JP3490507B2 - Hydraulic control device for construction machinery - Google Patents

Description

Description: BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control device for driving a plurality of hydraulic actuators by controlling the discharge flow rate of a variable displacement hydraulic pump by load sensing control, and more particularly to a control device for power generation. The present invention relates to a load sensing control type hydraulic control device for a construction machine having a hydraulic actuator that needs to be driven at a constant speed, such as a hydraulic motor. 2. Description of the Related Art A hydraulic motor for driving a generator for a lifting magnet mounted on a construction machine needs to be driven and controlled at a constant rotational speed so that the generator performs a power generation operation at a rated voltage. As a hydraulic pressure supply control device for controlling this type of hydraulic pressure supply, one disclosed in Japanese Utility Model Laid-Open No. 3-97474 is known. FIG. 6 shows the hydraulic supply control device. This hydraulic supply control device supplies the discharge oil discharged from the main hydraulic pump 1 to a hydraulic motor 3 for driving a generator 2 for a lifting magnet and a hydraulic cylinder device 4 for a working machine which is another hydraulic actuator. In addition, a flow path switching valve 5 and a flow rate priority valve 6 are provided in the middle of the supply path of the discharge oil. The flow path switching valve 5 is selectively supplied with pilot hydraulic pressure generated by a sub-hydraulic pump 9 by an electromagnetic opening / closing valve 8 controlled by an operation switch 7 to perform a switching operation. Supply and discharge pressure oil to
In the position, pressure oil is supplied to the control valve 10 of the hydraulic cylinder device 4 for the working machine. When the differential pressure across the throttle 11 becomes greater than the spring force of the return spring 6a, the flow priority valve 6 switches from position a to position b. When the flow path switching valve 5 is switched to the position a, the discharge flow rate of the main hydraulic pump 1 increases, the rotation speed of the hydraulic motor 3 becomes equal to or higher than a predetermined rotation speed, and the throttle 1
FIG. 7A shows that when the pressure difference before and after 1 becomes larger than the spring force.
7B, ie, the switching position for supplying the pressure oil only to the hydraulic motor 3 to the switching position in FIG. 7B, and supplies the excess hydraulic oil to the working machine hydraulic cylinder device 4. On the other hand, when the pressure difference to the hydraulic motor 3 decreases and the rotational speed of the hydraulic motor 3 decreases, so that the differential pressure across the throttle 11 becomes equal to or less than the spring force, the flow rate priority valve 6 is turned on as shown in FIG.
Returning from the switching position shown in FIG. 7B to the switching position shown in FIG.
To ensure the pressure oil flow rate to In FIG. 6, reference numeral 12 denotes a prime mover for driving the main hydraulic pump 1 and the sub hydraulic pump 9, reference numeral 14 denotes a lifting magnet (electromagnet), reference numeral 13 denotes a control panel for controlling energization of the lifting magnet 14.
Reference numerals 15 and 16 indicate relief valves, respectively. In the conventional apparatus shown in FIG. 6, if the load hydraulic pressure of the hydraulic cylinder device 4 for the working machine is higher than the load hydraulic pressure of the hydraulic motor 3, the excess hydraulic oil Flows to the side of the hydraulic motor 3 having a low load, as a result, the differential pressure across the throttle 11 further increases, and the flow priority valve 6 maintains the state of FIG. 7B. Therefore, the flow rate of the pressure oil to the hydraulic motor 3 increases, the rotation speed of the generator 2 cannot be maintained at the set rotation speed, and the attraction force of the lifting magnet 14 is not stable. In recent years, load sensing control for controlling the displacement of a variable displacement hydraulic pump so that the differential pressure between the pump pressure and the actuator load pressure becomes a reference value has been known. When installing a hydraulic motor that drives the generator of the lifting magnet, it is necessary to consider the flow rate fluctuation of the hydraulic motor due to the load pressure difference between the hydraulic cylinder device for the working machine and the hydraulic motor, assuming various working conditions. There is. SUMMARY OF THE INVENTION It is an object of the present invention to provide a load sensing hydraulic control device which maintains the flow rate of hydraulic oil to another hydraulic actuator such as a hydraulic motor at a predetermined constant value even when the load pressure of a hydraulic cylinder device for a work machine fluctuates. Another object of the present invention is to provide a hydraulic control device for a construction machine capable of keeping its output state constant. DESCRIPTION OF THE PREFERRED EMBODIMENTS The present invention will be described with reference to FIG. 1 showing an embodiment. The present invention relates to a variable displacement hydraulic pump 20 which is driven by a discharge oil from the hydraulic pump 20. First and second hydraulic actuators 4a, 3
And a control valve 22a provided between the hydraulic pump 20 and the first hydraulic actuator 4a to control the flow rate of the first hydraulic actuator 4a, and provided between the hydraulic pump 20 and the second hydraulic actuator 3 A flow compensating valve 40 for maintaining the flow rate of the second hydraulic actuator 3 constant by controlling the differential pressure across the throttle 41 to a predetermined value, and a displacement volume adjusting means for adjusting the displacement volume of the hydraulic pump 20 20a and a first hydraulic actuator 4a provided in the middle of a hydraulic supply path for the first hydraulic actuator 4a.
And a variable throttle valve 28 for controlling the flow rate supplied to the hydraulic pump 20 so that the differential pressure between the discharge pressure of the hydraulic pump 20 and the maximum value of the load pressure of the first and second hydraulic actuators 4a and 3 becomes a reference value. The displacement volume of the variable displacement hydraulic pump 20 is controlled by the volume adjusting means 20a, and when the displacement volume of the variable displacement hydraulic pump 20 is the maximum value and the differential pressure does not reach the reference value, the differential pressure is reduced to the flow rate compensating valve 40. The above-mentioned object is achieved by providing the control means 27 for reducing the opening of the variable throttle valve 28 until the predetermined value is reached. When the pressure of the second hydraulic actuator 3 is lower than the pressure of the first hydraulic actuator 4a, the flow rate of the second hydraulic actuator 3 is maintained at a predetermined constant value by the flow rate compensating valve 40. When the load pressure of the first hydraulic actuator 4a is lower than the load pressure of the second hydraulic actuator 3, the load pressure of the second hydraulic actuator 3 becomes the maximum load pressure, and the displacement volume adjusting means 20a outputs the second displacement.
The displacement (discharge capacity) of the hydraulic pump 20 is adjusted so that the differential pressure between the load pressure of the hydraulic actuator 3 and the discharge pressure of the hydraulic pump 20 becomes a predetermined reference value. When the discharge flow rate of the variable displacement hydraulic pump 20 is insufficient, that is, when the displacement volume of the variable displacement hydraulic pump 20 is the maximum value and the differential pressure does not reach the reference value, the differential pressure of the pressure across the control valve 22a is reduced. The opening of the variable throttle valve 28 is reduced until the differential pressure across the flow compensating valve 40 reaches a predetermined value. As a result, the flow rate supplied to the first hydraulic actuator 4a decreases, and a constant flow rate compensated by the flow rate compensation valve 40 is supplied to the hydraulic motor 3. In the means and means for solving the above problems which explain the constitution of the present invention, the drawings of the embodiments are used to make the present invention easier to understand. However, the present invention is not limited to this. FIG. 1 is a diagram showing the overall configuration of a hydraulic control device for construction equipment according to the present invention. The same parts as those in FIG. 5 are denoted by the same reference numerals, and differences will be mainly described. . FIG. 1 shows a flow rate compensating valve for controlling the flow rate of the hydraulic motor 3 at an inlet circuit of the hydraulic motor 3 by adding the above-described hydraulic motor circuit for the generator to a load sensing control type hydraulic control device. A pressure compensation valve 40 is provided, and the pressure PLm downstream of the pressure compensation valve 40 is also used as a load sensing pressure. In FIG. 1, reference numeral 20 denotes a variable displacement hydraulic pump driven by a prime mover 12 such as a diesel engine.
0a is the displacement volume (tilt amount) of the variable displacement hydraulic pump 20
Is a regulator that regulates Variable displacement hydraulic pump 2
As hydraulic actuators to which the discharge oil from 0 is supplied, in addition to the hydraulic motor 3, three hydraulic cylinder devices 4 a to 4 c for the working machine are provided. Performed by valve 21;
Supply and stop of the discharge oil to the hydraulic cylinder devices 4a to 4c and switching of the supply direction are performed by individual flow control valves 2 respectively.
2a to 22c. Hydraulic motor 3 and hydraulic cylinder device 4a
High pressure selection valves 23 to 25 are provided to take out the maximum pressure among the load pressures PLm, PLa, PLb, and PLc of 4c as load sensing pressure (maximum load pressure) PAmax, and this load sensing pressure PAmax is a differential pressure sensor. 26. The pump pressure PLS is also input to the differential pressure sensor 26. The differential pressure sensor 26 detects a differential pressure ΔPLS between the pump pressure PLS and the load sensing pressure PAmax, and the detection signal is input to the controller 27. . The controller 27 is implemented by a microcomputer or the like.
LS, a signal indicating the tilt angle of the variable displacement hydraulic pump 20 is input from the tilt angle sensor 20b, and the differential pressure Δ
The displacement volume of the variable displacement hydraulic pump 20 is controlled by the regulator 20a so that PLS becomes a predetermined reference value. Further, a signal from a switch 21A for operating the solenoid valve 21 is input to the controller 27. When the switch 21A is turned on, the controller 27 sets a flag for the lifting magnet use mode, and
When A is off, a work mode flag is set. When the switch 21A turns on, the solenoid valve 21 opens, and when it turns off, it closes. Pressure compensation valves 28 to 30 functioning as variable throttle valves corresponding to the cylinder devices 4a to 4c are provided in the middle of the oil passages for individually supplying the discharge oil to the hydraulic cylinder devices 4a to 4c for the working machine. Is provided. Pressure compensating valve 28
Is the spring force applied to the spool by the spring 31 and the load pressure PLa.
And the pressure compensating valve 28
And the control pressure PC generated by the proportional solenoid valve 32 operates according to an equilibrium relationship between the total pressure of the pressing force exerted on the spool by the outlet pressure PZa of the cylinder device 4a.
To control the flow rate of pressurized oil. The pressure compensating valve 29 is generated by the proportional solenoid valve 34 and the total force of the spring force applied to the spool by the spring 33 and the pressing force applied by the load pressure PLb to the spool, the pressing force applied by the outlet pressure PZb of the pressure compensating valve 29 to the spool. The control pressure PC operates according to the balance of the total force of the pressing forces exerted on the spool, and controls the flow rate of the pressure oil to the cylinder device 4b. The pressure compensating valve 30 includes a spring 35
The total force of the pressing force exerted on the spool by the spring force applied to the spool and the load pressure PLc, and the outlet pressure P of the pressure compensating valve 30
The pressing force exerted on the spool by Zc and the control pressure PC generated by the proportional solenoid valve 36 operate according to the equilibrium relationship of the total force of the pressing force exerted on the spool, and controls the flow rate of the pressure oil to the cylinder device 4c. Here, when the hydraulic cylinder devices 4a to 4c are operated in combination, the required flow rate of the actuator driven at the maximum load pressure is still insufficient even if the pump displacement q is increased to the maximum value qmax. A control pressure PC is generated, and the control pressure Pc is increased until ΔPLS reaches a reference value, ie, a spring pressure Pi. The pressure compensating valves 28 to 30 are equivalent to each other. As a representative example, the operation of the pressure compensating valve 28 will be described with reference to FIG. 2A shows that the pressure compensating valve 28 is in a minimum throttle state (fully open state), and FIG. 2B shows that the pressure compensating valve 28 is in a maximum throttle state. In FIG. 1, a pressure compensating valve 40 and a throttle 41 are provided in the middle of a hydraulic pressure supply path for the hydraulic motor 3. The pressure compensating valve 40 is provided with a spring force applied to the spool by the spring 42 and a pressure downstream of the throttle 41, that is, a total value of a pressing force exerted on the spool by the load pressure PLm of the hydraulic motor 3 and a pressure upstream of the throttle 41, That is, it operates according to the balance of the pressing force exerted on the spool by the outlet pressure PZm of the pressure compensating valve 40, and controls the flow rate of the pressure oil supplied to the hydraulic motor 3. The pressure compensating valve 40 operates between the minimum throttle state (full open state) shown in FIG. 3A and the maximum throttle state shown in FIG. As the pressure increases, the state moves from the minimum throttle state in FIG. 3A to the maximum throttle state in FIG. 3B, and the flow rate of the pressure oil supplied to the hydraulic motor 3 is reduced. In this embodiment, each of the pressure compensating valves 28 to 3
The set pressures of the springs 31, 33, 35, 42 of 0, 40 are all equal to Pi, and the spool pressure receiving areas of the outlet pressures PZa to PZc, PZm at the respective pressure compensating valves 28 to 30, 40 are AZ.
a to AZc, AZm, Ala to ALc, ALm for the spool pressure receiving area for load pressures PLa to PLc, PLm, and A for the spring pressure Pi
i, When the spool pressure receiving area of the control oil pressure PC is represented by Ac,
AZa = AZb = AZc = AZm = ALa = ALb = ALc = ALm, A
i = Ac, and the respective area ratios of AZa, AZb, AZc, AZm and Ai, and the respective area ratios of ALa, ALb, ALc, ALm and Ac are set to 1: 0.4. Further, referring to FIG.
If the differential pressure ΔPLS detected by the differential pressure sensor 26 does not reach the reference value even when the displacement volume of the variable displacement hydraulic pump 20 reaches the maximum value, the proportional pressure is set so that the differential pressure ΔPLS becomes the reference value. The control hydraulic pressure Pc output from the solenoid valves 32, 34, 36 is increased to increase the degree of restriction of the pressure compensating valves 28 to 30, and to reduce the flow rate of the hydraulic pressure supplied to the hydraulic cylinder devices 4a to 4c. The reference value of the differential pressure is the spring pressure P in the work mode.
i, 0 in the lifting magnet use mode.
4Pi. Here, a case where the hydraulic cylinder device 4a and the hydraulic motor 3 are operated in combination will be described as an example. (1) The load pressure of the hydraulic cylinder device 4a is
In this case, PLa> PLm,
A large amount of pressure oil tends to flow to the hydraulic motor 3 having a low load. At this time, the pressure compensating valve 40 operates so as to satisfy the following equation: ALm · PLm + Ai · Pi = AZm · PZm (1) Due to the pressure receiving area ratio in the design described above, the differential pressure PZm-PLm across the throttle 41 becomes 0.4 Pi. Here, assuming that the coefficient determined by the density of the hydraulic oil and the shape of the throttle 41 is C and the opening area of the throttle 41 is A41, the inflow flow rate Q3 of the hydraulic motor 3 is expressed by the following equation. Q3 = C · A41 (PZm−PLm) ^1/2 = C · A41 (0.4Pi) ^1/2 (2) Since the coefficient C and the spring pressure Pi are constants, the aperture of the diaphragm 41 is obtained. By selecting the area A41, the flow rate Q3 of the pressure oil to the hydraulic motor 3 becomes the set flow rate. As can be seen from the relationship between the discharge pressure PSL of the hydraulic pump 20 and the discharge flow rate Q shown in FIG. 4, the discharge flow rate of the hydraulic pump 20 at the maximum load pressure is larger than the set flow rate Q3 of the hydraulic motor 3. Is set. The inflow flow rate Q4a of the hydraulic cylinder device 4a is
The differential pressure between the maximum load pressure PLa and the discharge pressure PSL of the hydraulic pump 20, that is, the load pressure PLa and the outlet pressure PLa
When the hydraulic pump 20 discharges the pressure oil so that the differential pressure of Za becomes constant, it is expressed by the following equation. Q4a = C · A22a (PZa−PLa) ^1/2 (3) where A22a is the opening area of the internal passage of the pressure compensating valve 22a, and the flow rate Q4a is a value corresponding to the opening area A22a. . (1.1) PLa> PLm and Q <Q3 +
Q4a The discharge flow rate Q of the hydraulic pump 20 (the displacement q is the maximum value q
max) is smaller than the total value of the set flow rate Q3 of the hydraulic motor 3 and the required flow rate Q4a of the work implement actuator 4a, that is, even when Q <Q3 + Q4a, the relationship of Q> Q3 holds. The hydraulic oil flow rate Q3 is always kept at a set value, and the hydraulic cylinder device 4a is driven by the excess amount of the hydraulic oil Q-Q3. (2) When the load pressure of the hydraulic motor 3 is higher than the load pressure of the work implement actuator 4a, that is, P
The case of Lm> PLa is as follows. (2.1) PLa <PLm and Q ≧ Q3 + Q4a At this time, the discharge flow rate Q of the hydraulic pump 20 (the displacement volume q
Is qmax), but the flow rate Q4 required by the hydraulic cylinder device 4a is
a is larger than the sum of the set flow rate Q3 of the hydraulic motor 3 and
That is, if Q ≧ Q4a + Q3, the hydraulic pump 20 applies the differential pressure ΔP between the maximum load pressure PLm and the discharge pressure PLS of the hydraulic pump 20.
LS, that is, the pressure oil is discharged at a flow rate at which the differential pressure between the load pressure PLm and the outlet pressure PZa becomes constant, so that the pressure oil flow rate Q4a to the hydraulic cylinder device 4a is a value corresponding to the opening area A22a. The pressure oil flow Q3 to the motor 3 becomes the set flow. (2.2) PLa <PLm and Q <Q3 +
Q4a The discharge flow rate Q of the hydraulic pump 20 is smaller than the hydraulic cylinder device 4a.
Is smaller than the sum of the required flow rate Q4a and the set flow rate Q3 of the hydraulic motor 3, that is, when Q <Q4a + Q3,
In the conventional circuit in which the pressure oil flows preferentially to the hydraulic cylinder device 4a having a low load, the flow rate Q3 of the pressure oil to the hydraulic motor 3 decreases, and the set flow rate cannot be secured. When the difference .DELTA.PLS between the load pressure PLm of the hydraulic motor 3 which is the maximum load pressure and the discharge pressure PLS of the hydraulic pump 20 becomes smaller than the set differential pressure (reference value) Pi, the controller 2
7 is a hydraulic pump 20 for making this differential pressure a set differential pressure.
, A command to increase the discharge flow rate, that is, a command to increase the tilt angle. When the displacement volume of the hydraulic pump 20 has already reached the maximum value, the pressure oil flow rate does not increase any more. When the controller 27 detects that the pressure difference ΔPLS does not reach the reference value Pi despite the discharge capacity q of the hydraulic pump 20 being the maximum value qmax, the controller 27 sets the proportional solenoid valve 3
2 to increase the control oil pressure PC and shift the pressure compensating valve 28 to the throttle state shown in FIG. At this time, the variable pressure compensating valve 28 operates such that: ALa · PLa + Ai · Pi = AZa · PZa + Ac · Pc (4) Here, by substituting the design area ratio, it is expressed as follows: PZa−PLa = 0.4 (Pi−Pc) (5) Therefore, when the equation (5) is modified, the following equation is obtained: PC = Pi − {(PZa−PLa) /0.4} (6) As can be seen from equation (2), the hydraulic motor 3
In order to set the flow rate Q3 of the pressurized oil to be supplied to the set value, the pressure before and after the throttle 41 may be set to 0.4 Pi. PLa <PL
m, the pressure PLm of the hydraulic motor 3 is input to the differential pressure sensor 26 as the load sensing maximum load pressure PAmax, so the differential pressure ΔP used for the load sensing control is
If LS is controlled to 0.4 Pi, the set flow rate flows to the hydraulic motor 3. Therefore, the controller 27
The drive of the proportional solenoid valve 32 is controlled until the detected value ΔPLS becomes 0.4 Pi. At this time, the control pressure Pc is expressed by the following equation. PC = (ΔPLS−PZa + PLa) /0.4 (7) When the control oil pressure PC is applied to the pressure compensating valve 28, the degree of restriction of the pressure compensating valve 28 increases.
The pressure oil flow Q4a to the hydraulic cylinder device 4a is reduced. When the pressure difference ΔPLS becomes 0.4 Pi, the control pressure Pc is maintained. Therefore, the flow rate Q of the pressure oil supplied to the hydraulic motor 3
3 is kept at the set value. FIG. 5 shows this control flow. In this load sensing control type hydraulic control device, the differential pressure ΔP
If LS is smaller than the reference value Pi (Yes at Step S1), the displacement angle of the hydraulic pump 20 is increased within the range where the displacement volume q of the hydraulic pump 20 is equal to or less than the maximum value qmax (Yes at Step S2). The removal volume q is increased (step S3). On the other hand, when the differential pressure ΔPLS is equal to the reference value Pi (Yes at Step S4), the tilt angle at that time is maintained (Step S5), and when the differential pressure ΔPLS is larger than the reference value Pi (No at Step S4). If the displacement volume q of the hydraulic pump 20 is equal to or more than the minimum value qmin (Yes at step S6), the displacement angle of the hydraulic pump 20 is reduced by decreasing the tilt angle of the hydraulic pump 20.
Is reduced (step S7). If the pressure difference .DELTA.PLS is smaller than the reference value Pi in a state where the displacement volume q of the hydraulic pump 20 has reached the maximum value qmax (No at step S2), the control to generate the proportional solenoid valves 32, 34 and 36 is performed. The hydraulic pressure Pc is increased to increase the degree of restriction of the pressure compensating valves 28 to 30, and the flow rate of the hydraulic pressure supplied to the hydraulic cylinder devices 4a to 4c is reduced (step S8). If it is determined in step S9 that the mode is the lifting magnet use mode, the pressure difference ΔPLS between the inlet pressure PZn of the flow control valves 22a to 22c and the load sensing pressure PAmax.
Becomes equal to the reference value 0.4 Pi (Yes at step S10).
The control pressure Pc is increased. Thus, in the lifting magnet use mode, PLa <PLm and Q <Q
Even in the case of 3 + Q4a, the flow rate of the hydraulic oil to the hydraulic motor 3 is ensured preferentially to the set value. If the operation mode is determined in step S9, the inlet pressures PZn of the flow control valves 22a to 22c are determined.
Pressure difference ΔPLS between the load sensing pressure PAmax and the reference value P
The control pressure Pc is increased until i is reached (Yes at step S11). The control pressure PC in the conventionally known load sensing circuit and the advantages of the present invention will be described. Now, in the circuit of FIG. 1, the hydraulic motor 3 and the working machine hydraulic cylinder device 4a are operated in combination, and the hydraulic motor 3
Is higher than the load oil pressure of the cylinder device 4a,
When the total value of the hydraulic oil flow required to drive the hydraulic motor 3 at the set rotation speed and the hydraulic oil flow required by the hydraulic cylinder device 4a is larger than the maximum discharge flow of the variable displacement hydraulic pump 20, the following operation is performed. To try to send pressure oil to the hydraulic motor 3 side. The load pressure PLm of the hydraulic motor 3 is selected as the load sensing pressure PAmax and input to the differential pressure sensor 26, and the differential pressure sensor 26 receives the load sensing pressure PAmax.
And the pressure difference between the pump pressure PLS and the controller 27 is detected and input to the controller 27. The controller 27 determines that the pressure difference ΔPLS is not equal to the reference value even though the pump displacement is the maximum value, and proportionally adjusts the pressure difference between the front and rear of the control valve 22a of the cylinder device 4a to the reference value of ΔPLS. The control pressure Pc is applied from the solenoid valve 32 to the pressure compensating valve 28. As a result, the hydraulic cylinder device 4
The pressure oil flow supplied to a decreases, and the pressure oil flow to the hydraulic motor 3 increases. However, the control pressure Pc controls ΔPLS to the spring pressure Pi so that the pressure difference between the upstream and downstream of the control valve 22a is equal to the differential pressure ΔPLS between the maximum load pressure Pamax and the pump discharge pressure PLS in the load sensing control. is there. Therefore, the flow rate control such that the hydraulic motor 3 rotates at the set rotation speed is not performed, and the generator 2 cannot be driven at the set rotation speed stably. On the other hand, in this embodiment, when the lifting magnet is used, the load sensing differential pressure ΔPLS is controlled to 0.4 Pi so that the flow rate of the power generating hydraulic motor 3 becomes a set value. In the operation mode without using the lifting magnet, the load sensing differential pressure ΔPLS is set to Pi.
To effectively use the flow rate. In the above embodiment, the pressure compensating valves 28 to 3
0 and 40 are arranged on the upstream side of the flow control valves 22a to 22c and the throttle 41, respectively.
8 to 30 and 40 may be arranged downstream of the flow control valves 22 a to 22 c and the throttle 41. Also, the pressure compensating valve 2
8 to 30 may be provided in the return side circuits of the hydraulic cylinder devices 4a to 4c. Furthermore, if a variable throttle function is added to the flow control valves 22a to 22c, the flow control valves 22a to 22c
2c may also serve as the pressure compensating valves 28 to 30. In the above, the driving of the generator 2 by the lifting magnet 13 has been described. However, the present invention can be applied to, for example, a case in which a winch is driven at a constant rotation in addition to the generator, and the hydraulic motor 3 is used. The present invention can also be applied to a hydraulic actuator that performs a linear motion, such as a hydraulic cylinder device other than an actuator that performs a simple rotary motion. Further, the main spool may be directly throttled without using the variable pressure compensating valve. As described above, according to the present invention, when the load pressure of the second hydraulic actuator such as a hydraulic motor is lower than the load pressure of the first hydraulic actuator such as a hydraulic cylinder, The flow rate is controlled to a predetermined constant value by a flow rate compensating valve provided in the circuit of the second hydraulic actuator. When the load pressure of the second hydraulic actuator is higher than the load pressure of the first hydraulic actuator,
If the displacement volume of the variable displacement hydraulic pump is the maximum value and the differential pressure between the front and rear pressures of the control valve of the first hydraulic actuator does not reach the reference value, the throttle of the flow compensation valve of the second hydraulic actuator is restricted. The variable throttle valve of the first hydraulic actuator is throttled until the differential pressure before and after reaches a predetermined value for obtaining the predetermined flow rate, so that the flow rate of the hydraulic oil supplied to the first hydraulic actuator decreases, The flow rate of the pressure oil diverted to the hydraulic actuator is set to the set flow rate, and the driving state of the second hydraulic actuator is kept constant. Thus, the driving state of the second hydraulic actuator is kept constant under any state. For example, if the second hydraulic actuator is a hydraulic motor that drives a generator of a lifting magnet, the generator is always set. Driven at the rotation speed, the attraction force of the lifting magnet becomes stable irrespective of the load fluctuation of the hydraulic cylinder device for the working machine.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a diagram showing an overall configuration of an embodiment of a hydraulic control device for a construction machine according to the present invention. 2 (a) and 2 (b) are circuit diagrams showing details of a pressure compensating valve upstream of a control valve of the hydraulic control device of FIG. 1 in different states. 3A and 3B are circuit diagrams showing details of a pressure compensating valve upstream of a motor of the hydraulic control device of FIG. 1 in different states. FIG. 4 is a graph showing a relationship between a discharge pressure and a discharge flow rate of a variable displacement hydraulic pump. FIG. 5 is a flowchart showing a control flow of a hydraulic control device for a construction machine according to the present invention. FIG. 6 is a hydraulic circuit diagram showing an example of a conventional hydraulic control device. 7A and 7B are circuit diagrams showing details of a flow rate priority valve of the hydraulic control device of FIG. 6 in different states. [Description of Signs] 1 Main hydraulic pump 3 Hydraulic motor 4, 4a-4c Hydraulic cylinder device 20 for work equipment 20 Variable displacement hydraulic pump 20a Regulator 20b Tilting angle sensor 21 Electromagnetic on-off valve 21A Switches 22a-22c Flow control valves 23-25 High pressure selection valve 26 Differential pressure sensor 27 Controller 28-30 Pressure compensation valve 32, 34, 36 Proportional solenoid valve 40 Pressure compensation valve 41 Throttle

──────────────────────────────────────────────────続 き Continued on the front page (58) Field surveyed (Int. Cl. ⁷ , DB name) F15B 11/00-11/22 F04B 49/00

Claims (1)

(57) [Claim 1] A variable displacement hydraulic pump, first and second hydraulic actuators driven by oil discharged from the hydraulic pump, the hydraulic pump and the first hydraulic pressure A control valve provided between the hydraulic pump and the second hydraulic actuator, the control valve being provided between the hydraulic pump and the second hydraulic actuator, for controlling a differential pressure across the throttle to a predetermined value; A flow compensating valve for keeping the flow rate of the second hydraulic actuator constant, a displacement volume adjusting means for adjusting a displacement volume of the hydraulic pump, and a middle of a hydraulic supply path for the first hydraulic actuator. A variable throttle valve provided on the first hydraulic actuator to control a flow rate supplied to the first hydraulic actuator; a discharge pressure of the hydraulic pump and the first and second hydraulic actuators; The displacement of the variable displacement hydraulic pump is controlled so that the differential pressure from the maximum value of the load pressure becomes the reference value, and the displacement of the variable displacement hydraulic pump is the maximum value and the differential pressure is Control means for reducing the opening of the variable throttle valve until the differential pressure reaches the predetermined value of the flow rate compensating valve when the reference pressure does not reach the reference value.