JP2749733B2 - Hydraulic control equipment for hydraulic construction machinery - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic system capable of selecting and using one of a plurality of types of front attachments having different required PQ diagrams, such as a hydraulic shovel bucket and a vibrator. The present invention relates to a hydraulic control device for a construction machine.

[0002]

2. Description of the Related Art Since a hydraulic excavator mainly performs bucket work, a PQ diagram (pressure flow characteristic) of a hydraulic circuit is designed to be optimal for bucket work.
In general, heat balance of an engine or a hydraulic device is designed based on a Q diagram. However, for example, if a vibro operation is performed with the bucket PQ diagram, the flow rate under a high load is insufficient, and the workability deteriorates.

[0003]

The problem to be solved by the present invention is disclosed in
As disclosed in Japanese Patent Publication No. 167042, it is conceivable to increase the engine speed by a predetermined amount at a high load to increase the flow rate at a high load. However, if only an increase in the flow rate under a high load is attempted during the vibrating operation, the cooling performance is insufficient in the heat balance designed with the PQ diagram suitable for the bucket operation, and the engine may be overheated. Therefore, if the heat balance is designed to cope with the PQ diagram for the vibro from the beginning, when the bucket is used, the specification becomes overspecified, which not only increases the cost but also increases the size of the entire apparatus.

An object of the present invention is to provide a heat balance based on a PQ diagram suitable for a bucket work or the like without deteriorating the heat balance in an attachment work requiring a large flow rate under a high load such as a vibro work. An object of the present invention is to provide a hydraulic control device for a hydraulic construction machine with improved workability.

[0005]

The present invention will be described with reference to FIGS. 1 to 3 showing one embodiment. The present invention relates to a variable displacement hydraulic pump 1 driven by a prime mover 13, Hydraulic actuator (4) for front attachment, which is driven by oil discharged from the engine, and a control valve (2) for controlling the flow rate and direction to the hydraulic actuator (4)
And a displacement volume changing means 40 for changing a displacement volume of the hydraulic pump 1 in accordance with at least a load of the hydraulic actuator (4). As a front attachment hydraulic actuator, a first actuator for bucket work, The present invention is applied to a hydraulic control device of a hydraulic construction machine capable of selectively using the second actuator 4 which requires a small flow rate at a low load and a large flow rate at a high load as compared with the bucket work. The purpose of the above is to provide the second actuator 4
Is used, the pressure flow characteristic of the variable displacement hydraulic pump 1 is set to a smaller flow rate at a low load than when the first actuator is used, and to a larger flow rate at a high load than when the first actuator is used. This is achieved by providing the pump discharge flow rate control means 50 that performs the above. The pump discharge flow rate control means 50 of the second aspect increases the rotation speed of the prime mover 13 at a high load. The pump discharge flow control means 50 in claim 3 reduces the displacement volume of the variable displacement hydraulic pump 1 at a low load. The control device according to claim 4, wherein
Required flow rate reducing means 21 for reducing the maximum required flow rate of the control valve 2 when the actuator is used, and the pump discharge flow rate controlling means 50 controls the variable displacement hydraulic pump 1 And a load sensing means 61 (FIG. 2) for controlling the displacement volume of the first actuator and reducing the target differential pressure when the second actuator is used than when the first actuator is used. The pump discharge flow control means 50 according to claim 5 increases the displacement volume of the variable displacement hydraulic pump 1 at a high load.

[0006]

For example, when the second actuator, which is the vibrating hydraulic motor 4, is used, the pressure flow characteristic of the variable displacement hydraulic pump 1 is lower when the load is lower than when the first actuator, which is the bucket hydraulic cylinder, is used. Set a smaller flow rate and a larger flow rate at high load. Thereby, the workability under a high load is improved without deteriorating the overall heat balance.

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 easy to understand. However, the present invention is not limited to this.

[0008]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS One embodiment of the present invention will be described with reference to FIGS. FIG. 1 is a diagram illustrating a hydraulic circuit of a wheel-type hydraulic excavator that performs load sensing control and input torque limiting control, and illustrates a case where a vibro attachment is used. Reference numeral 1 denotes a variable displacement hydraulic pump driven by the engine 13. The number of revolutions of the engine 13 is controlled by moving the governor lever 13b of the governor 20a to the pulse motor 14a in accordance with the amount of depression of the fuel lever 57a or the travel pedal 3.
Is controlled by turning. Then, the variable displacement hydraulic pump 1 is driven in accordance with the engine speed, and the discharge oil is guided to the vibrating hydraulic motor 4 via the auxiliary attachment control valve 2 and the boom is controlled via the boom control valve 5. It is guided to the driving oil cylinder 6. The control valve 5 is operated by operating the boom operation lever 7 and is controlled by the output pressure of the boom pilot valve 8. Reference numerals 9 and 10 denote pressure compensating valves, which are used for compensating that the hydraulic motor 4 and the hydraulic cylinder 6 operate independently. The differential pressure between the control valves is constant regardless of the pump discharge pressure. It has the function of In FIG. 1, only the essential parts of the present invention are shown, and a traveling system, a turning system, and the like are omitted.

The load sensing control means the pressure before and after the control valve 2 for the work attachment or the control valve 5 for the boom, that is, the inlet pressure (pump pressure) and the outlet pressure of the control valves 2 and 5 (in FIG. The displacement volume of the variable displacement hydraulic pump 1 (hereinafter referred to as tilt) is set so that the differential pressure between the load pressure of the cylinder 6 and the pressure on the high pressure side selected by the shuttle valve 11 and called the load sensing pressure) becomes a constant value. In this case, the pump pressure is maintained higher than the load sensing pressure by a predetermined target value. In this embodiment, the load sensing control is performed electronically as described later. By performing the load sensing control as described above, the pump tilt angle is controlled such that the pump discharge flow rate is equal to the required flow rate of the control valve 2 or 5, and no extra flow rate is discharged. Since fuel efficiency is eliminated, fuel efficiency is improved and operability is also good.

The input torque limiting control means that the actual engine speed Nr, the target engine speed Nθ based on the governor lever position, and the pump pressure Pp are used to determine the pump tilt angle q.
The pump displacement (tilt angle) is controlled so that the pump input torque determined by the pump pressure Pp does not exceed the engine output torque.

The tilt angle of the variable displacement hydraulic pump 1, that is, the displacement volume, is controlled by a tilt angle control device 40.
The tilt angle control device 40 controls a piston position by a hydraulic pump 41 driven by the engine 13, a pair of solenoid valves 42 and 43, and pressure oil from the hydraulic pump 41 according to switching of the solenoid valves 42 and 43. The tilt angle of the hydraulic pump 1 is controlled according to the piston position of the servo cylinder 44. Here, the pair of solenoid valves 42 and 43 are switch-controlled by the controller 50.

Reference numeral 51 denotes a tilt angle sensor for detecting the tilt angle θs of the hydraulic pump 1, and 52 denotes a discharge pressure Pp of the hydraulic pump 1.
Is a pressure sensor for detecting the rotation speed N of the engine 13.
The rotational speed sensor 54 for detecting r is the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the actuator (in FIG. 1, the larger one of the load pressure of the hydraulic motor 4 and the load pressure of the hydraulic cylinder 6; (Selected by the valve 11). Reference numeral 55 denotes a potentiometer for detecting the rotation amount Nθ of the governor lever 13b, and reference numeral 56 denotes a pressure sensor for detecting the pressure Pt of the pilot valve 12 in accordance with the operation amount of the travel pedal 3, and the detection results of these sensors. Is input to the controller 50. Reference numeral 57 denotes a rotation speed setting device for commanding a target rotation speed X in accordance with the amount of manual operation of the fuel lever 57a, and the command signal is also input to the controller 50.

In FIG. 1, reference numeral 20 denotes a pilot valve for operating a vibrator, reference numeral 21 denotes an electromagnetic proportional pressure reducing valve, and reference numeral 22 denotes an attachment selection switch. When the selection switch 22 is switched to the contact a during the vibro operation, the voltage from the controller 50 is output. The signal is applied to the solenoid portion of the electromagnetic proportional pressure reducing valve 21, and the pressure of the hydraulic pressure source 23 is reduced, for example, from 70 kg / cm ² to 25 kg / cm ² and applied to the pilot valve 21 for vibro. When, for example, a crusher is used as an attachment other than the vibro, the selection switch 22 is switched to the contact b side. It should be noted that the switching to the b contact side is also performed during normal bucket work. Thereby, the battery voltage is applied to the solenoid portion of the electromagnetic proportional pressure reducing valve 21, and the pressure reducing valve 21 applies the pressure of the hydraulic pressure source 23 to the pilot valve 20 without reducing the pressure. Thus, the pilot valve 20
The maximum output pressure of the pilot valve 20 is reduced by reducing the input pressure of the control valve 2 during the vibrating operation, and the maximum opening area of the control valve 2 is limited. Therefore, in the hydraulic circuit adopting the load sensing control as in the present embodiment, in combination with the control by the load sensing control unit 61 described later, only the amount by which the maximum required flow rate of the vibro control valve 2 is reduced is reduced. The maximum tilt angle of the hydraulic pump 1 is regulated, and the maximum discharge flow rate is reduced as compared with the time of bucket work.

The controller 50 has a tilt angle control circuit 60 for controlling the pump tilt angle as shown in FIG. 2, and this control circuit 60 is provided with a load sensing control unit (hereinafter referred to as LS).
Control unit) 61, a torque control unit 62, and a minimum value selection unit 6.
3 and a servo control unit 64.

The LS control section 61 has a setting section 61a for setting a target differential pressure ΔPLSR1 for a vibrating operation.
The difference Δ (P) between the differential pressure ΔPLSD detected at
LS) is calculated by the deviation device 61b, and the deviation Δ (PLS)
The amount of change ΔθL of the target value is calculated by the function generator 61c based on Further, the change amount ΔθL of the target value is integrated by the integrator 61d and output as a target pump tilt angle θL for load sensing control.

The torque control unit 62 includes a target torque calculation unit 62
The target torque calculation unit 62a determines the torque Tr exerted by the engine from the governor lever position Nθ detected by the potentiometer 55, and calculates the engine speed Nr detected by the speed sensor 53. And a deviation ΔT from the governor lever position Nθ detected by the potentiometer 55 is calculated, and the deviation ΔT is added to the torque Tr to calculate a target torque Tpo. The pump discharge pressure Pp detected by the pressure sensor 52 is input to a reciprocal calculator 62b for bucket work to calculate a reciprocal 1 / Pp, and the pump pressure Pp is calculated by a reciprocal calculator 6 for vibro work.
The reciprocal α / Pp is also calculated by inputting it to 2c. In addition,
α is a value larger than 1.

Reference numeral 62h denotes a switch linked to the vibrator selection switch 22. When the switch 22 is switched to the a contact for vibro, the switch 62h is switched to the a contact, and is switched to the b contact except during the operation of the vibro. If the pump discharge pressure Pp detected by the pressure sensor 52 becomes equal to or higher than Pup when the switch 62h is switched to the contact a during the vibrating operation, the function generator 62
d outputs a high level signal to switch the switch 62e to the a side, and outputs a low level signal from the function generator 62d to switch the switch 62e to the b side when the pump discharge pressure Pp falls below Pd. The reciprocal 1 / Pp of the pump pressure Pp calculated by the reciprocal calculating unit 62b
The reciprocal α / Pp of the pump pressure Pp calculated by the reciprocal calculation unit 62c is connected to the contact a of the switch 62e, and any reciprocal is selected from the switch 62e. The 1 / Pp or α / Pp selected by the switch 62e is multiplied by the target torque Tpo by the multiplier 62f, thereby obtaining θps (ΔTpo = θp).
* Pp). Further, this θps is filtered by a first-order lag element filter 62g and output as a target pump tilt angle θT for input torque limiting control.

The minimum value selector 63 selects the smaller one of the two target tilt angles θL and θT and inputs the smaller value to the servo controller 64 as the tilt angle command value θr. In the servo control unit 64, a deviation Δθ between the input tilt angle command value θr and the tilt angle feedback value θs detected by the tilt angle sensor 51 is obtained by a deviation device 64a. On / off signals of the solenoid valves 42 and 43 are output from the function generator 64b based on the function generator 64b. Thus, the tilt angle control device 40 is controlled so that the pump tilt angle θs matches the tilt angle command value θr.

The controller 50 also includes an engine speed control circuit 7 for controlling the engine speed as shown in FIG.
Has zero. The function generator 71 outputs the target rotation speed Nx based on the rotation speed command value (the operation amount of the fuel lever 57a) x sent from the rotation speed setting device 57, and the function generator 72 outputs the target rotation speed Nx from the traveling pilot pressure sensor 56. The target rotation speed Nx is output based on the traveling pilot pressure Pt sent. The maximum value of these target rotation speeds is selected by the maximum value selection circuit 73 and connected to the a contact of the switch 74.
The number-of-revolutions setting value circuit 75 is connected to the contact b of the switch 74.
The switch 74 selects one of the rotation speeds.

Reference numeral 79 denotes a switch linked to the vibro selection switch 22.
When the switch is switched to the contact, the switch 79 is switched to the contact a, and is switched to the contact b except when the vibro work is performed. When the switch 79 is switched to the "a" contact during the vibrating operation, the switch 74 is switched by the function generator 76, and is switched to the "b" contact when the pump pressure Pp becomes equal to or more than Pup, and is switched to the "a" contact when the pump pressure Pp becomes equal to or less than Pd. Therefore, when the pump pressure is equal to or higher than Pup, the rotation speed increase set value Nmax is output from the switch 74, and when the pump pressure is equal to or lower than Pd, the engine target rotation speed selected by the maximum value selection circuit 73 is output from the switch 74. Where N
max> Nxmax, Ntmax.

The output Na of the switch 74 is deviated from the governor lever position Nθ by a deviator 77. When the deviation ΔN is equal to or more than a predetermined positive value, the function generator 78 sends a signal to the pulse motor 14 in the forward rotation direction. When the deviation ΔN is equal to or smaller than a predetermined value on the negative side, the function generator 76 supplies a signal in the reverse rotation direction to the pulse motor 27.

Next, the operation of the present embodiment configured as described above will be described. Select switch 22 during vibro work
To the contact a side. Switch 6 of torque control unit 62
2h and each of the switches 79 of the engine speed control circuit 70 are switched to the a contact for vibro. The upper limit of the input pressure of the pilot valve 20 that operates the vibro control valve 2 is limited by the electromagnetic proportional pressure reducing valve 21, and the maximum opening area of the control valve 2, that is, the maximum required flow rate is limited. The tilt angle of the variable displacement hydraulic pump 1 is controlled to a value determined under the control of the tilt angle control circuit 60, and the engine speed becomes a value controlled by the speed control circuit 70.

If the pump pressure Pp is lower than the above-mentioned Pup, the switch 62e of the torque control unit 62 is switched to the contact b side to select the reciprocal calculation unit 62b, and based on the actual engine speed Nr and the target engine speed Nθ, Is calculated as 1 / P calculated by the reciprocal calculator 62b.
p is multiplied by the multiplier 62f and input to the minimum value selection circuit 63 as θT. At this time, the differential pressure sensor 54
ΘL calculated by the load sensing control unit 61 so that the differential pressure ΔPLSD detected at
Is input to the minimum value selection circuit 63. Since the maximum required flow rate of the control valve 2 is limited during the vibrating operation, the maximum value of θL during the vibrating operation is determined by the torque control unit 62.
It is smaller than the maximum value of T.

On the other hand, when the pump pressure Pp is lower than Pup, the switch 74 of the engine speed control circuit 70 is switched to the contact a, and the output of the maximum value selection circuit 73 is selected. Therefore, the engine speed is limited by the fuel lever 5
The target rotation speed Nx determined by the operation amount x of 7a and the target rotation speed Nt determined by the traveling pilot pressure Pt due to the depression of the traveling pedal 3 are controlled to be larger.

The maximum discharge flow rate Q _V max of the variable displacement hydraulic pump 1 when the engine speed is maximized in such a vibrating operation state is as shown by broken lines V1-V2 in the PQ diagram of FIG. is limited to a maximum discharge flow rate Q less than _B max than Vibro work shown by the solid line B1-B2. Also,
In the region between the pressure P2 corresponding to V2 and the pressure Pup,
Since θT determined by the torque control unit 62 is smaller than θL determined by the LS control unit 61, the tilt angle of the variable displacement hydraulic pump 1 is determined according to the θT, and
The characteristic is such that the discharge flow rate Q decreases as the pump pressure Pp increases, as indicated by V2-V3 in the PQ diagram of FIG.

Next, when the pump pressure Pp becomes equal to or higher than the predetermined pressure Pup, the function generator 62d of the torque control section 62 outputs a high-level signal, the switch 62e is switched to the contact a side, and the reciprocal calculation section 62c calculates. Α / Pp is selected and multiplied by the target torque Tpo by the multiplier 62a. At this time, θL calculated by the LS control unit 61 when the control valve 2 for vibro is operated at the full stroke and requesting the maximum flow rate is larger than θT calculated by the torque control unit 62. The tilt angle of the variable displacement hydraulic pump 1 is controlled to θT according to the pump pressure Pp. here,
At the time of the vibro work, since the reciprocal of the target torque Tpo is α / Pp (α> 1), the value of θT is larger than that of the work other than the vibro work.

On the other hand, the function generator 76 of the engine speed control circuit 70 also outputs a high level signal to switch the switch 74 to the contact b. As a result, Nmax (> Nxmax, Ntmax) output from the rotation speed increase set value circuit 75 is input to the deviation device 77 as the target rotation speed Na.
The deviation from the governor lever position Nθ is taken, and the engine speed is controlled to Nmax. By such tilt angle control and engine speed control, the P-
The Q diagram has characteristics as shown by broken lines V3-V4-V5 in FIG.

When the pump pressure Pp drops from the high load region and becomes lower than the pressure Pd (<Pup), the torque control unit 62
And the switch 74 of the engine speed control circuit 70 are both switched to the contact b side. As a result, the torque control unit 62 calculates the target torque θT from the reciprocal 1 / Pp calculated by the reciprocal calculation unit 62a, and sets the engine target rotation speed from Nmax to the value selected by the maximum value selection circuit 73. Reduced. At this time, although illustration is omitted, control is performed so that the engine speed gradually decreases. Therefore, as shown in FIG. 4, the PQ diagram shows the dashed line V with decreasing pump pressure Pp.
The characteristic is represented by θT calculated by the torque control unit 62, as shown by 5-V6-V7-V2.

[0029] Thus, the pump pressure Pp is in a time Vibro work below Pup, the maximum discharge flow rate Q _V max indicated by the dashed line V1-V2 in FIG. 4, the maximum discharge other than Vibro work flow Q _B The pump discharge flow rate is controlled so as to be smaller than max and the pump discharge flow rate when the pump pressure Pp is equal to or higher than Pup is larger than the pump discharge flow rate other than in the vibro-blow operation. As a result, the pressure Pu during the vibro work
The pump discharge flow rate can be increased in a high load region of p or more, and workability is improved. Then, the maximum pump discharge flow rate Qvmax pump pressure is less than Pup, (e.g., bucket work) other than Vibro work because with less than the maximum discharge flow rate Q _B max, the overall heat balance during Vibro work bucket This can be equivalent to a heat balance at the time of work or the like, and the workability of the vibro can be improved without increasing the size of the engine cooling system or the cooling system of the hydraulic equipment.

In the above embodiment, the load sensing control unit and the input torque limit control unit are provided in the controller, and the tilt angle control is performed by the solenoid valve. However, the configuration may be as shown in FIG. . That is, the load sensing control unit and the input torque limit control unit are connected to the load sensing regulation.
The controller 111 and the torque regulator 25 may be used, respectively, and the tilt angle control may be performed by the regulators 111 and 25 using hydraulic pressure. In this case, the spring force of the target differential pressure setting spring 111a provided on the load sensing regulator 111 is equal to the electromagnetic proportional solenoid 11
1b, the spring force can be reduced by, for example, operating the electromagnetic proportional solenoid 111b by a controller at the time of vibrating work, and increasing the spring force by the electromagnetic proportional solenoid 111b at times other than the vibrating work. Further, the spring force of the spring 25a of the torque regulator 25 is increased by the electromagnetic proportional solenoid 25b when the pump pressure Pp is equal to or higher than Pup. Even with such a device,
The same operation and effect as those of the above-described embodiment can be obtained.
Reference numeral 26 denotes a tilt cylinder.

In the above, both the engine speed and the pump tilt angle are increased in the region where the pump pressure Pp is equal to or higher than Pup. However, only the engine speed or the pump tilt angle is increased. Is also good. FIG. 6 is a PQ diagram when only the engine speed is increased, and FIG.
Shows a PQ diagram when only the pump tilt angle is increased. Note that these controls may be configured so as to perform the engine speed control and the pump tilt angle control of the above-described embodiment, respectively, and therefore, specific circuits are omitted.

In the above description, the input pressure of the pilot valve 20 for vibro is reduced by the pressure reducing valve 21 at the time of the vibrating operation to limit the maximum opening area of the control valve 2 for vibro, but the pressure reducing valve 21 is omitted. The target differential pressure ΔPLSR of the load sensing control unit 61 may be set to two values, large and small, and a small target differential pressure may be selected in conjunction with the operation of the selection switch 22 during the vibrating operation.

In the above, the case where the load sensing control is adopted has been described. However, the present invention can be applied to a hydraulic construction machine which does not perform the load sensing control. Although the first actuator has been described as a vibratory hydraulic motor, the actuator is not limited to a vibratory hydraulic motor as long as the actuator requires a PQ diagram similar to the hydraulic motor 4.

In the configuration of the above embodiment, the engine 1
3 is a prime mover, a hydraulic motor 4 is a first hydraulic actuator, a bucket cylinder (not shown) is a second hydraulic actuator, a tilt angle control device 40 is a displacement volume changing means, a torque control circuit 60 and an engine speed. Control circuit 70
Constitute pump discharge flow rate control means.

[0035]

According to the present invention, a hydraulic actuator for a first front attachment such as a bucket cylinder and a second front which requires a larger flow rate under a high load than the first hydraulic actuator such as a vibrator motor. When using the second hydraulic actuator in a hydraulic construction machine in which the hydraulic actuator for attachment can be selectively used, by increasing the pump discharge flow rate under a high load and decreasing the discharge flow rate under a low load, the engine and hydraulic pressure can be reduced. The heat balance in the case of using the second hydraulic actuator can be made equivalent to the heat balance in the case of using the first hydraulic actuator without increasing the cooling capacity of the device, and the heat balance can be reduced without increasing the size of the cooling device. 2 to improve workability during high-load work using the hydraulic actuator. Can.

[Brief description of the drawings]

FIG. 1 is a diagram showing an overall configuration of a hydraulic control device according to an embodiment of the present invention.

FIG. 2 is a block diagram showing a tilt angle control circuit provided in the controller of FIG. 1;

FIG. 3 is a block diagram showing an engine speed control circuit provided in the controller of FIG. 1;

FIG. 4 is a PQ diagram of the variable displacement hydraulic pump of FIG. 1;

FIG. 5 is a diagram showing an embodiment of a hydraulic load sensing control unit and an input torque limit control unit;

FIG. 6 is a PQ diagram of another embodiment.

FIG. 7 is a PQ diagram of another embodiment.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Variable displacement hydraulic pump 2 Vibro control valve 4 Hydraulic motor 20 Pilot valve for vibro 22 Selection switch 40 Tilt angle control device 50 Controller 60 Torque control circuit 61 LS control section 61a, 61b Target differential pressure setting section 61c Selection switch 62 Torque Control section 62b, 62c Reciprocal calculation section 62e Switch 70 Engine speed control circuit 74 Switch 75 Speed up setting value circuit

Claims (5)

(57) [Claims] 1. A variable displacement hydraulic pump driven by a prime mover, a front attachment hydraulic actuator driven by oil discharged from the hydraulic pump, a control valve for controlling a flow rate and a direction to the hydraulic actuator, A displacement volume changing means for changing a displacement volume of the hydraulic pump in accordance with a load of the hydraulic actuator, wherein the front attachment hydraulic actuator is compared with a first actuator at the time of bucket work and at the time of bucket work. A hydraulic control device for a hydraulic construction machine, which can selectively use a second actuator which requires a small flow rate at a low load and a large flow rate at a high load, wherein a variable displacement hydraulic pressure is used when the second actuator is used. The pressure flow characteristics of the pump are lower than when using the first actuator when the load is low. Set the small flow rate, high load hydraulic control system for a hydraulic construction machine characterized by comprising a pump discharge flow rate control means for setting the large flow rate than when using the first actuator. 2. The hydraulic control device for a hydraulic construction machine according to claim 1, wherein the pump discharge flow rate control means increases the number of revolutions of the prime mover under a high load. 3. The hydraulic control device for a hydraulic construction machine according to claim 1, wherein said pump discharge flow rate control means reduces a displacement volume of said variable displacement hydraulic pump at a low load. 4. The control device according to claim 3, further comprising a required flow rate reducing means for reducing a maximum required flow rate of the control valve when the second actuator is used, wherein the pump discharge flow rate controlling means is arranged before and after the control valve. Load sensing means for controlling the displacement of the variable displacement hydraulic pump so that the pressure difference becomes the target differential pressure; and a target for reducing the target differential pressure when using the second actuator than when using the first actuator. A hydraulic control device for a hydraulic construction machine, comprising: a differential pressure switching unit. 5. The hydraulic control device for a hydraulic construction machine according to claim 1, wherein said pump discharge flow rate control means increases a displacement volume of said variable displacement hydraulic pump at a high load.