on/off by S12.
By controlling the displacement of the main pump 200
The discharge amount is controlled to be equal to the target discharge amount QO calculated by the controller 229.

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

The contents of the calculations performed by the calculation section of the controller 229 are shown in FIG. 29 in a functional block diagram.

In FIG. 29, blocks 240, 241242 are
The differential pressure ΔP [S detected by the differential pressure gauge 43 is calculated as the load sensing compensation differential pressure, that is, the target differential pressure ΔP LSO
This is a block for determining the target discharge differential pressure ff1QΔp. In this example, the differential pressure target discharge! QΔp is determined based on the following formula.

QΔp = g(ΔPLS) Σ 1ku 1 (Δ p tso −Δ Pl,
5)=ll(Δp tso−ΔPLS) +QO−1Δ
QΔI)+QO-1...(1) However, KI: Integral gain ΔPLSO: Target differential pressure Qo-1: Discharge amount target value output in the previous control cycle ΔQΔp: Differential pressure of control 1 cycle time Target discharge amount This is an example in which the increment, that is, the differential pressure target discharge IQΔp is calculated by the integral FFIIt:IJ method of the deviation between the target differential pressure ΔPLSO and the actual differential pressure, and blocks 240 and 241 are calculated based on the differential pressure ΔPLSO.
Calculate Kl (ΔP LSO - ΔPLS) from Control 1
This is to find the increment ΔQΔp of the differential pressure target discharge amount per cycle time, and in block 242, the increment ΔQΔp and the discharge amount target value Q o−1 output in the previous control cycle are calculated.
and 7 to obtain equation (1).

In this example, QΔp was obtained using an integral control method, but a different method is used, for example, QΔρ=Ko(Δp tso−ΔP LS)...
・(2) However, Kl) is a proportional control method expressed by proportional gain, or equation (1) and (2)
It may also be determined using a proportional/integral control method that adds the equations.

Block 243 is a function block that determines the input limit target discharge JiQT from the discharge pressure ps of the main pump 200 detected by the pressure detector 224 and the input torque limit function f (PS) stored in advance. FIG. 30 shows the input torque limiting function f(Ps). Main pump 2
The input torque of 00 is proportional to the displacement of the main pump 200, that is, the product of the tilting amount of the swash plate and the discharge pressure ps. Therefore, the input torque limiting function f (Ps) uses a hyperbola or an approximate hyperbola. That is, it is a function expressed by the equation of TP: input limit torque: proportionality constant. The input limit target discharge iQT can be determined from this input torque limit function f (Ps ) and the discharge pressure Ps.

The magnitude of the differential pressure target discharge amount QΔρ and the input limit target discharge amount QT obtained in the above manner is determined in the minimum value selection block 204, and when QΔp≦Q■, the discharge amount target value Qo is set as QΔρ. , and if QΔo >QT, select QT. That is, the differential pressure target discharge amount QΔp
The smaller of input limit target discharge amount QT is the discharge amount target (f
i is selected as QO, and the input limit target discharge amount determined by the discharge 1 target value QO or the input torque limit function (PS) is not exceeded.

In blocks 255, 256, and 257, the solenoid valve 20 of the discharge amount control device 209 is
The operation command signals 311 and 312 for 7.238 are generated once.

Specifically, first in block 255, ZQO-Q
θ is calculated, and the discharge amount target value Qo and the detected discharge amount Qθ are calculated.
Find the deviation Z. Then block 256.257
, when the deviation Z exceeds a preset dead zone Δ, the operation command signal S11 or S1 is
2 is generated.

That is, if the deviation Z is positive and becomes less than the dead zone Δ, block 25 is executed.
At step 6, the operation command signal S11 is generated, and the solenoid valve 237 of the discharge amount control device 209 is turned on.

As a result, as described above, the tilt angle of the main pump 200 increases, and the discharge amount Qθ is controlled to match the discharge amount target value Qo. When the deviation Z is negative and becomes less than or equal to the dead zone Δ, an operation command signal S12 is generated in block 257, and the solenoid valve 2
37 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 the detected discharge amount Qθ is controlled to match 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 limit target discharge amount QT, the discharge amount of the main pump 2°O is set to the differential pressure target discharge amount ff1QΔp. controlled, main pump 200
The differential pressure ΔPLS between the discharge pressure and the maximum load pressure is maintained at the target differential pressure ΔPLSO. That is, load sensing control is performed to maintain the differential pressure ΔP[S constant. On the other hand, when the differential pressure target discharge rate QΔp becomes larger than the input limit target discharge rate QT, the input limit target discharge rate 0 is selected as the discharge rate target value QO, and the discharge rate is set so as not to exceed the input limit target discharge rate QT. controlled. That is, input restriction control of the main pump 200 is performed.

On the other hand, the differential pressure target discharge amount QΔρ and the input limit target discharge amount QT
is determined as a deviation in block 258, and the target discharge amount surface difference ΔQ
seek.

Next, in blocks 259, 260, and 261, a basic value, that is, a basic correction value, for the total consumable flow rate correction control of the branch compensating valves 205, 206 (see FIG. 27) is determined from the target discharge amount 10 difference ΔQ obtained in block 258. Calculate Qns. 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)=1nS in Σ・ΔQKlns−ΔQ
+ Q ns-1 = ΔQ ns + Q ns-1 = - (4) However, KInS
: Integral gain QnS-1: Basic correction value ΔQ ns output in the previous control cycle: Increment of the basic correction value of control m1 cycle time, that is, in block 259, control 1 is calculated from the target discharge amount deviation ΔQ obtained in block 258. The increment ΔQns of the basic correction value per cycle time is determined by K1n5·ΔQ. Then, an addition block 260 adds this value to the basic correction value Qns-1 output in the previous control cycle to obtain an intermediate value Q'nS, and a block 261 having a limiter function shown in FIG. When ns<O, QnS
=O, and when Q'nS≧0, Q'ns <
When Q'nsc, a basic correction value Qns that increases in proportion to the increase in Q'ns is output, and Q'nS≧Q'nsc.
In this case, the basic correction value QnS is determined so that Q nS=Q n5IlaX. Here Q nsn+a:x
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 block 261 is further applied to function block 262 provided for each actuator 201 and 202.
, 263 and a different operation command signal 321
．． Get 322.

Basic correction value Q stored in function blocks 262 and 263
ns and the operation command signal S21. The relationship with S22 is the 32nd
In the figure, 264 is a characteristic for the operation command signal S21, and 265 is a characteristic for the operation command signal S22. Further, 266 is a characteristic that does not change the basic correction value Qns. That is, the operation command signal S21 is the basic correction value Q.
The operation command signal S22 is corrected to a value larger than ns, and the operation command signal S22 is corrected to a value smaller than the basic correction value QnS.

Operation command signal S21. determined in blocks 262 and 263.
S22 is an electromagnetic proportional pressure reducing valve 216.21 shown in FIG.
7, and the electromagnetic proportional pressure reducing valves 216 and 217 are driven by this signal to generate control pressure at the corresponding level, and this is transmitted to the drive section 205c of the branch compensation valve 205 and 206.
206c.

As a result, the second control force in the valve-opening direction applied to the shunt compensation valves 205 and 206 is greater than that in the case where the basic correction value Qns or the command signal is output.
, and is corrected to become larger in the shunt compensating valve 206 , and correspondingly, the shunt compensating valve 205
．． The diversion ratio by 206 is corrected.

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

For example, when the boom pilot valve 211 is slightly operated to drive the flow rate control valve 204 and the boom is operated independently, the required flow rate is small, so the differential pressure target discharge amount QΔp is set as the input limit target discharge in the controller 229. A value smaller than the quantity QT is calculated, and the differential pressure target discharge iQΔρ is selected as the discharge quantity target value Qo. This lick, main pump 20
Load sensing control is performed in which the differential pressure ΔP[S between the discharge pressure of 0 and the maximum load pressure is maintained at the target differential pressure Δptso. On the other hand, the basic correction value Qns is calculated to be zero, the second control force in the valve opening direction is applied to the branch compensation valves 205 and 206 only by the force of the springs 212 and 213, and the boom cylinder 202 is given a second control force in the valve opening direction. A flow rate is supplied according to the opening degree.

When operating the pilot valves 210 and 211 at the same time to perform a combined operation of, for example, swinging and raising the boom, the required flow rate is large and the load pressure of the swing motor 201 is high, so the controller 229 inputs the differential pressure target discharge amount QΔp. A value larger than the limit target discharge iQT is calculated, and either the input limit target discharge ff1Q■ or the discharge amount target value QO is selected. Therefore, the discharge amount of the main pump 200 is controlled so as not to exceed the input limit 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. This basic correction value Q
The electromagnetic proportional pressure reducing valve 21 uses nS as an operation command signal.
6,217, the shunt compensation valve 205.2
The second control force in the valve opening direction of 06 is decreased at the same rate, and the target value of the differential pressure across the flow rate control valves 203 and 204 is decreased at the same rate. As a result, the flow rate supplied to the flow control valves 203 and 204 decreases at the same rate, and the actuator 2
The total amount of pressure oil consumed in 01202 can be reduced without changing the distribution ratio between the two, and the speed ratio of both actuators 201202 can be maintained. This is referred to herein as total consumable flow rate correction control.

In this embodiment, when this total consumable flow rate correction control is performed, the basic correction value Qns is further modified and a different operation command signal S21. S22 is determined and outputted to the electromagnetic proportional pressure reducing valves 216 and 217.

Because of this, the second control force in the valve opening direction applied to the shunt compensation valves 205 and 206 becomes smaller in the shunt compensation valve 205 compared to the case where the basic correction value Qns is output as a command signal. In 206, the total consumable flow rate is corrected to be larger, and while performing correction control, further,
Dividing control is performed so 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 first embodiment,
The combined operation of turning and raising the boom can be reliably performed, the boom raising speed is fast, and the turning is relatively gentle, thereby improving the combined operation and making effective use of energy.

As described above, in this embodiment as well, substantially the same effects as in the first embodiment can be obtained in the combined operation of swing and boom.

Mouse 51 Do 11 Man The fifth embodiment of the present invention will be explained with reference to FIGS. 33 to 38. In FIG. It is attached.

In FIG. 33, the hydraulic drive system of this embodiment has basically the same configuration as the fourth embodiment shown in FIG. 27. Therefore, explanations of parts having the same configuration will be omitted. Main pump 20
The discharge pipe 207 of the main pump 200 has a relief valve 300 that causes the pressure oil from the main pump 200 to flow out into the tank when it reaches the relief setting pressure, and prevents the pump discharge pressure from becoming higher than the set pressure. The pressure oil from the pump 200 is applied to the high pressure side of the swing motor 201 and the boom cylinder 202 at a load pressure (hereinafter referred to as the maximum load pressure PaIllax).
An unload valve 301 is connected thereto which causes the water to flow out into the tank when the pressure reaches the sum of the unload setting pressure and the unload setting pressure.

The discharge amount of the main pump 200 is determined by the swash plate of the main pump 200.
Drive cylinder 3 that drives 00a and increases/decreases displacement volume
02a, and an electromagnetic control valve 302b that controls the supply and discharge of pressure oil to the drive cylinder 302a and adjusts the displacement of the drive cylinder. 303 is a relief valve that sets the swing relief pressure of the swing motor 201.

The pilot valves 210 and 211 include the pilot valves 210 and 211.
, 211 are provided, respectively. Also, the main pump 200 is operated by an operator.
A selection device 306 is provided for externally selecting and setting a target value of the discharge pressure.

Detection signals from the displacement detector 223, discharge pressure detector 224, differential pressure detector 225, pilot pressure detectors 304, 305, and selection device 306 are input to the controller 307, where after predetermined calculations are performed, the discharge Quantity control device 3
02 electromagnetic control valve 302b and electromagnetic proportional pressure reducing valve 216,
Operation command signal S1 to drive parts 216c and 217c of 217
and outputs S21 and S22.

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 represents the differential pressure Δ
This is a function block that calculates the target discharge amount QO of the main pump 200 that maintains the differential pressure ΔPLS at the target differential pressure ΔPLSO from P[S. Differential pressure Δ stored in this function block 310
PLS and target discharge amount Q.

This functional relationship is shown in FIG. 35. This functional relationship is such that the target discharge amount Qo increases in proportion to the decrease in the differential pressure ΔP[S. Note that the target discharge amount QO is determined by blocks 240 to 2 shown in FIG. 29 in the fourth embodiment described above.
Similarly to 42, the calculation may be performed using an integral control method.

The target discharge ff1QO is the discharge amount Qθ of the main pump 200 detected by the displacement detector 223 in the addition block 311.
The deviation ΔQ from the
Output to b. As a result, the electromagnetic control valve 302b is driven, and the discharge pressure Ps is increased to the actuator 201. ．． The second block 313, in which the discharge amount of the main pump 200 is controlled so as to be higher than the maximum load pressure P alax of the main pump 202 by a constant value ΔP LSQ, calculates the control force signal 11 from the differential pressure ΔP[S.
This control force signal 11 is a function block that calculates
The main pump 200 is subjected to load sensing control by the discharge amount control device 302, and when the differential pressure ΔP[S does not reach the target differential pressure Δp tso even if the discharge amount of the main pump 200 reaches the maximum, the flow compensation valve 205 .206 drive unit 2
Control force Nc1.05c, 206c imparts. N2 is increased, and the second control force f-Nc1. in the valve opening direction is increased. By reducing Na3, that is, by reducing the target value of the differential pressure across the flow control valves 203 and 204, the increase in the absolute amount of the flow rate of pressure oil supplied to each actuator 201 and 202 is suppressed. The pump discharge amount is distributed according to the opening ratio of 203 and 204, that is, the ratio of the required flow rate. The function (system) between the differential pressure ΔP[S stored in the function block 313 and the control force signal 11 is shown in No. 36I7I. This functional relationship basically corresponds to the rotation shown in FIG. 4A of the first embodiment. Note that the control force signal 11 is used as the first command value of the control force Nc2 applied by the drive unit 206a to the branch compensation valve 206.

The block 314 is a function block that calculates the control force signal 12 that maintains the discharge pressure Ps at the target discharge pressure Pso using a proportional control method from the discharge pressure Ps of the main pump 200 detected by the discharge pressure detector 224. Signal 12 is used to obtain the second command value of control force Nc2. This function block 314 is a target discharge pressure Pso
It is configured such that it can be changed by a command signal r from a selection device 306. FIG. 37 shows the functional relationship between the discharge pressure Ps stored in the function block 314, the control force signal 12, and the command signal r. In FIG. 37, the function that is set when the command signal r is at the minimum value is shown in FIG. The related target discharge pressure is shown in pso.

Blocks 315 and 316 are blocks for calculating a control force signal i3 that maintains 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, and controls The force signal i3 is the second of the control force Nc2 together with the control force signal 12.
It is used once to obtain the command value. Here, block 3
15, the rate of change i3 of the control force signal i3 is determined from the discharge pressure Ps based on the functional relationship stored in advance, and this rate of change i3 is integrated by the block 316 to obtain the control force signal i3.
seek. Block 315, like block 314, is configured such that target discharge pressure Pso can be changed by command signal r from selection device 306. The function (system is shown in FIG. 38) between the discharge pressure Ps stored in the function block 315, the rate of change i3 of the control force signal i3, and the command signal r6. Note that also in FIG. A target discharge pressure related to a function that is set at a certain time is indicated by Pso.

The control force signal 12 obtained by the function block 3111 and the control force signal i3 obtained by the integration block 316 are added to the addition block 3.
17, and the second command value of the control force Nc2 applied by the drive unit 206a of the branch compensation valve 206 is determined. The first command value 11 of the control force NC2 obtained in the function block 313 and the second command value 11 of the control force Nc2 obtained in the addition block 317
The command value i2+13 is judged to be large or small in the minimum value selection block 3118, 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 turns an ON signal into a switch block when there are detection signals for both pilot pressure a1 or a2 and pilot pressure b1 or b2. 320, and outputs an OFF signal to the switch block 320 at other times. Switch block 320 connects AND block 319 to O
When the FF signal is output, it is held at the position shown, and the first command value 11 obtained in function block 313 is selected, and when the ON signal is output from AND block 319, the minimum value selected in block 318 is value, ie the first command value 11 or the second command value i2 +i3. As a result, when one of the pilot valves 210, 211 is operated, that is, when turning or operating the boom alone,
The first command value 11 is selected, and the pilot valves 210, 2
11 is operated, that is, in a combined operation of swing and boom, the first command value 11 and the second command value i2
The minimum value of +i3 is selected.

The control force signal 11 as the command value of the control force NC1 for the shunt compensating valve 205, obtained in the function block 313, passes through the amplification block 321, becomes the operation command signal S21, and is output to the electromagnetic proportional pressure reducing valve 216.

Further, the first command value 11 or the second command value i2+t3 selected by the switch block 320 is applied to the amplification block 3
22 and then the electromagnetic proportional pressure reducing valve 2 as an operation command signal 322.
17.

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

For example, when operating the boom pilot valve 211 to drive the flow rate control valve 204 to operate the boom independently,
Main pump 200 discharge pressure PS and boom cylinder 202
The differential pressure ΔPLS with respect to the load pressure is detected by the differential pressure detector 225, and in the controller 307, the function block 3
10, the corresponding target discharge amount QO is calculated, and as described above, the operation command signal S1 is output to the electromagnetic control valve 302b of the discharge amount control device 302, and the discharge amount is adjusted so that the differential pressure ΔPLS matches the target differential pressure ΔP LSO. is controlled.

At this time, in the function block 313, the differential pressure Δ
The control force signal 11 corresponding to PLS is determined as the first command value of the control force Nc2 of the shunt compensation valve 206, and
Only the pilot valve 211 is operated and the AND block 32
Since an OFF signal is output from 0, the first command value 11 is selected in the switch block 320, and this is output as the operation command signal 322 to the electromagnetic proportional pressure reducing valve 217. The spring 21 is attached to the shunt compensating valve 206.
The control force N corresponding to the control force signal 11 is opposite to the force f of 3.
c2 acts, and the second control force fi1 in the valve opening direction is applied to the branch compensating valve 206. Here, the control force signal 11, i.e., ilo, when the differential pressure ΔPLS is at the target differential pressure ΔP LSO is such that the corresponding control force Nc2 is equal to fO, which was explained with reference to FIG. 4A in the first embodiment. Since they are set to match, the branch compensation valve 206 and the flow control valve 204
Since the differential pressure across the front and rear of the
The boom cylinder 202 is supplied with a flow rate according to the opening degree of the flow control valve 204 . At this time, the operation command signal S21 corresponding to the control force signal 11 is simultaneously output to the electromagnetic proportional pressure reducing valve 216, and the branch compensation valve 205 similarly operates to maintain a predetermined differential pressure.

Even when the swing motor 201 is driven by a single swing operation, the operations of the branch compensating valves 205 and 206 are substantially the same as in the case of the above-described boom single operation.

When operating the pilot valves 210 and 211 simultaneously to perform a combined operation of turning and raising the boom, for example, the operator first operates the selection device 306 to output the corresponding command signal r, and then the function blocks 314 and 3 of the controller 307
Adjust 15 characteristics.

That is, the target discharge pressure Pso of the main pump 200 is set to a value suitable for the combined operation of swinging and boom raising.4 Specifically, in this combined operation, since the rotating body driven by the swinging motor 201 is an inertial load, the turning The motor 201 serves as an actuator on the high load pressure side, and its load pressure normally rises to the relief pressure set by the relief valve 303. From this, the target discharge pressure Pso is determined by adding the relief pressure of the swing motor 2゜1 to the load sensing compensation differential pressure Δ.
The pressure is set to be lower than the pressure obtained by adding p tso and higher than the pressure obtained by adding the differential pressure ΔP LSO to the load pressure of the boom cylinder 202.

Next, the pilot valves 210 and 211 are operated to open the flow control valves 203 and 204, and a combined operation of turning and raising the boom is started. At this time, the discharge pressure p of the main pump 200
s increases due to the load sensing control of the discharge amount control device 302, and in the process, when the discharge pressure Ps becomes larger than the target discharge pressure Pso, the function block 3
14, a relatively small control force signal 12 corresponding to the discharge pressure ps is determined, and at the same time, a relatively small control force signal i3 corresponding to the discharge pressure is determined in the function block 315 and the integration block 316. , the relatively small addition value i in the addition block 317
2 +i3 is calculated.

On the other hand, since the main pump 200 is under load sensing control at this time, the differential pressure ΔP[S is the target differential pressure ΔP LS
near the function block 31 of the controller 307
3, the control force signal 11 corresponding to the differential pressure Δp tso is determined.

Here, the functional relationship of block 313 and block 314.
The functional relationship of 315 is that the discharge pressure Ps is the target discharge pressure Ps
The added value i2+i3 when it is near o and the differential pressure ΔP[S
Control force signal 1 when is near the target differential pressure ΔP LSO
1 is almost equal to 1. As a result, the additional value i2 + i3 when the discharge pressure Ps is about to exceed the target discharge pressure pso is
Control force signal 1 when S is near target differential pressure ΔP [SO
1, il>i2+i3, and the minimum value selection block 318 selects the added value i2 +i3, that is, the second command value.

In this case, both pilot valves 210 and 211 are being operated, so the AND block 319 outputs
The N signal is output and the switch block 320 is switched to a position that selects the output of the minimum value block 318. Therefore, in the switch block 320, the second
The command value 12+i3 is selected, and this is the operation command signal 3.
22 to the electromagnetic proportional pressure reducing valve 217. Also,
An operation command signal 321 corresponding to the control force signal 11 is output to the electromagnetic proportional pressure reducing valve 216 .

In this way, the operation command signal 321. As a result of S22 being output, the second control force N in the valve opening direction is applied to the branch compensating valve 205.
f-il is applied as c1, and f-(i2+i3
) is given. Here f(i2±i3)>f-il
It is. Therefore, at the start of the combined operation of turning and raising the boom, the boom cylinder 202, which is on the low load pressure side,
The degree to which the shunt compensation valve 206 is throttled is reduced, and the normal control force Nc2=i1 is applied to the boom cylinder 204.
A larger flow rate is supplied than if . As a result, the increase in the discharge pressure of the main pump 200 is suppressed,
The discharge pressure will be 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 more than pso, so the flow rate of the pressure oil supplied to the swing motor 201 increases due to the swing. The load pressure is lower than when it rises to the relief pressure, and the swing motor 201 is driven at an appropriate speed without pressure oil being relieved. This makes it possible to perform a combined operation of turning and raising the boom, where the boom raising speed is high and the turning speed is relatively slow, and energy loss during turning acceleration can be reduced.

During the combined operation of swinging and boom raising as described above, when the swing is accelerated and reaches a steady speed, the swing motor 20
The load pressure of the main pump 200 decreases, and the discharge pressure of the main pump 200, which is controlled by load sensing, also decreases and becomes equal to or less than the target discharge amount Pso. Discharge pressure is target discharge amount Ps
o or less, the values of the control force signal 12 obtained in the function block 314 and the control force signal i3 obtained in the blocks 314 and 316 become large, and the second command value 12-1-f3 obtained in the addition block 318 also increases. It becomes relatively large, and the functional relationship of block 313 and block 31 described above
From the setting relationship of the functional relationship of 4,415, il < i2
+i3. Therefore, the first command ti i 1 is selected in the minimum value selection block 318, and the operation command signal S22 corresponding to the first command value 11 is output to the electromagnetic proportional pressure reducing valve 217.

As a result, the conventional f-il is applied to the branch flow compensation valve 206 as the second control force Nc2 in the valve opening direction, and at this time, the branch flow compensation valve 205 is also applied in the same valve opening direction. A second control force f11 is applied. Thereby, the differential pressures across the flow control valves 203 and 204 are controlled to be equal, and the swing motor 201 and the boom cylinder 202 are supplied with flow rates as requested by the pilot valves 210 and 211. That is, the flow rate of the pressure oil supplied to the swing motor 201 increases, and the desired swing speed can be obtained. In this way, after the turning acceleration, it is possible to realize a complex operation intended by an operator with a relatively fast turning speed.

As described above, in this embodiment, the boom cylinder 202 is an actuator that drives a load with small inertia.
main pump 200 by controlling the flow rate supplied to
Since the drive pressure of the swing motor 201, which is an actuator that drives a load with large inertia, is controlled by arbitrarily controlling the discharge pressure of the speed can be fast and the turning speed can be relatively slow, improving operability, reducing energy loss during complex operations, and enabling economical operation.

Furthermore, according to this embodiment, the characteristics of the function blocks 314 and 315 can be changed as appropriate by operating the selection device 306, and the target discharge pressure pso of the main pump 200 can be changed, so that the matching of the swing and boom raising can be adjusted as appropriate. Can be set.

In the above embodiment, in order to achieve both control responsiveness and stability, a function of the proportional control method is used as a means for obtaining a control force signal for controlling the discharge pressure Ps to be maintained at the target value pso in the controller 307. Block 3
14 and the integral control type function blocks 315 and 316 are used, but it will be obvious that either one may be used to determine the control force signal.

11 Pentagram U The sixth embodiment of the present invention will be explained with reference to FIGS. 39 to 44. In FIG. 0, the fourth embodiment shown in FIG. 27 and the fifth embodiment shown in FIG. Components equivalent to those in the example are given the same reference numerals.

In FIG. 39, the hydraulic drive system of this embodiment basically has the same configuration as the fourth embodiment shown in FIG. 27, and the explanation of that part will be omitted. However, the differential pressure ΔPLS between the discharge pressure ps of the main pump 200 and the maximum load pressure Pa1aX
The output signal from the differential pressure detector 225 that detects the is expressed as Edp. In addition, as in the fifth embodiment shown in FIG. 33, the discharge pipe 207 of the main pump 200 is configured such that when the pressure oil from the main pump 200 reaches the relief setting pressure, it flows out into the tank, and the pump discharge pressure is adjusted accordingly. A relief valve 300 is provided to prevent the pressure from becoming higher than the set pressure, and pressure oil from the main pump 200 is supplied to the high pressure side of the swing motor 201 and the boom cylinder 202 at a load pressure (hereinafter referred to as maximum load pressure P). An unload valve (not shown) is provided which allows the liquid to flow out into the tank when the pressure reaches the sum of the unload setting pressure (referred to as anax) and the unload setting pressure, and prevents the pressure from exceeding the specified pressure.

Furthermore, the main pump 200 is provided with a displacement detector 223 that detects its displacement, and the displacement detector 223 outputs a signal Eθ corresponding to the detected displacement. The discharge rate of the main pump 200 is controlled by a discharge rate control device 400 using a load sensing control method, which corresponds to the discharge rate control device 302 of the fifth embodiment.
00 consists of a tilting drive device 400a that drives the swash plate 200a of the main pump 200 to increase/decrease the displacement volume, and an electromagnetic proportional pressure reducing valve 400b that outputs control pressure to this tilting drive device and adjusts its displacement. There is.

Then, it is detected that pilot pressure is applied to pilot lines 401a and 401b that lead pilot pressure from a swing pilot valve (not shown) to the driving part of the flow rate control valve 203, respectively, and signals E402 and E40 are sent.
Operation detectors 402 and 403 that output 3 are provided. The rotation motor 20 is operated by an operator.
A selection device 406 is provided for selecting and setting the rate of increase in the flow rate of the pressure oil supplied to the pump 1, and the selection device 406 outputs a signal ES according to the setting at that time.

Signal Edp from differential pressure detector 225, i-operation detector 402
．． The signals E4Q2 and E403 from 403, the signal ES from the selection device 406, and the signal Eθ from the displacement detector 223 are input to the controller 407, where after predetermined calculations are performed, the signals are sent to the electromagnetic proportional pressure reducing valves 216 and 217. While outputting operation command signals E216 and E217,
An operation command signal E400 is output to the electromagnetic proportional pressure reducing valve 400b of the discharge amount control device 400.

In this embodiment, the selection device 1F 406, as shown in FIG. 40, consists of a voltage setting device including a variable resistor 408, and when the operator changes the position of the movable contact, a corresponding voltage level is set. Ru. This voltage value is the signal B
The signal ES is taken into the controller 407 as a
Sent to PU. As shown in the flowchart in FIG. 41, the CPU reads the A/D conversion value of this signal Es in step S1, sets ΔE=A/D conversion value in step S2, and issues an operation command to the electromagnetic proportional pressure reducing valve 216. Amount of change Δ per cycle of signal E216
Find E. This amount of change ΔE is used in the controller 407 to obtain the operation command signal E216.

The six flowcharts shown in FIG. 42 are the operation command signals E216 for the electromagnetic proportional pressure reducing valves 216, 217,
This shows the calculation procedure of E217, and the method of obtaining the operation command signal E400 for the electromagnetic proportional pressure reducing valve 400b of the discharge amount control device 400 is the same as the method of obtaining the operation command signal S1 in the fifth embodiment shown in FIG. Since they are essentially the same,
The explanation will be omitted.

First, in step S10, the signal E(lΩ, , E40
2E403, Load ES 6 Then step S
11, a basic drive signal EHL for the electromagnetic proportional pressure reducing valves 216 and 217 is calculated from the differential pressure signal Edp and a pre-stored functional relationship. This basic drive signal E11 [ is caused by the fact that the main pump 200 is load sensing controlled by the discharge amount control device 400, and even if the discharge amount of the main pump 200 reaches the maximum, the differential pressure ΔP[S does not reach the target differential pressure ΔP LSO. Sometimes, the drive part 205 of the shunt compensation valve 205° 206
c, the control force Ncl, Na3 applied by 206c is increased, and the second control force f-Ncl, Na3 in the valve opening direction is decreased, that is, the target value of the differential pressure across the flow control valves 203 and 204 is decreased. , the flow rate of the pressure oil supplied to each actuator 201, 202 is distributed according to the opening ratio of the flow rate control valves 203, 204, that is, the ratio of the required flow rate, although the increase in the absolute amount is suppressed. be. FIG. 43 shows the functional relationship between the differential pressure ΔPLS and the drive signal Ell for determining the basic drive signal EHL. This functional relationship is
Differential pressure ΔPLS and control force signal 11 shown in FIG. 36 described above
The relationship is essentially the same as that of

Next, in step S12, the operation command signal E40
2 or E403 is input, and if it is not input, the process proceeds to step S13, and the drive signal EH of the electromagnetic proportional pressure reducing valve 216 is set as EH-EH14AX, where EHHAX is the drive signal EH At this time, the control force NC1 of the drive unit 205c becomes the maximum, and the branch compensating valve 205 is held at the fully open position against the force f of the spring 212. Operation command signal E402 or E40
If 3 is input, the process advances to step S14 and EH
Determine whether L<E H-1-ΔE. That is, the drive signal EHL is determined from the drive signal EH-1 of the electromagnetic proportional pressure reducing valve 216 obtained in the previous control cycle by the selection device 4 described above.
It is determined whether the value is smaller than the value obtained by subtracting the amount of change ΔE set by 06. Here, EIIL is E H-
If it is determined that it is smaller than 1-ΔE, the process proceeds to step S15, where EH=Elf-1-ΔE is set, and E
If it is determined that El
=EHL. That is, the drive signal El+ is determined so that the maximum rate of change of the drive signal El+ matches ΔE.

Subsequently, in step S17, El-1=EH is set, in step 318, the drive signal EH is output as the operation command signal E216, and in step S19, the basic drive signal E[1F is output as the operation command signal E217. As a result, the control force NC1 applied by the drive section 205C of the branch flow compensation valve 205 is controlled to match the basic drive signal E11[, and its rate of change is limited to ΔE or less.6 The drive section of the branch flow compensation valve 206 As before, the control force Nc2 applied by the motor 206C is controlled to match the basic drive signal EIIL4.

Next, the operation of this embodiment configured as above will be explained.

First, when no flow control valve is operated and the actuator is not driven, the operation detection signal E402 or E403 is not input to the controller 407, so in step 812 of the flowchart shown in FIG. A determination of NO is made, and step S1
3, the drive signal EH of the electromagnetic proportional valve pressure valve 216 is set to the maximum value EHHAX. For this reason, the shunt compensation valve 20
5 is held in the fully closed position. On the other hand, the electromagnetic proportional pressure reducing valve 21
7, the basic drive signal E11F is the operation command signal E
217, but at this time, the discharge pressure Ps of the main pump 200 corresponding to the set pressure (>ΔPLSO) is secured by the unload valve (not shown), so the function shown in FIG. 43 is set in step S11. A relatively small basic drive signal EHL is required from the relationship,
The flow compensating valve 206 is held in the fully open position by the force f of the spring 213.

When operating a boom pilot valve (not shown) to drive the flow rate control valve 204 and operating the boom independently,
Discharge pressure Ps of main pump 200 and boom cylinder 202
The differential pressure ΔPLS with respect to the load pressure is detected by the differential pressure detector 225, and the controller 407 calculates an operation command signal E400 for keeping the differential pressure ΔP[S constant. The discharge amount of the main pump 200 is controlled accordingly.

At this time, the controller 407 sends an operation command signal E216 to the electromagnetic proportional pressure reducing valve 216,217.
, E217 are calculated. Here, in this case, since the swing flow control valve 203 is not driven, the operation detection signal E402 or E403 is not input, and the electromagnetic proportional valve pressure valve 216 is The drive signal EH is set to a maximum of 1iaEHMAX, and the shunt compensation valve 205 is held at the fully open position. On the other hand, for the boom shunt compensation valve 206, in step Sll, a basic drive signal E corresponding to the differential pressure ΔPLS near the target differential pressure ΔP LSO is calculated from the functional relationship shown in FIG. The drive signal EHL is output to the electromagnetic proportional pressure reducing valve 217 as an operation command signal E217. Here, the fourth
The functional relationship of No. 3 is substantially the same as the functional relationship shown in FIG. 36 described above. Therefore, the branch compensation valve 206 is held at the fully open position by the second control force of f-NO2 against the first control force in the valve closing direction based on the pressure difference across the flow control valve 204, and the boom cylinder 202 Nara JiiftlJ Goben 2
A flow rate corresponding to the opening degree of 04 is supplied.

Independent operation of swing motor 201 or flow control valve 203
, 204 at the same time to perform a combined operation of turning and raising the boom, for example, the operator first operates the selection device 406 to output the flow rate increase speed signal Es, and as described above, the operator operates the selection device 406 to output the flow rate increase speed signal Es. Set the amount of change ΔE per cycle. Specifically, if the turning acceleration is desired to be gradual, the change amount ΔE is set to a small value, and if it is desired to be accelerated, the change amount ΔE is set to a large value.

Next, the flow rate control valve 203 is driven alone or both the flow rate control valve 203 and the flow rate control valve 204 are driven simultaneously to start a single swing operation or a combined swing and boom raising operation. At this time, the discharge pressure Ps of the main pump 200 is set to a differential pressure Δ by load sensing control of the discharge amount control device 400.
Rise while maintaining p tso.

At the same time, the controller 407 also sends an operation command signal E2 to the electromagnetic proportional pressure reducing valves 216 and 217.
1G and E217 are calculated. Here, in this case, since the swing flow control valve 203 is driven and the operation detection signal E402 or E403 is input, a YES determination is made in step 312 shown in FIG.
Drive signal EH is obtained by calculations in steps 314 to 816. That is, a drive signal E)I is obtained that sets the basic drive signal EHL to a target value and limits the rate of change to ΔE or less. Then, this drive signal Ell is outputted to the electromagnetic proportional valve 216 as the operation command signal E21G, and the shunt compensation valve 205
It gradually starts to open from the fully closed position at a speed corresponding to the amount of change ΔE, and correspondingly, pressure oil is supplied to the swing motor 201 at a rate of increase in flow rate corresponding to the amount of change ΔE.

In this way, the swing motor 201 is driven with an acceleration corresponding to the amount of change ΔE.

Here, the relationship between the time t during the turning operation, the drive signal EH, and the flow rate increase speed signal Es is shown in FIG.

After the start of the turn, the drive signal EH decreases at a gradient corresponding to the amount of change ΔE. The gradient becomes larger as the flow rate increase rate signal ES, that is, the amount of change ΔE becomes larger. 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 driving acceleration of the swing motor 201.

On the other hand, for the boom branch flow compensation valve 206, in step S11, as in the case of independent operation of the boom, basic drive corresponding to the differential pressure ΔPLS near the target differential pressure ΔPLSQ is performed based on the functional relationship shown in FIG. A signal EI11 is calculated, and this basic drive signal EHI is outputted to the electromagnetic ratio S pressure reducing valve 217 as an operation command signal E217. That is, the shunt compensation valve 206 receives the signal E217 in opposition to the force of the spring 213.
A control force Nc2 corresponding to is applied in the valve opening direction. As a result, in the case of a single turning operation, the flow branch compensation valve 206 is f
- It is held in the fully open position by the second control force of Na3. In addition, in the case of a combined operation of turning and raising the boom, the boom cylinder 202 becomes the actuator on the low load pressure side.
The shunt compensation valve 206 adjusts the differential pressure across the flow rate control valve 204 to f-
It is narrowed down to hold at NC2.

In the case of a combined operation of turning and raising the boom, 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 PLS decreases, the value of the basic drive signal E11 calculated in step S11 of FIG. 42 increases,
The shunt compensation valves 205 and 206 are actuators 201 and 2
The absolute amount of pressure oil supplied to 02 is limited, and the flow rate distribution is controlled to suit.

After the start of the swing operation, when the swing reaches a speed corresponding to the opening degree (required flow rate) of the flow rate control valve 203, the flow rate compensation valve 205
The control force NC1 applied by the drive unit 205c in step S
The value reaches a value corresponding to the drive signal E11[ calculated in step S11, and EII=Eln is always calculated in step S16. Therefore, at this point, the second control force f-Nc1.
f-Na3 is equal, and in the case of a combined operation of turning and boom raising, the respective actuators 201 and 202
A flow rate proportional to the adjustment of the flow rate control valves 203 and 204 is supplied to the flow rate control valves 203 and 204, and a combined operation of turning and boom raising can be performed at the required speed ratio.

As described above, in this embodiment, the rate of increase in the flow rate of the pressure oil supplied to the swing motor 201 can be arbitrarily set at the start of the swing operation, so that in the combined operation of swing and boom raising, the combined At the start of operation, the flow rate ratio of pressure oil supplied to both actuators can be arbitrarily changed,
Combined operations can be performed at the optimal speed ratio for the task.

Furthermore, since the rate of increase in the flow rate of the pressure oil supplied to the swing motor 201 can be set arbitrarily at the start of the swing operation, a sudden increase in the swing load pressure can be suppressed, and the pressure that is throttled and discarded by the swing relief valve can be set arbitrarily. Oil is reduced and energy loss can be reduced. Furthermore, if the flow rate increase rate is set relatively low, the drive pressure of the swing motor can be kept below the relief pressure, which makes it possible to further reduce energy loss and also reduce the discharge pressure of the main pump 200. Therefore, when the main pump 200 is subjected to horsepower limit control # (input torque it/J limit control), the discharge amount can be increased in accordance with the reduction in discharge pressure, increasing the amount of pressure oil supplied to the boom cylinder, Drive speed can be increased.

Sixth Example 1'/Modification A first modification of the sixth embodiment will be explained 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
It consists of a switching device including a movable contactor 409 for ~D. Contacts A to C are connected to the input terminal D11°Di2 of the CPU in the controller 407A. Di 3, and input terminal Di 1° Di 2. Di 3 is connected to the power supply via resistors 410a, 410b, 410c. With such a configuration, the movable contact 409
For example, when the input terminal Di 1 is in a position where it contacts the contact C as shown in the figure, the input terminal Di 1 is grounded and the voltage becomes 0, and the other input terminals Di 2 . Di 3 is maintained in a state where the power supply voltage is applied.

In the controller 407A, input terminals D1 and Di
2. The flow rate increase rate is set as shown in FIG. 46 according to the voltage state of Di3. In step 820, it is determined whether the voltage at the input terminal Di3 is 0, and if it is 0, in step S21, the change 18E per cycle of the operation command signal E21G for the electromagnetic proportional pressure reducing valve 216 is determined.
is set to a pre-stored value ΔEA. If the voltage at the input terminal Di3 is not O, the process proceeds to step 322, where it is determined whether the voltage at the input terminal D2 is 0, and if it is O, the amount of change ΔE is set to a pre-stored value ΔEB in step S23. If the voltage at the input terminal Di2 is not 0, the process proceeds to step 324, where it is determined whether the voltage at the input terminal D1 is O, and if it is 0, the amount of change Δ is determined at step S25.
Set E to a pre-stored value ΔEC. Finally, if the voltage at the input terminal Di1 is not 0, the process proceeds to step S26, where the amount of change ΔE is set to a pre-stored value ΔED.

By switching the position of the movable contactor 409 as described above, the amount of change ΔE can be set according to the position.

Next, the second modification of the sixth embodiment is shown in FIGS. 39 and 4.
This will be explained with reference to FIG. In 1) in FIG. 47, the same steps as those shown in FIG. It was designed to be done only at certain times.

In the hydraulic drive system of this embodiment, as shown by imaginary lines in FIG. An operation detector 405 is further provided which detects that pilot pressure is applied to the pilot line 404a on the side corresponding to the increase and outputs a signal B 405, and a signal E405 is sent to the controller 407.

In the controller 407, in step S30 shown in FIG.
3. In addition to the BS, the detection signal E 405 from the operation detector 405 is read, and in addition to the judgment of step S12, the operation detection signal E4 is read in in step S31.
05 is input, and only when this is satisfied does the process proceed to steps S14 to S16, and a drive signal El+ is calculated that sets the basic drive signal EHL as a target value and limits the amount of change to ΔE or less. .

According to the present embodiment, it is possible to control the rate of increase in the flow rate of the pressure oil supplied to the swing motor only during the combined operation of swing and boom raising, and to control the acceleration of the swing.

〔Effect of the invention〕

In the hydraulic drive device for construction machinery of the present invention, configured as described above, individual pressure compensation characteristics are given to the first and second flow compensation valves, and the first and second actuators are simultaneously driven. During combined operations, it is possible to provide the optimum flow splitting ratio according to the type of actuator and improve operability and/or work efficiency.

[Brief explanation of drawings]

FIG. 1 is a circuit diagram showing the entire hydraulic drive system for construction machinery according to the first embodiment of the present invention, FIG. 2 is a schematic diagram showing the configuration of a controller, and FIG. 3 is a schematic diagram showing the configuration of a controller. FIG. 4 is a functional block diagram showing the contents of the calculation;
Figure A is a diagram showing the functional relationship between the differential pressure ΔPIS and the value of the control force Pc1 to be applied to the shunt compensation valve related to the swing motor, and Figure 4B is a diagram showing the functional relationship between the differential pressure ΔPLS and the value of the control force Pc1 to be applied to the shunt compensation valve related to the travel motor. Control force F c2. to be applied to the valve. F
Fig. 4C is a diagram showing the functional relationship with the value of c3;
FIG. 4D is a diagram showing the functional relationship between the differential pressure ΔP[S and the value of the control force Fc4 to be applied to the shunt compensation valve related to the boom cylinder, and FIG. $IJ power to be given to the shunt compensation valve related to c5. FIG. 5 is a diagram showing the functional relationships shown in FIGS. 4A to 4D, and FIG. 6 is a diagram showing the functional relationship between the oil temperature Th and the correction coefficient. FIG. 7 is a side view of a hydraulic excavator to which the hydraulic drive system of this embodiment is applied, FIG. 8 is a top view of the hydraulic excavator, and FIG. 12 are diagrams showing four modified examples of the functional relationship between the differential pressure ΔP[S and the value of the control force Fct to be applied to the shunt compensation valve related to the swing motor, and FIGS. FIG. 14 is a diagram showing two modified examples of the functional relationship between the differential pressure ΔP[S and the value of the control force Fc2°Fe2 to be applied to the shunt compensation valve related to the travel motor, and FIG. FIG. 16 is a circuit diagram showing the entire hydraulic drive device according to the second embodiment of the invention, FIG. 16 is a functional block diagram showing the contents of calculations performed by the controller, and FIG. 17 is a circuit diagram showing the contents of the calculations performed by the controller. 18 is a circuit diagram showing the entire hydraulic drive device according to an example, FIG. 18 is a functional block diagram showing the contents of calculations performed by the controller, and FIG.
FIG. 20 is a diagram showing a plurality of functional relationships between control forces Fcl to Fc6, and FIG.
Figure 1 is a diagram showing the relationship between the differential pressure across the flow rate control valve for the boom and the supply flow rate when performing the same complex operation, and Figure 22
The figure shows the relationship between the differential pressure across the swing flow control valve and the supply flow rate when performing the same combined operation. Figure 23 shows the relationship between the supply flow rate and the pressure difference between the front and rear flow control valves when performing the same combined operation. FIG. 24 is a diagram summarizing the functional relationships selected when carrying out a combined operation of the arm and baguette intended for shaping the ground, etc. without making it flat. FIG. 25 is a functional block diagram showing the contents of calculations performed by the controller in a modification of the third embodiment, and FIG. 26 is a diagram showing an embodiment of the control pressure generation circuit. FIG. 27 is a circuit diagram showing a hydraulic drive device according to a fourth embodiment of the present invention, FIG. 28 is a schematic diagram showing the configuration of a discharge amount control device, and FIG. 29 is a circuit diagram showing a hydraulic drive device according to a fourth embodiment of the present invention. FIG. 30 is a functional block diagram showing the content of calculations performed by the controller; FIG. 30 is a diagram showing the relationship between discharge pressure and input limit target discharge amount;
The figure shows a limiter function for determining the basic correction [Qns from the intermediate 1iiQ'ns, and FIG. 32 shows the basic correction value Qns and the operation command signal S21. FIG. 33 is a circuit diagram showing the hydraulic drive device according to the fifth embodiment of the present invention, and FIG. 34 is a functional block diagram showing the contents of calculations performed by the controller. 35 is a diagram showing the 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 force signal 11, and FIG. The figure shows the functional relationship between the discharge pressure Ps, the control force signal 12, and the command signal r, and FIG. 38 shows the functional relationship between the discharge pressure PS, the rate of change i3 of the control force signal i3, and the 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 the configuration of a selection device, and FIG. 41 is a diagram showing the configuration of a selection device. Flow 5f yard showing the procedure for calculating the amount of change ΔE according to the operation of
FIG. 42 is a flow chart showing the calculation contents performed by the controller, and FIG. 43 is a diagram showing the functional relationship between the differential pressure ΔP[S and the basic drive signal E11].
The figure is a diagram showing the relationship between the time t at the start of the turning operation, the drive signal EH, and the flow rate increase speed signal Es, and FIG.
FIG. 46 is a diagram showing the configuration of a selection device according to a first modification of the sixth embodiment, and FIG.
The figure is a flowchart showing the calculation contents performed by the controller in the second modification of the sixth embodiment. Explanation of symbols FIGS. 1 to 14 45-50... Spring (driving means) 35c-40c... Five drive parts (driving means)... Differential pressure detector (first means) 61... Controller (second means) 62a ~62f
... Solenoid proportional pressure reducing valve (control pressure generation means) 80 to 85 ... Function block (first calculation means) 90
~92...Delay element block (second calculation means) 60
...Oil temperature detector 86...Function block (sixth calculation means) 87-89
... Multiplication block (fourth calculation means) Fig. 15 45A to 50A ... Drive section (drive means) 63 ... Pilot pump (pilot pressure supply means) 64 ... Relief valve (pilot pressure supply means) )11
3... Pressure reducing valve (pilot pressure supply means) Figures 1 to 14 120... Selection device 'll (fourth means) 61B...
Controller (second means) 80B to 85B...Function block (fifth calculation means) 27th to 32nd figures 200...Main pump 201...Swivel motor (first actuator) 22
4...Discharge pressure detector (fifth means) 225...Differential pressure detector (first means) 205.206...Diversion compensation valve 229...Controller (second means) 240~ 24
2...Block (10th calculation means) 243...Block (11th calculation means) 258...Block (1st calculation means)
2 calculation means) 259 to 263...blocks (13th
306...Selection device (sixth means) 307...Controller (second means) 313...
Function block (sixth calculation means) 314, 315...
Function block (seventh calculation means) 316... Integration block (seventh calculation means) 317... Addition block (seventh calculation means) 318... Minimum value selection block (seventh calculation means)
402, 403, 405... Operation detector (seventh means) 406... Selection device (eighth means) 407... Controller (second means) ) Figure 2

Claims (15)

[Claims] (1) A hydraulic pump, at least first and second hydraulic actuators driven by pressure oil supplied from the hydraulic pump, and controlling the flow of the pressure oil supplied to these first and second actuators, respectively. first and second flow control valves that control the first and second flow rate control valves, and first and second flow branch compensation valves that control the first differential pressure that occurs between the inlets and outlets of the first and second flow control valves, respectively; Discharge amount control means for controlling the flow rate of pressure oil discharged from the hydraulic pump in response to a second differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second actuators. , the first and second flow compensation valves each include a drive means that applies a control force based on the second pressure difference to the corresponding flow compensation valve and sets a target value of the first pressure difference. A hydraulic drive device for construction machinery comprising: a first means for determining the second differential pressure from the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second actuators; and at least 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 branch compensating valves based on the second differential pressure determined by the second flow compensation valve; first and second control pressure generating means provided corresponding to the first and second flow compensating valves, respectively;
The first and second control pressures each generate a control pressure according to the individual value determined by the second means, and output this to the drive means of the first and second flow compensation valves, respectively. A hydraulic drive device characterized by having a generating means. (2) In the hydraulic drive system for construction machinery according to claim 1, the second means corresponds to the second differential pressure determined by the first means and the first and second flow compensation valves. A hydraulic drive device comprising a first calculating means for calculating values of first and second control forces corresponding to the second differential pressure from first and second functions set in advance. . (3) The hydraulic drive device for construction machinery according to claim 2, wherein the first actuator is an actuator that drives an inertial load, and the second actuator is an actuator that drives a normal load. The second function is configured so 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 second differential pressure and the first and second differential pressures. A hydraulic drive device characterized in that a relationship with a control force value is determined. (4) The hydraulic drive device for construction machinery according to claim 2, wherein the first actuator is an actuator that drives an inertial load, and the second actuator is an actuator that drives a normal load. The first function related to the actuator is configured such that when the second differential pressure increases beyond a predetermined value, an increase in the target value of the first differential pressure is suppressed. A hydraulic drive device characterized in that a relationship between the value of the control force and the value of the control force is determined. (5) In the hydraulic drive system for construction machinery according to claim 2, wherein the first and second actuators are travel actuators, both the first and second functions are based on the first differential pressure. A hydraulic drive device, characterized in that a relationship between the second differential pressure and the values of the first and second control forces is determined such that the target value is greater than the second differential pressure. (6) The hydraulic drive device for construction machinery according to claim 2, wherein the first actuator is one of the actuators for traveling, and the second actuator is an actuator for excavation work. is the first
A relatively large time delay is given to a change in the value of the first control force obtained from the function, and a relatively small time delay is given to a change in the value of the second control force obtained from the second function. A hydraulic drive device further comprising second calculation means for providing a time delay. (7) In the hydraulic drive device for construction machinery according to claim 2, wherein the first actuator is a hydraulic motor and the second actuator is a hydraulic cylinder, the temperature of the pressure oil discharged from the hydraulic pump is detected. The second means further comprises a third calculation means for calculating a temperature correction coefficient from the temperature of the pressure oil detected by the third means and a preset third function. , further comprising fourth calculation 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. Hydraulic drive device. (8) The hydraulic drive device for construction machinery according to claim 1, which is operated from the outside and outputs a selection command signal according to the type of work or the content of the work to be performed by driving the first and second actuators. further comprising a fourth means;
The second means calculates the second differential pressure obtained by the first means, fourth and fifth functions preset corresponding to the first and second flow compensation valves, respectively, and the second differential pressure determined by the first means. A hydraulic drive device comprising a fifth calculation means for calculating the values of the third and fourth control forces from the selection command signal output from the fourth means. (9) In the hydraulic drive device for construction machinery according to claim 8, the fifth calculation means includes a plurality of functions having different characteristics as the fourth and fifth functions, and the fourth
one of the plurality of functions is selected in response to the selection command signal output from the means, and the second difference is calculated from the second differential pressure obtained by the first means and the selected function. A hydraulic drive device characterized in that values of third and fourth control forces corresponding to pressure are determined. (10) Claim 1, wherein the first actuator is an actuator that drives an inertial load, and the second actuator is an actuator that drives a normal load.
The hydraulic drive system for construction machinery described above further includes a fifth means for detecting the discharge pressure of the hydraulic pump, and the second means is configured to detect a second differential pressure determined by the first means and a second differential pressure determined in advance by the first means. A fifth control force value corresponding to the second differential pressure is determined from the set sixth function, and this value is set as a value of the control force to be applied by the driving means of the first branch compensation valve. The calculation means of No. 6 calculates a value of a sixth control force for maintaining the discharge pressure at a predetermined value from the discharge pressure detected by the fifth means and a preset seventh function, and the fifth control Of the force and the sixth control force, the one that increases the target value of the first differential pressure is selected as the second control force.
A hydraulic drive device characterized in that it has a seventh calculating means for determining the value of the control force to be applied by the driving means of the branch compensating valve. (11) The hydraulic drive device for construction machinery according to claim 10, further comprising a sixth means that is operated from the outside and outputs a selection command signal related to the predetermined value of the discharge pressure, and the seventh calculation means is selected by the selection command signal.
A hydraulic drive device characterized in that the predetermined value of the discharge pressure can be changed by changing the characteristics of the function. (12) Claim 1, wherein the first actuator is an actuator that drives an inertial load, and the second actuator is an actuator that drives a normal load.
In the hydraulic drive system for construction machinery described above, a seventh means for detecting driving of the first actuator;
and an eighth means for setting a rate of increase in the flow rate of the pressure oil supplied through the branch flow compensation valve, the second means comprising:
The second differential pressure obtained by the first means and the eighth differential pressure set in advance
and an eighth calculation means for determining the value of the seventh control force corresponding to the second differential pressure from the function of and the seventh
when the start of driving of the first actuator is detected by the means, an eighth control that changes the value of the seventh control force as a target value at a rate equal to or less than the amount of change corresponding to the rate of increase in the flow rate; The value of the force is determined, and this eighth control force is applied to the first control force.
A hydraulic drive device characterized in that it has a ninth calculating means for determining the value of the control force to be applied by the driving means of the branch compensating valve. (13) The hydraulic drive device for construction machinery according to claim 12, further comprising ninth means for detecting the drive of the second actuator, and the ninth calculating means is configured to detect the driving of the seventh actuator.
A hydraulic drive system for construction machinery, characterized in that the value of the eighth control force is determined when the start of driving of the first and second actuators is detected by the ninth means. (14) The hydraulic drive system for construction machinery according to claim 1, further comprising a tenth means for detecting the discharge pressure of the hydraulic pump, and the second means detects the discharge pressure determined by the first means. a tenth calculation means for calculating a differential pressure target discharge amount of the hydraulic pump to maintain the differential pressure constant from the differential pressure of the second one; and a preset input limit for the hydraulic pump based on the discharge pressure detected by the tenth means; an eleventh calculation means for calculating an input limit target discharge amount of the hydraulic pump from a function; a twelfth calculation means for calculating a deviation between the differential pressure target discharge amount and the input limit target discharge amount; When the input limited target discharge amount is selected as the discharge amount target value of the hydraulic pump from among the input limited target discharge amounts, based on the deviation of the target discharge amount,
13. Calculating individual values as control force values to be applied by the driving means of each of the first and second flow branch compensation valves.
A hydraulic drive device characterized by having a calculation means. (15) In the hydraulic drive device for construction machinery according to claim 1, the first-mentioned driving means is provided in the first and second branch compensation valves and urges each of the branch compensation valves in the valve opening direction. and pilot pressure supply means for directing a substantially constant common pilot pressure to the further drive means, said first-mentioned drive means comprising:
A hydraulic drive device, wherein each of the first and second branch compensation valves is disposed on a side that urges the valve in a valve closing direction.