JPH0351502A - Control device of load sensing control hydraulic circuit - Google Patents

Abstract

PURPOSE:To set a target rotational frequency to low rotation so as to reduce fuel cost and noise by automatically increasing the rotational frequency of a prime mover when the discharge quantity of a hydraulic pump is saturated. CONSTITUTION:When a directional control valve 59 is set in the forward position and a pilot pressure reducing valve 58 is stepped in the condition where a setting device 68 sets a target rotational frequency smaller than the maximum rotational frequency of an engine 63, corresponding to the operation, pressure oil is supplied to a traveling motor 3. In the case where the load is comparatively small in low-speed running on a flat road, a control unit 40 conducts load sensing control in such a manner that the differential pressure between the discharge pressure and the running load pressure is constant. In case of conducting high speed running, when the demand flow of a flow control valve 4 is increased, and the discharge quantity of a hydraulic pump 1 is to be saturated, a target discharge quantity deviation is computed as a positive value, and the corresponding correction rotational frequency is increased while it is positive to increase the engine rotating speed. Thus, the target rotational frequency can be set to low rotation so as to reduce fuel cost and noise.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to a load sensing control hydraulic circuit for hydraulic machines such as hydraulic excavators and hydraulic cranes. The present invention relates to a control device for a load sensing control hydraulic circuit that is configured to maintain a constant differential pressure higher than the evening load pressure (in the case of single operation) or the maximum load pressure of a plurality of actuators (in the case of combined operation).

[Conventional technology]

In recent years, load sensing control hydraulic circuits have been attracting attention as hydraulic circuits for hydraulic machines such as hydraulic excavators and hydraulic cranes. This hydraulic circuit is connected between a hydraulic pump driven by a prime mover and a plurality of hydraulic actuators, and includes a pressure-compensated flow control valve that controls the flow rate of pressure oil supplied to the actuators in accordance with an operation signal from an operation means. and a load sensing regulator that maintains the discharge pressure of the hydraulic pump higher than the load pressure of the actuator (in the case of single operation) or the maximum load pressure of the plurality of actuators (in the case of combined operation) by a constant differential pressure. A flow control valve with pressure compensation has a pressure compensation function that controls the flow rate at a constant level regardless of fluctuations in load pressure or hydraulic pump discharge pressure.
A flow rate l4 proportional to the amount of operation is supplied to each actuator, and independence of operation of each actuator is ensured during combined operation of a plurality of actuators. In addition, the load sensing regulator saves energy by providing the minimum flow rate that can accommodate the maximum load pressure of the actuator for the discharge amount of the hydraulic pump.

[Problem to be solved by the invention]

However, this load sensing control hydraulic circuit has the following problems specific to load sensing control.

The discharge amount of a variable displacement hydraulic pump is determined by the displacement, and in the case of a swash plate type, it is determined by the product of the amount of tilting of the swash plate and the number of revolutions, and the discharge amount increases as the tilt displacement increases. The amount of tilting of this swash plate has a maximum amount of tilting determined from the structure, and at one set rotation speed of the prime mover, the discharge amount of the hydraulic pump becomes maximum at this maximum amount of tilting. Furthermore, when the input torque of the hydraulic pump exceeds the output torque of the prime mover, the rotational speed of the prime mover decreases, and in the worst case, the engine stalls. Therefore, in order to avoid this, input torque limit control of the hydraulic pump is generally performed15. This control limits the increase in the amount of tilting of the swash plate and limits the maximum value of the discharge amount when the input torque of the hydraulic pump attempts to exceed the output torque of the prime mover.

As described above, a hydraulic pump has a maximum discharge amount that is determined by its structure and a maximum discharge amount that is limited by input torque limit control, and in any case, there is a limit, that is, a maximum possible discharge amount, for the discharge amount of a hydraulic pump. Therefore, when operating a single or multiple hydraulic actuators, if the required flow rate or the total required flow rate commanded by the operating lever becomes larger than the maximum possible discharge rate of the hydraulic pump, load sensing control is applied to the discharge rate (tilt). Even if you try to increase the amount (transfer amount), you will not be able to increase it. That is, the discharge amount of the hydraulic pump becomes saturated. When the discharge amount of the hydraulic pump becomes saturated in this way, the discharge pressure of the hydraulic pump decreases, making it impossible to maintain a constant differential pressure against the maximum load pressure.In the case of single operation, the actuator moves at the speed commanded by the control lever. During combined operation, most of the pressure oil discharged from the hydraulic pump flows to the actuator on the low pressure side, and no pressure oil is supplied to the actuator on the high pressure side, making it impossible to perform smooth combined operation.

One possible way to solve the above problem is to set the pump capacity to a large capacity according to the maximum required flow rate or the total maximum required flow rate of the actuator, but this increases the cost of parts and is economically disadvantageous. It is disadvantageous. In addition, the above problem can be alleviated by always setting the engine speed at a high rotation speed, but this reduces fuel consumption, noise, and engine speed.
This is unfavorable in terms of durability of equipment such as pumps.

The purpose of the present invention is to reduce fuel consumption and noise by normally setting the prime mover at low rotation speed, and when the hydraulic pump discharge volume is saturated, the engine rotation speed is automatically increased to increase the pump discharge volume. The object of the present invention is to provide a control device for a load sensing control hydraulic circuit that can obtain a desired actuator speed and excellent combined operability.

[Means to solve the problem]

In order to achieve the above object, according to a first aspect of the present invention, there is provided a prime mover, a variable displacement hydraulic pump driven by the prime mover, and at least one pump driven by pressure oil discharged from the hydraulic pump. one hydraulic actuator, connected between the hydraulic pump and each hydraulic actuator,
In a control device for a load sensing control hydraulic circuit comprising a flow control valve that controls a flow rate of pressure oil supplied to a hydraulic actuator according to an operation signal of an operation means, a rotation speed for setting a target rotation speed Nθ of the prime mover. a setting means, a first detection means for detecting a differential pressure between the discharge pressure of the hydraulic pump and a load pressure of the actuator, a second detection means for detecting the discharge pressure of the hydraulic pump, and the first detection means. a first means for calculating a differential pressure target discharge amount QΔp of a hydraulic pump that maintains the differential pressure constant from a differential pressure signal of the means; and at least a pressure signal of the second detecting means and a preset hydraulic pump a second means for calculating an input limited target discharge amount QT of the hydraulic pump from an input limiting function; and a second means for calculating an input limited target discharge amount QT of the hydraulic pump; Qo, and a third means for controlling the discharge amount of the hydraulic pump so as not to exceed the input limit target discharge amount QT; and when the input limit target discharge amount QT is selected by the third means; , a fourth means for calculating the corrected rotation speed Nns of the prime mover based on at least the differential pressure target discharge amount QΔp and the input limited target discharge amount QT, and the corrected rotation speed Nns is calculated by the fourth means. If not, the rotation speed of the prime mover is controlled based on the target rotation speed Nθ set by the rotation speed setting means, and when the corrected rotation speed Nns is calculated, the target rotation speed Nθ and the correction rotation speed Nns are
and a fifth means for controlling the rotational speed of the prime mover to increase based on the following.

Further, in order to achieve the above object, according to the second invention, a third detection means for detecting the operation amount of the operation means is further provided, and in place of the fifth means, the fourth means is used. When the corrected rotation speed Nns is not calculated, or when the operation amount of the operation means detected by the third detection means is less than or equal to a predetermined value, the target rotation speed set by the rotation speed setting means is set. The rotational speed of the prime mover is controlled based on Nθ, and when the corrected rotational speed Nips is calculated by the fourth means and the operation amount of the operating means exceeds a predetermined value, the target rotational speed Nθ and the fourth
There is provided a control device for a load-sensing hydraulically driven turntable, comprising a sixth means for controlling the rotational speed of the prime mover to increase based on the corrected rotational speed Nns calculated by the calculation means. Ru.

Furthermore, in order to achieve the above object, according to a third aspect of the present invention, a seventh aspect of the present invention calculates a target rotational speed Nl of the prime mover according to the operation amount of the operation means detected by the third detection means.
further comprising means, in place of the fifth means or sixth means, when the corrected rotational speed Nns is not calculated by the fourth means, or when the corrected rotation speed Nns is detected by the third detection means. When the amount of operation of the operating means is less than a predetermined value, a value that is greater than 20 of the target rotation speed Nθ set by the target rotation speed setting means and the target rotation speed NIl calculated by the seventh means is selected. , the rotational speed of the prime mover is controlled based on this value, and the fourth means adjusts the corrected rotational speed Nns.
is calculated and the operation amount of the operating means exceeds a predetermined value, the target rotational speed Nθ set by the target rotational speed setting means, the target rotational speed NI calculated by the seventh means, and the fourth There is provided an eighth control hydraulic circuit control device that controls the rotation speed of the prime mover to increase based on the corrected rotation speed Nn@ calculated by the calculation means.

[Effect]

In the control device for the load sensing control hydraulic circuit according to the first aspect of the present invention, when the differential pressure target discharge amount QΔp is selected as the discharge amount target value Qo by the third means, the discharge pressure of the hydraulic pump and the load pressure of the hydraulic actuator Or the differential pressure between the maximum load pressure of multiple actuators is the differential pressure target discharge amount Q
The discharge amount of the hydraulic pump is controlled by load sensing so that Δ·p. At this time, in the fourth means, since the input limit target discharge amount QT is not selected in the third means, the correction rotation speed Nns is not calculated, and the fifth means calculates the target rotation speed set in the rotation speed setting means. The rotational speed of the prime mover is controlled based on the number Nθ.

On the other hand, when the input limited target discharge amount QT is selected as the discharge amount target value Qo by the third means, the discharge amount of the hydraulic pump is input limited controlled based on the input limited target discharge amount QT.

That is, the discharge amount of the hydraulic pump becomes saturated. At this time, the fourth
In the means, since the input limit target discharge amount QT is selected by the third means, the corrected rotation speed Nns is calculated, and the fifth means calculates the correction rotation speed Nθ and the target rotation speed Nθ set by the rotation speed setting means. The rotation speed of the prime mover is controlled to increase based on the number Nns. This increases the pump discharge amount and eliminates the saturation of the discharge amount.

In this way, in the first aspect of the present invention, the rotation speed of the prime mover automatically increases when the discharge amount of the hydraulic pump is saturated, so it is possible to set the target rotation speed Nθ to a low rotation speed, and the target rotation speed Nθ By setting the speed to low speed, fuel efficiency is reduced.
Not only can noise be reduced, but also the durability of equipment can be improved.

Further, even if the discharge amount of the hydraulic pump 22 is saturated, a desired actuator speed can be obtained and a reliable operation can be performed in a compound operation.

In the load sensing control hydraulic circuit control device according to the second aspect of the present invention, the sixth means increases the rotation speed of the prime mover when the corrected rotation speed Nns is calculated and the operation amount of the operating means exceeds a predetermined value. Since the control is performed, the same effect as the first aspect of the present invention can be obtained, and even if the corrected rotation speed N++s is calculated, the prime mover is controlled based on the target rotation speed Nθ in the range where the operation amount of the operating means is below a predetermined value. Therefore, even if a large load is applied to the actuator and the pump discharge amount is input-limited and saturated, the rotation speed of the prime mover does not increase, and fluctuations in the rotation speed of the prime mover during fine operation are prevented. This can be prevented and improves operability.

In addition, since there is no fluctuation in the rotational speed of the prime mover during slight operation, it is not unpleasant to the user's ears, and furthermore, it is possible to prevent deterioration of fuel efficiency due to frequent fluctuations in the rotational speed of the prime mover. Furthermore, when the manipulated variable exceeds a predetermined value, the rotational speed of the prime mover increases, so that the drive state of the prime mover matches the operator's intention, providing a good operating feeling.

In the load sensing control hydraulic circuit control device according to the third aspect of the present invention, the eighth means increases the rotation speed of the prime mover when the corrected rotation speed Nns is calculated and the operation amount of the operating means exceeds a predetermined value. Since the control is performed, the first and second
The same effect as the present invention can be obtained, and even if the corrected rotation speed Nns is calculated, the operation amount of the operating means is within a predetermined value or less, and the target rotation speed Nθ is the target calculated by the seventh means. When the rotational speed is smaller than NA', the prime mover is controlled based on the target rotational speed Nθ, so that the same effect as in the second aspect of the present invention can be obtained. Further, when the corrected rotational speed Nns is not calculated, or even if the corrected rotational speed Nns is calculated, the operation amount of the operating means is within a predetermined value or less, and the target rotational speed Nθ is greater than the target rotational speed Nl. Sometimes, the rotation speed of the prime mover is controlled based on the target rotation speed NA'. Therefore, the rotational speed of the prime mover is controlled in conjunction with the operating amount of the operating means that controls the flow rate control valve, so it is similar to the invention described in Patent Application No. 24, No. 24 of 1983, filed by the present applicant. It becomes possible to control the prime mover using both the rotation speed setting means and the operation means, and it is possible to set the target rotation speed Nθ low to improve fuel efficiency, and to obtain a powerful operation feeling proportional to the amount of operation of the operation means. This improves operability.

〔Example〕

Hereinafter, preferred embodiments of the present invention will be described with reference to the drawings.

First Embodiment A first embodiment of the present invention will be explained with reference to FIGS. 1 to 12.

In FIG. 1, the port-sensing control hydraulic circuit related to the control device of this embodiment includes, for example, a swash plate type variable displacement hydraulic pump 1, and first and second pumps driven by pressure oil from the hydraulic pump 1. 2 hydraulic actuators 2 and 3, and a first actuator that is arranged between the hydraulic pump 1 and the first actuator 2 and controls the flow rate and flow direction of the pressure oil supplied from the hydraulic pump 1 to the first actuator. Flow control valve 4
and a second pressure compensation valve 6, which is disposed between the hydraulic pump 1 and the 225th actuator 3, and controls the flow rate and flow direction of the pressure oil supplied from the hydraulic pump 1 to the second actuator. It has a flow rate control valve 5 and a second pressure compensation valve 7.

In this embodiment, the hydraulic circuit is applied to the wheeled hydraulic excavator shown in FIG.
is, for example, a travel motor that drives the rear wheels 50 via a transmission 51 and a propeller shaft 52, and the second actuator 3 is, for example, a boom cylinder that raises and lowers a poom 53 that is a part of the front attachment.

Returning to FIG. 1, the first pressure compensation valve 6 is connected to the hydraulic pump 1 via a pressure oil supply pipe 20 on its inlet side, and connected to the flow rate control valve 4 via a check valve 22 on its outlet side. Connected. The flow control valve 4 has a pressure compensation valve 6 on its inlet side.
and to the tank 10 via a return line 24, and the outlet side is connected to the first actuator 2 via a main line 25,26. In the main pipe 25.26,
26 Corresponding to the fact that the first actuator 2 is a travel motor, a counterbalance valve 54, a relief valve 55a,
55b, a replenishment circuit 56 is provided.

The second pressure compensation valve 7 is connected to the hydraulic pump 1 via a conduit 21 and a pressure oil supply conduit 20 on the artificial side, and is connected to the flow rate control valve 5 via a check valve 23 on the outlet side. The inlet side of the flow control valve 5 is connected to the pressure compensation valve 7 and to the tank 10 via a return line 29, and the outlet side is connected to the second actuator 3 via a main line 27.28. .

The pressure compensation valve 6 has two closing direction operating pilot pressure chambers 6
A, 6b and one release direction operation pilot chamber 6c opposite thereto.
One of the two closing direction operating pilot pressure chambers 6a, 6b 6
The inlet pressure of the flow rate control valve 4 is applied to the pressure chamber 6b via the conduit 30, and the control pressure from the electromagnetic proportional valve 9, which will be described later, is applied to the other pressure chamber 6b via the conduit 31. The operating pilot pressure chamber 6C includes a flow control valve 4 and a first
Pressure between the actuator 2 and the actuator 2 is applied via the conduit 32a. Further, the pressure compensating valve 6 is provided with a spring 6d that always biases the valve in the releasing direction.

The pressure compensating valve 7 is similarly constructed. That is, the pressure compensation valve 7 has two closing direction operating pilot pressure chambers 7a and 7b.
and one release direction operation pilot chamber 7c opposite thereto.
One of the two closing direction operating pilot pressure chambers 7a and 7b, 7a, has a
The inlet pressure of the flow control valve 5 is applied via the conduit 33;
The output pressure of the electromagnetic proportional valve 9 is connected to the other pressure chamber 7b through the conduit 34.
is applied through the release direction operating pilot pressure chamber 7c.
The pressure between the flow control valve 5 and the second actuator 3 is applied via the conduit 35a.

Further, the pressure compensating valve 7 is provided with a spring 7d that always biases the valve in the releasing direction.

The operation of the pressure compensation valve 6 will be briefly explained. When the control pressure from the electromagnetic proportional valve 9 is 0 (zero), the pressure compensation valve 6 allows the inlet pressure of the flow rate control valve 4 to be lowered from the pipe line 30 to the pilot chamber 6.
The outlet pressure of the flow rate control valve 4 is introduced into the pilot chamber 6c from the pipe line 32a, and is biased by the spring 6d. Therefore, the pressure compensation valve 6 always controls the flow rate from the hydraulic pump 1 so that the differential pressure between the inlet pressure and the outlet pressure of the flow rate control valve 4 is constant at the equivalent pressure of the spring 6d. Therefore, the flow rate flowing through the flow rate control valve 4 is between the discharge line 20 of the hydraulic pump 1 and the main line 25 or 26 of the hydraulic actuator 2.
Pressure compensation control is performed so that it does not change even if the pressure difference changes. The pressure compensation valve 7 operates similarly.

When the control pressure is output from the electromagnetic proportional valve 9, this pressure is applied to the pressure chambers 6b, 7b via the conduits 31, 34, and acts in a direction to cancel the urging force of the opposing springs 6d, 7d. As a result, the pressure difference between the inlet pressure and the outlet pressure of the flow control valves 4 and 5 is controlled to decrease in proportion to the increase in control pressure, and the flow rate flowing through the flow control valves 4 and 5 decreases.
Total consumable flow rate correction control is performed to limit the total flow rate of the pressure oil consumed by the actuator 2.3 without changing the distribution ratio between the two.

In the illustrated embodiment, the flow rate control valves 4 and 5 are configured in a hydraulic pipe 29 lot operation type, and are connected to pilot lines 36a and 3.
6b and 37a. It has a pilot chamber connected to 37b and is controlled by pilot pressure transmitted to the pilot line.

Pilot line 36a. 36b is connected to a pilot circuit 57 for running. Pilot circuit 5 for driving
7 has a pedal-operated pilot pressure reducing valve 58 for instructing the traveling speed, and a manually-operated direction switching valve 59 for instructing forward or reverse travel. Pilot pressure reducing valve 58
When the directional control valve 59 is switched to either position, the pilot pressure reducing valve 58 switches the pilot pump 60
A pilot pressure corresponding to the amount of depression is generated based on the pressure oil from the directional control valve 59.
6a, 36b, and the flow rate control valve 4 is switched accordingly. When the pilot pressure reducing valve 58 is released, the pilot pressure is reduced to the slow return valve 61.
The pressure is gradually reduced through the flow rate control valve 4, and the flow rate control valve 4 also gradually returns to the neutral position 30.

The pilot pipes 37a and 37b are connected to a pilot circuit operated by an operating lever (not shown) that determines the speed and driving direction of the second actuator, that is, the boom cylinder 3, and the operation occurs depending on the operating amount and operating direction of the operating lever. Pilot pressure is transmitted to any of these pilot lines to control switching of the flow control valve 5.

The flow rate control valve 4 and the pressure compensation valve 6 are combined together to
The flow rate control valve 5 and the pressure compensation valve 7 are combined to form one flow rate control valve with pressure compensation.

Flow control valve 4, . 5 are connected to pilot lines 32 and 35 for detecting the load pressures of the first and second hydraulic actuators 2 and 3, respectively.
．． 35 is further connected to a pipe 38 via a high pressure selection valve 12, and the load pressure on the high pressure side selected by the high pressure selection valve 12 (hereinafter referred to as maximum load pressure) is connected to the differential pressure gauge 4 via the pipe 38.
3 led to 31. The discharge pressure of the hydraulic pump 1 is also introduced to the differential pressure gauge 43 via the conduit 39 . Differential pressure gauge 43 is hydraulic pump 1
The differential pressure between the discharge pressure and the maximum load pressure is detected and a differential pressure signal ΔP is output.

The pressure oil supply pipe 20 of the hydraulic pump 1 includes a pressure detector 14 that detects the discharge pressure of the hydraulic pump 1 and outputs a pressure signal P.
The hydraulic pump 1 is provided with a tilt angle meter 15 that detects the tilt angle of a variable displacement mechanism such as a swash plate and outputs a tilt angle signal Qθ.

The discharge amount of the hydraulic pump 1 is controlled by a discharge amount control device 16 linked to the displacement variable mechanism. The discharge amount control device 16 is configured, for example, as an electro-hydraulic servo hydraulic drive device, as shown in FIG.

That is, the discharge amount control device 16 controls the displacement variable mechanism 1 of the variable displacement hydraulic pump 1, which is composed of a swash plate, an oblique shaft, or the like.
6a, and the servo piston 16b is housed in a servo cylinder 16c. The cylinder chamber of the servo cylinder 32 16c is divided into a left chamber 16d and a right chamber 16e by the servo piston 16b, and the cross-sectional area D of the left chamber 16d is larger than the cross-sectional area d of the right chamber 16e.

The left chamber 16d and the right chamber 16e of the servo cylinder 16c are connected to each other via conduits 16f and 16i, and the conduit 16
A pilot pump 8 is connected to i, a solenoid valve 16 is interposed between the conduit 16f and the conduit 16i, and the conduit 16f is further connected to the tank 10 by a conduit 16j equipped with a solenoid valve 16h.
is in contact with. The electromagnetic valves 16g and 16h are normally closed valves (a function that returns to the closed state when no electricity is applied), and are each excited (turned on) and switched by a load sensing control signal Q'0 from the control unit 40, which will be described later. It will be done.

When the solenoid valve 16g is excited (turned on) and switched to the switching position B, the left chamber 16d of the servo cylinder 16c communicates with the hydraulic pressure source 8, and the servo piston 16b moves to the third position due to the difference in area between the left chamber 16d and the right chamber 16e. Move to the right as shown in the diagram. As a result, the tilting angle of the variable displacement mechanism 16a of the hydraulic pump 1 increases, and the discharge amount increases. Solenoid valve 16g and solenoid valve 16h
is demagnetized (turned off) and both return to the switching position A, the oil passage of the left chamber 16d is cut off, and the servo piston 16
b is held stationary at that position. As a result, the tilt angle of the variable displacement mechanism 16a of the hydraulic pump 1 is kept constant, and the discharge amount is kept constant. Solenoid valve 16
When h is excited (turned on) and switched to the switching position B, the left chamber 16d and the tank 10 communicate with each other, the pressure in the left chamber 16d decreases, and the servo piston 16b is moved by the pressure in the right chamber 16e as shown in FIG. moved to the left. As a result, the tilting angle of the variable displacement mechanism 16a of the hydraulic pump 1 is reduced, and the pump discharge amount is also reduced.

In this way, the solenoid valve 16g. 16h and the tilt angle of the hydraulic pump 1, the pump discharge amount is controlled to be equal to the target discharge amount Qo calculated by the control unit 40 as described later.

34 Returning to FIG. 3 again, the hydraulic pump 1 is driven by the engine 63. The engine 63 is preferably a diesel engine equipped with a fuel injection device 64 with an all-speed governor, and the fuel injection device 64 has a governor lever 65, and by rotating the governor lever 65, the fuel injection amount is adjusted. The governor lever 65 is driven by a pulse motor 66, and the rotation amount of the pulse motor 66, that is, the displacement amount of the governor lever 65 is detected as the rotation speed of the engine 63 by a potentiometer 67, and the potentiometer 67 outputs an actual rotation speed signal Nrp.

Further, the engine 63 is provided with a rotation speed setting device 68 that is manually operated by an operator, includes an operating device for setting a target rotation speed of the engine 63, and outputs a target rotation speed signal Nθ.

Preferably, the rotation speed setting device 68 is configured such that the maximum value of the target rotation speed Nθ set thereby is equal to or lower than the maximum rotation speed of the engine 63. As the operating device, any device can be used, such as a potentiometer that outputs an electric signal according to the displacement amount 35 of the operating lever, a dial-type potentiometer, a push-button-type up/down switch, etc.

Pressure signal P1 from pressure detector 14 Tilt angle signal Qθ from tilt angle meter 15, differential pressure signal ΔP1 from differential pressure gauge 43 Actual rotational speed signal Nap from potentiometer 67, and target from rotational speed setting device 68 The rotational speed signal Nθ is input to the control unit 40, and the control unit 40 creates a load sensing control signal QlO, an engine control signal N'rθ, and a total consumable flow rate correction control signal Qns based on these input signals. It is output to the discharge amount control device 16, the pulse motor 66, and the electromagnetic proportional valve 9.

The control unit 40 is composed of a microcomputer,
FIG. 4 shows an outline of its configuration.

That is, the control unit 40 controls the pressure signal P1 and the tilt angle signal Q.
θ, differential pressure signal ΔP1, actual rotational speed signal Nrp, and target rotational speed signal Nθ, A/D converter 4 converts into digital signals.
0a, a central processing unit 40b, a memory 40c for storing control procedure programs, 36, an output D/A converter 40d, an output I/O interface 40e, and an electromagnetic proportional valve 9. amplifier 4
0f and solenoid valve 16g. Amplifier 40 connected to 16h
g, 40h, and the amplifier 4 connected to the pulse motor 66.
0i.

Next, details of the operation procedure performed by the control unit 40 will be explained using the flowchart shown in FIG. I will explain.

First, in step 100, the pressure signal P1 from the pressure detector 14 is
Read the tilting angle signal Qθ from the tilting angle meter 15, the differential pressure signal ΔP from the differential pressure gauge 43, the actual rotational speed signal N+p from the potentiometer 67, and the target rotational speed signal Nθ from the rotational speed setting device 68, Remember.

Subsequently, in step 101, the input limit target discharge amount QT is determined from the pressure signal P and the input torque limit function f (P) inputted in advance. FIG. 6 shows the input torque limit function. In FIG. 6, the horizontal axis is the discharge pressure P, and the vertical axis is the input limit target discharge amount QT based on the input torque limit function f (P). The input torque of the hydraulic pump 37 is proportional to the product of the tilting amount Qθ of the hydraulic pump 1 and the discharge pressure P. Therefore, the input torque limiting function f (
P) uses a hyperbola or an approximate hyperbola.

That is, r However, TP: Input limit torque κ: proportionality constant It is a function expressed by the formula.

The input limit target discharge amount QT can be determined from this input torque limit function f (P) and the discharge pressure P.

Returning to FIG. 6 again, in step 102, the differential pressure, that is, the differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the actuator 2.3 is determined from the differential pressure signal ΔP. The differential pressure target discharge amount QΔp to be kept constant is determined.

An example of how to obtain this will be explained with reference to FIG. FIG. 7 is a block diagram showing a method for determining the differential pressure target discharge amount QΔp from the differential pressure signal ΔP of the differential pressure gauge 43. In this example, the differential pressure target discharge amount QΔp is determined based on the following formula. Desired.

QΔp=g(Δp)=ΣK+(ΔPO−ΔP)=K! (
ΔPo−ΔP) +Qo−1ΔQΔp+Qo−1
...(2) However, KI: Integral gain ΔPa ・Target differential pressure Qo-1: Discharge amount target value output in the previous control cycle ΔQΔp: Differential pressure in control 1 cycle time Increment of target discharge amount, that is, differential pressure This is an example in which the target discharge amount QΔp is calculated by an integral control method of the deviation between the target differential pressure ΔPO and the actual differential pressure, and in FIG. 7, block 120 calculates Kl(ΔP
o - ΔP) to find the increment ΔQΔp of the differential pressure target discharge amount per one control cycle time, and in block 121, this ΔQΔp and the discharge volume target value Qo−1 output in the previous control cycle are calculated. By adding these, equation (2) is obtained.

In this example, QΔp was determined by an integral control method of ΔPO−ΔP, but a different method, for example, QΔp
= Kp (ΔPo - ΔP) ・(3) However, Kp is the proportional control method expressed by proportional gain, or the equation (2) and (3)
It may also be determined by employing a proportional/integral control method in which equations are added.

Returning to FIG. 5 again, in step 103, step 10l. ,1
The target discharge amount deviation ΔQ between the differential pressure target discharge amount QΔp obtained in step 02 and the input limit target discharge amount QT is determined, and the sign of the deviation ΔQ is determined in step 104. If positive, the process proceeds to step 105 to determine the discharge amount. QT is selected as the target value Qo, and if it is negative, the process proceeds to step 106 and QΔp is selected as the discharge amount target value Qo.

That is, the differential pressure target discharge amount QΔp and the input limit target discharge amount QT
The smaller one is selected as the discharge amount target value Qo, and the discharge amount target value Qo is prevented from exceeding the input limit target discharge amount QT determined by the input torque limit function f(P).

Next, the process moves to step 107. In step 107, the corrected rotation speed Nns of the engine 4063 is calculated from the target discharge amount deviation ΔQ obtained in step 103. An example of how to obtain this will be explained with reference to FIG. FIG. 8 is a block diagram showing a method for calculating the corrected rotational speed Nns from the target discharge amount deviation ΔQ. In this example, the corrected rotational speed Nns is determined by an integral control method based on the following equation.

Nns=h(ΔQ)=ΣKIns ・ΔQK Ins
ΔQ + N ns-1ΔN ns+ N ns-
1-(4) However, Kins: Integral gain Nns-1: Corrected rotational speed Nns output in the previous control cycle ΔNns: Increment of the corrected rotational speed in one control cycle time, that is, in Fig. 8, first obtained in step 103 From the target discharge amount deviation ΔQ, in block 130, a corrected rotational speed increment Δ'Nns per one control cycle time, that is, K lns
・Find ΔQ. Then, in the adder 131, this value is added to the corrected rotation speed Nns-1 output in the previous control cycle to obtain an intermediate value N'ns. 0, N
When 'ns≧0, a correction rotation speed Nns that increases in proportion to the increase in N'ns is output, and when N'ns≧Q'ns
The corrected rotational speed N++s is determined so that Nns=Nnsmax at c.

In this example, the corrected rotational speed Nns was determined using an integral control method, but similarly to the differential pressure target discharge amount QΔp described above, the relationship between Nns and ΔQ may also be determined using a proportional control method or a proportional-integral control method. good.

Next, in step 108, the corrected rotation speed Nns and the target rotation speed N
Determine the size of o, and if Nns≧Nθ, step 109
The process proceeds to step 110 and selects Nns as the final target rotational speed Nr θ, and if Nns<Nθ, the process proceeds to step 110 and selects Nθ as the final target rotational speed Nr θ. That is,
The larger of Nns and Nθ is selected.

Next, the process moves to step 111. In step 111, a total consumable flow correction value Qns for controlling the electromagnetic proportional valve 9 is calculated from the target discharge amount deviation ΔQ obtained in step 103. An example of how to obtain this is explained in FIG. 9. FIG. 9 is a block diagram showing a method for calculating the total consumable flow rate correction value Qns from the target discharge amount deviation ΔQ. As can be seen from this figure, in this example, the correction value Qn
Similarly to the corrected rotational speed Nns, s is determined by an integral control method using a block 133, an adder 134, and a limiter 135. Note that in this case as well, the correction value Qns may be obtained using the proportional control method or the proportional/integral control method.

Returning to FIG. 5 again, in step 112, the product of the tilt angle signal Qθ, which is the output of the tilt angle meter 15, and the actual rotational speed signal Nrp, which is the output of the potentiometer 67, is further multiplied by a coefficient K to obtain the actual pump discharge amount. Q is determined, and a control signal Q I O of the discharge amount control device 16 is generated from the discharge amount target value Qo of the hydraulic pump 1 obtained in steps 105 and 106 and this actual pump discharge amount Q, and the control signal Q I o is output to the discharge amount control device 16 via the 1/O interface 40e and amplifiers 40g and 40h of the control unit 40 shown in FIG. The tilting amount Qθ of the pump 1 is controlled.

43 FIG. 10 shows a flowchart of the operation procedure performed in step 112 above. First, in step 139, Q=K−Q
Calculate θ·Nrp to find the actual pump discharge amount Q. Next, in step 140, z=QoQ is calculated, and the deviation 2 between the discharge amount target value Qo and the actual pump discharge amount Q is determined. Next, in step 141, the magnitude of the absolute value of deviation 2 and the value Δ2 that defines a preset dead zone is determined. If the absolute value of deviation Z is larger than the set value Δ2, the process proceeds to step 142, and the sign of deviation 2 is determined. Determine. If deviation 2 is positive here, step 14
3, a control signal Q'0 is output to turn on the solenoid valve 16g of the discharge amount control device 16 and turn off the solenoid valve 16h. As a result, as described above, the tilting angle of the hydraulic pump 1 increases, and the tilting angle is controlled so that the actual pump discharge amount Q matches the discharge amount target value Qo. If deviation 2 is negative, proceed to step 144, turn off solenoid valve 16g, and turn off solenoid valve 1.
Outputs a control signal Q lo that turns ON 6h. As a result, the tilting angle of the hydraulic pump 1 is reduced, and the actual pump discharge amount Q is controlled 44 to match the discharge amount target value Qo.

If the absolute value of the deviation Z is smaller than the set value Δ2 in step 141, the process proceeds to step 145, where both the solenoid valves {6g and 16h are turned off. As a result, the tilt angle of the hydraulic pump 1 is maintained, and the pump discharge amount is controlled to be constant.

As a result of controlling the discharge amount of the hydraulic pump 1 as described above,
In step 106, the differential pressure target discharge amount QΔp becomes the discharge amount target value Qo.
When selected, the discharge amount of the hydraulic pump 1 is controlled to be the differential pressure target discharge amount QΔp, load sensing control is performed to maintain the differential pressure between the pump discharge pressure and the maximum load pressure constant, and step 105 When the input limit target discharge amount QT is selected as the discharge amount target value Qa, the input torque limit control is performed so that the discharge amount of the hydraulic pump does not exceed the input limit target discharge amount QT.

Returning to FIG. 5 again, in step 113, the actual rotational speed signal N+p, which is the output of the potentiometer 67, and steps 109, 1
From the target rotation speed Nrθ obtained in step 10, the pulse motor 66
A control signal N'+p 45 is generated, and the control signal N'rp is sent to the I/O interface 40e and amplifier 40i of the control unit 40 shown in FIG.
output to the pulse motor 66 via the engine speed N
Control is performed so that rp becomes the target rotational speed N+θ.

FIG. 11 shows a flowchart of the operation procedure performed in step 113 above. First, in step 150, A=Nrp
−Nr θ is calculated, and the actual rotation speed Ntp and target rotation speed Nr are calculated.
Find the deviation A of θ. Next, in step 151, the magnitude of the absolute value of the deviation A and the value ΔA that defines a preset dead zone is determined. If the absolute value of the deviation A is larger than the set value ΔA, the process proceeds to step 152, and the sign of the deviation A is determined. Determine. If the deviation A is positive here, the process proceeds to step 153, where a control signal N'rθ for driving the pulse motor 66 in the normal rotation direction is output. As a result, the fuel injection amount of the fuel injection device 64 increases,
The actual rotation speed Nrp is controlled to match the target rotation speed N+θ. If the deviation A is negative, the process proceeds to step 154 and outputs a control signal N'tθ for driving the pulse motor 66 in the reverse direction. As a result, the fuel injection amount of the fuel injection device 64 decreases, and the actual rotational speed N+p
is controlled so that it matches the target rotational speed N+θ.

If the absolute value of the deviation A is smaller than the set value ΔA in step 151, the process proceeds to step 155 and the output of the control signal is stopped. As a result, the rotation of the pulse motor 66 is stopped, and the engine speed is maintained constant.

Returning again to FIG. 5, in step 114, the correction value Qns is output as a control signal to the electromagnetic proportional valve 9 via the D/A converter 40d and amplifier 40f of the control unit 40 shown in FIG. The pressure compensation valves 6 and 7 shown in FIG. As a result, when the differential pressure target discharge amount QΔp is smaller than the input limit target discharge amount QT, block 13 of step 111
5 (FIG. 9), the correction value Qns becomes O, and the pressure compensation valve 6
, 7 perform pressure compensation control as set by the springs 6d and 7d, and when the differential pressure target discharge amount QΔp becomes larger than the input limit target discharge amount QT, the correction value Qns is set in step 111.
increases as the target discharge amount deviation ΔQ increases with Qnsmax being the maximum value, weakening the control force of the springs 6d and 7d of the pressure compensation valves 6 and 7 in the valve opening direction.

As a result, the set differential pressure between the pressure compensation valves 6 and 7 is reduced, and total consumable flow rate correction control is performed.

Of the above control procedures, the control system for the hydraulic pump 1 and the control system for the engine 63 are summarized in a control block diagram and shown in FIG. 12. In the figure, block 200 corresponds to step 101 in FIG. 5, and calculates the input limit target discharge amount QT using the input torque limit function shown in FIG.
2 and 203 correspond to the step 102, among which the addition block 201 and the proportional calculation block 202 correspond to the differential pressure target discharge amount increment calculation section 120 in FIG.
3 corresponds to the adder 121, and the differential pressure target discharge amount QΔp is calculated by these blocks. Block 204 is the fifth
Corresponding to steps 104, 105, and 106 in the figure, the smaller of the two target discharge amounts QT and QΔp is the discharge amount target value Q.
selected as a.

Blocks 205, 206, 207, and 208 correspond to step 107 in FIG.
5 and the proportional calculation block 206 correspond to the correction rotation speed increment calculation section 130 in FIG. . block 213
corresponds to steps 108, 109, and 110 in FIG.
is a servo control block corresponding to step 113, in which the pulse motor 66 is controlled based on the target rotation speed Nrθ and the actual rotation speed Nrp.

Blocks 209, 210, 211, and 212 correspond to step 112 in FIG.
0, the addition block 209 corresponds to the procedure 140, and the blocks 210.211 correspond to the procedure 141.
-145, and output control signals Q'o to the solenoid valves 16g and 16, respectively.

Next, the operation of this embodiment configured as above will be described in 49. explain.

Assume that the rotation speed setting device 68 is now in a state of setting a target rotation speed Nθ smaller than the maximum rotation speed of the engine 63. In such a state, when the operator switches the directional control valve 59 to the forward position and depresses the pilot pressure reducing valve 58 with the intention of driving the vehicle forward alone, the switching position of the directional control valve 59 and the amount of depression of the pilot pressure reducing valve 58 change. Accordingly, the flow control valve 4 is switched, and pressure oil is supplied to the first actuator, ie, the travel motor 2, accordingly.

At this time, when the load applied to the travel motor 2 is relatively small, and the amount of depression of the pilot pressure reducing valve 58 is small, such as when traveling at low speed on a flat road, and the required flow rate of the flow rate control valve 4 is small, the present embodiment In the control device, a value larger than the differential pressure target discharge amount QΔp is calculated as the input limit target discharge amount QT, the differential pressure target discharge amount QΔp is selected as the discharge amount target value Qo, and the pump discharge pressure and the running load pressure are Hydraulic pump 1 is controlled by pressure sensing so that the differential pressure between the two is constant. Therefore, the vehicle can travel forward at a desired speed depending on the amount of depression. At this time, the final target rotation speed N ``θ
is the target rotation speed N set by the rotation speed setting device 68.
θ has been selected, and the engine 63 is controlled to a constant speed corresponding to its target rotation speed.

When the required flow rate of the flow control valve 4 becomes significantly large, or when the load pressure increases and the pump discharge pressure increases, such as when driving at high speeds or when driving on a slope, the input limit target discharge amount A value smaller than the differential pressure target discharge amount QΔp is calculated as QT, and the input limit target discharge amount QT is selected as the discharge amount target value Qo. For this reason, the discharge amount of the hydraulic pump is controlled to limit the input torque, and tends to be saturated.

At this time, the target discharge amount deviation ΔQ is simultaneously calculated as a positive value, and the corrected rotation speed Nns is calculated based on this. This corrected rotation speed Nns increases while the target discharge amount deviation ΔQ is calculated as a positive value, and eventually becomes larger than the target rotation speed Nθ set by the rotation speed setting device 68. For this reason, the engine 63
The final target rotational speed N' θ is the corrected rotational speed Nns
is selected, and the engine speed increases. As a result of the engine speed increasing in this way, the discharge amount of the hydraulic pump 1 increases, the saturation of the pump discharge amount is eliminated, and the pump discharge amount is adjusted so that the differential pressure between the pump discharge pressure and the running load pressure is constant. Even in this case, the vehicle can travel forward at a desired speed according to the amount of depression.

Note that when the target discharge amount deviation ΔQ is calculated as a positive value, the total consumable flow rate correction value Qns is also calculated and the pressure compensation valve 6 is attempted to be controlled, but due to the above-mentioned engine speed increase control. As a result of keeping the differential pressure between the pump discharge pressure and the running load pressure constant, the total consumable flow rate correction value Qn
s is no longer calculated, and control of the pressure compensation valve 6 is released.

Next, a case will be considered in which the target rotation speed No. is similarly set by the rotation speed setting device 68 to a value smaller than the maximum rotation speed of the engine 63 and a combined operation of travel and boom is performed. In this case, in the work mode where the amount of depression of the pilot pressure reducing valve 5528 is small, the amount of operation of the second actuator, that is, the operation lever (not shown) of the boom cylinder 3 is small, and the total required flow rate of the flow rate control valve 4.5 is small. In the control device of this embodiment, a value larger than the target discharge rate Qo is calculated as the input restriction target discharge rate QT, and the differential pressure target discharge rate QΔp is selected as the target discharge rate Qo, similar to the case of the driving-only operation. The hydraulic pump 1 is load-sensed controlled so that the differential pressure between the pump discharge pressure and the running load pressure is constant, and a desired combined operation can be performed.

If the total required flow rate of the flow control valves 4 and 5 increases or the load pressure of either of them increases significantly, the input limited discharge amount Q
When a value smaller than the differential pressure target discharge amount QΔpX is calculated as T, a value smaller than the differential pressure target discharge amount QΔp is calculated as the input limit target discharge amount QT. For this reason, the input limited target discharge amount QT is selected as the discharge amount target value Qo, and the discharge amount of the hydraulic pump is controlled to limit the input torque, and attempts to reach saturation, but at the same time the target discharge amount The amount deviation ΔQ is calculated as a positive value, and the corrected rotation speed Nr+s is calculated based on this,
Engine 63 is controlled to increase its rotational speed. Therefore, the discharge amount of the hydraulic pump 1 increases, and the differential pressure between the pump discharge pressure and the traveling load pressure is again maintained constant, and the combined operation of traveling and the boom can be continued in accordance with the respective operation amounts. I can do it.

When the total required flow rate of the flow control valves 4 and 5 further increases, or when the load pressure further increases and the engine 63 controlled by the corrected rotation speed Nns reaches the maximum rotation speed, the pump discharge amount also reaches the maximum possible discharge amount. This reaches the point where it becomes impossible to increase the pump discharge amount any further. In such a state, the differential pressure between the pump discharge pressure and the maximum load pressure is reduced by the input torque limit control of the hydraulic pump, the total consumable flow correction value Qns is calculated based on the target discharge amount deviation ΔQ, and the electromagnetic proportional valve 9
The control force in the closing direction of the pressure compensation valve 6.7 is increased by the control pressure from Throttled, flow control valves 4, 5
The diversion ratio of the flow rate is ensured. As a result, although the combined operation speed decreases as a whole, the drive speed ratio of the actuator is kept constant, so that smooth combined operation can be continued.

As explained above, according to this embodiment, the hydraulic pump 1
When the discharge amount is saturated, the rotation speed of engine 1 is automatically increased and the pump discharge amount is increased, so that the speed of the actuator decreases due to the saturation of the pump discharge amount until the maximum engine rotation speed is reached, and the operation of the combined operation. It is possible to prevent a decrease in performance and ensure excellent operability. Furthermore, since the target rotational speed Nθ of the engine 63 can be set to a low rotational speed, fuel consumption and noise can be reduced, and the durability of the equipment can be improved.

In addition, after the engine reaches the maximum rotation speed, the total consumable flow rate correction control is performed by controlling the pressure compensation valve.
Similarly, complex operations can be performed smoothly.

55 Modification of the First Embodiment A modification of the first embodiment will be explained with reference to the flowchart shown in FIG. In the figure, steps equivalent to those in the flowchart shown in FIG. 5 are given the same reference numerals. In this embodiment, control to increase the rotational speed of the engine 1 and control to correct the total consumable flow rate of the pressure compensating valves 6 and 7 are reliably performed separately.

In FIG. 13, the steps up to steps 105 and 106 for determining the discharge amount target value Qo are the same as in the first embodiment. After this, in this embodiment, the actual rotational speed of the engine 63 detected by the potentiometer 67 reaches the maximum rotational speed Nmax.
Determine whether it has been reached. Here, the maximum rotation speed Nm
If it is determined that ax has not been reached, the steps ↓07, 1 excluding steps 111 and 114 shown in FIG. 5 of the first embodiment are executed.
08. Proceed to 19, 110, 112A, 113. Thereby, control other than control of the pressure compensation valves 6 and 7, that is, load sensing control of the hydraulic pump 1 and control of increasing the rotational speed of the engine 63 is performed. Maximum rotation speed N in step 115
If it is determined that the maximum has been reached 56, steps 111 and 112 shown in FIG. 5 of the first embodiment are performed.
Proceeding to B, 114, control other than control to increase the rotational speed of the engine 63, that is, load sensing control of the hydraulic pump 1 and control of the pressure compensation valves 6 and 7 is performed.

According to this embodiment, the pressure compensating valves 6 and 7 are not controlled before the engine 63 reaches the maximum rotation speed.
Since the pressure compensation valves 6 and 7 are not controlled until the rotation speed of the engine 63 reaches the maximum, interference between the control to increase the rotation speed of the engine 63 and the control of the pressure compensation valves 6 and 7 can be prevented, resulting in a stable engine. 63 rotation speed increase control can be performed.

In the above embodiment, control of the pressure compensation valves 6 and 7 is started after the engine speed Nrp reaches the maximum speed Nmax, but when the engine speed Nrp reaches a predetermined value lower than the maximum speed Nmax, Control of the pressure compensation valves 6 and 7 may be started after reaching the value Nc. In this case, the engine speed is prevented from increasing unnecessarily, and noise and fuel consumption can be reduced.

57 Another modification of the first embodiment will be explained with reference to FIG. 14.

In this embodiment, switching between engine speed increase control and pressure compensation valve control is performed using an index other than the engine speed.

In FIG. 14, in step 100A, in addition to the pressure signal P1, the tilting angle signal Qθ, the differential pressure signal ΔP1, the actual rotational speed signal Nrp, and the target rotational speed signal Nθ, the pilot pressure of the actuator operating device, for example, the traveling pilot pressure pt, is Load. In this case, a pressure detector is connected to the outlet side of the pilot pressure reducing valve 58 as in the embodiment shown in FIG. 16, which will be described later.
Detects driving pilot pressure. And in step 116,
Determine whether P(>0, that is, whether the traveling pilot pressure PI is rising. If the traveling pilot pressure Pt is rising, steps 107, 108, 109, 110.112
Proceeding to A, 113, load sensing control of the hydraulic pump 1 and control of increasing the rotational speed of the engine 63 are performed. If the traveling pilot pressure PI is not set, follow steps 111 and 11.
Proceeding to step 2A, 114, load sensing control of the hydraulic pump 1 and control of the pressure compensation valves 6 and 7 are performed.

According to this embodiment, when the discharge amount of the hydraulic pump 1 is saturated, engine speed increase control is selected for driving, and flow compensation valve control is selected for actuator operations other than driving. For driving where there is relatively little fluctuation in flow rate and slow response of engine control is not a problem, sufficient flow rate can be supplied by controlling the increase in engine speed, and for other actuators. By performing total consumable flow rate correction control using highly responsive pressure compensation valve control without performing engine speed increase control, it is possible to perform optimal control that matches the characteristics of each actuator.

5 9 Second Embodiment A second embodiment of the present invention will be described with reference to FIGS. 15 to 17. In the drawings, the same reference numerals are given to the same members or procedures as those shown in FIGS. 1 to 12.

In FIG. 15, a pressure detector 62 is connected to the outlet side of the pilot pressure reducing valve 58 to detect the pilot pressure, that is, the operating amount of the pilot pressure reducing valve, and the pressure detector 62 outputs an operating amount signal X. This operation amount signal X is the control unit 40
The signal is input to an A/D converter 40a (see FIG. 4) which is an input section of A. The other hardware structure is the same as the embodiment shown in FIG.

FIG. 16 shows a flowchart of the control procedure program of this embodiment stored in the memory of the control unit 40A. For convenience of illustration, FIG. 16 shows only the procedures related to engine control, and the procedures related to load sensing control of the hydraulic pump 1 and control of the pressure compensation valves 6 and 7 are omitted.

In FIG. 16, in step 100B, pressure signal P160, tilting angle signal Qθ, differential pressure signal ΔP1, actual rotational speed signal Nrll
In addition to s and target rotational speed signal Nθ, pressure detection means 62
Read the manipulated variable signal X. Thereafter, steps 101 to 107 are the same as in the first embodiment shown in FIG. 16, and step 103 calculates the discharge amount deviation ΔQ,
At 05 and 106, the discharge amount target value Qo is calculated, and step 1
In step 07, the corrected rotation speed Nns is calculated.

Next, in this embodiment, in step 170, it is determined whether the operation amount X of the pilot pressure reducing valve 58 detected by the pressure detector 62 is larger than a predetermined value xk, and if it is determined that it is larger, the first Procedure 108, 1 similar to the example
09, 110, when the corrected rotation speed Nns becomes larger than the target rotation speed Nθ set by the rotation speed setting device 68, this value is selected as the final target rotation speed Nrθ,
The rotation speed is controlled to increase. If it is determined in step 170 that the manipulated variable X is less than or equal to the predetermined value xk, the process proceeds to step 110 and selects the set target rotation speed Nθ as the final target rotation speed Nrθ.

61 FIG. 17 is a control block diagram showing steps 170 and subsequent steps in the above flowchart. In the figure, blocks having the same functions as those shown in FIG. 12 are given the same reference numerals. The block 220 outputs a level 1 signal α when the manipulated variable X detected by the pressure detection means 62 exceeds a predetermined value xk, and the signal α is sent to the switch 221. The switch 221 is open when there is no signal α, and blocks the transmission of the corrected rotational speed Nns to the maximum value selection block 213, and is closed when the signal α is generated, and transmits the corrected rotational speed Nns to the maximum value selection block 213. introduce.

According to this embodiment configured as described above, when the operation amount X of the pilot pressure reducing valve 58 exceeds the predetermined value xk, the rotation speed increase control of the engine 63 is performed when the corrected rotation speed N++s is calculated. Therefore, in the range where the manipulated variable X exceeds the predetermined value xk, the same effects as in the t-th embodiment can be obtained. In addition, in the range where the manipulated variable X is less than or equal to the predetermined value xk,
Even if the corrected rotational speed NI has not been calculated, the engine 63 is controlled based on the target rotational speed Nθ. Therefore, when the operation amount of the pilot 62 pressure reducing valve 58 is reduced and the slow speed running is performed, the load is reduced at the time of starting. Therefore, even if the hydraulic pump 1 is temporarily controlled to limit the input torque, the engine speed is not controlled to increase, and fluctuations in the engine speed at the time of starting can be prevented, and operability is improved. Furthermore, since there is no change in the engine speed when running at a slow speed, it is not unpleasant to the driver's ears, and it is possible to prevent fuel consumption from deteriorating due to frequent fluctuations in the engine speed. Furthermore,
When the manipulated variable X exceeds the predetermined value xk, engine speed increase control is performed, so that the driving state of the engine matches the operator's intention and a good operational feeling can be obtained.

Third Embodiment A third embodiment of the present invention will be explained with reference to FIGS. 18 and 19. In the figures, the same reference numerals are given to the same members or procedures as those shown in FIGS. 1 to 12 and 15 to 17.

FIG. 18 is a flowchart showing the control procedure program of this embodiment, and like FIG. 6, only the procedure related to engine control is shown.

63 In FIG. 18, steps 100B to 107 are shown in step 16.
This is the same as the second embodiment shown in the figure. Next, in step 17, a second value that increases in accordance with the manipulated variable Target rotation speed N
Calculate I. Next, proceeding to step 170, the pressure sensor 6
It is determined whether the operating amount X of the pilot pressure reducing valve 58 detected in step 2 is larger than a predetermined value It is determined whether the rotation speed is larger than the target rotation speed Nl. If it is determined in step 172 that Nns≧Nl, steps 108 and 1 similar to the first embodiment are performed.
09, 110, and when the corrected rotation speed Nns becomes larger than the target rotation speed Nθ set by the rotation speed setting device 68, this value is selected as the final target rotation speed Nrθ, and rotation speed increase control is performed. Ru.

If it is determined in step 170 that the manipulated variable X is less than or equal to the predetermined value It is determined whether the target rotation speed NI is larger than the first target rotation speed Nθ, and Nl<No.
If it is determined that
θ is selected as the final target rotation speed Nr θ, and Nθ
Normal rotational speed control is performed based on the following. If it is determined in step 173 that Nl≧Nθ, the process proceeds to step 174, where the second target rotation speed In is selected as the final target rotation speed Nrθ, and the rotation speed of the engine 63 is controlled based on NI.

FIG. 19 is a control block diagram showing steps 171 and subsequent steps in the above flowchart. In the figure, blocks having the same functions as those shown in FIGS. 12 and 17 are given the same reference numerals. Block 223 inputs the operation amount X of the pilot pressure reducing valve 58 detected by the pressure detector 62,
This is a block that calculates the target rotation speed Nl that increases according to the manipulated variable X by the function of NA'=f(x).
5 are both sent to the maximum value selection block 224, the larger value of both is selected, and its output is sent to the maximum value selection block 21.
3 is output.

According to this embodiment configured as described above, in the range where the operation amount X of the pilot pressure reducing valve 58 exceeds the predetermined value xk, the corrected rotation speed Nns is calculated and the correction rotation speed Nns is
When the rotational speed is larger than the target rotational speed Nj2, control is performed to increase the rotational speed of the prime mover, so that the same effects as in the first embodiment can be obtained. In addition, in a range where the manipulated variable X is equal to or less than the predetermined value Since the control is performed, it is possible to obtain the same effects as in the second embodiment regarding slow speed running.

Furthermore, even when the corrected rotational speed Nns is not calculated or even when the corrected rotational speed Nns is calculated, the operation amount When it is larger than the number Nl, the rotational speed of the engine 63 is controlled based on the target rotational speed Nl. Therefore, the rotation speed of the engine 63 is controlled in conjunction with the operation amount of the pilot pressure reducing valve 58, which is an operating means for controlling the flow rate control valve 4.
Similar to the invention described in No. 150, both the rotation speed setting device 68 and the pilot pressure reducing valve 58 (operating means) are used to control the engine 63.
This makes it possible to set the target rotational speed Nθ low and improve fuel efficiency, and it is also possible to obtain a powerful operation feeling that is proportional to the amount of operation of the operating means, improving operability.

Fourth Embodiment A fourth embodiment of the present invention will be explained with reference to FIG. In the figures, members equivalent to those shown in FIGS. 1, 12, etc. are designated by the same reference numerals. In the first to third embodiments, the input limit target discharge amount QT was determined from the discharge pressure P and the input torque limit function f (P) shown in FIG. Input limit target value Q using
This is to determine T.

67 In FIG. 20, a block 250 displays the actual rotation speed Nrp detected by the potentiometer 67 and the rotation speed setting device 68.
This is an addition block that calculates the rotation speed deviation ΔN by comparing the target rotation speed Nθ set in The input limit target discharge amount calculation block 251 is inputted. As shown in the figure, in block 251, an input torque limiting function fl(P, ΔN) is preset with pump discharge pressure P and rotational speed deviation ΔN as parameters, and the input torque limiting function fl(P, ΔN)
is the target discharge amount QT as the rotation speed deviation ΔN increases.
The functional relationship is such that the product of P and discharge pressure P becomes small. In block 251, this input torque limiting function f
The input limit target discharge amount QT is calculated from l(P, ΔN). Other blocks are the same as those shown in FIG.

According to this embodiment, the input torque limit control 68 of the hydraulic pump 1 is performed so that the product of the target discharge amount QT and the discharge pressure P becomes smaller as the rotational speed deviation ΔN increases, so that stalling of the engine 63 is more effectively prevented. This makes it possible to use the engine's output horsepower to the maximum extent possible.

Fifth Embodiment A fifth embodiment of the present invention will be explained with reference to FIG. 21. In the figures, members equivalent to those shown in FIGS. 1, 12, etc. are designated by the same reference numerals. In this embodiment, when controlling the increase in engine speed, the discharge amount of the hydraulic pump is inputted to limit target discharge amount QT.
This is an example of preventing interference between load sensing control and engine speed increase control.

That is, in the embodiments shown in FIGS. 1 and 12, when the pump discharge amount is saturated and the differential pressure target discharge amount QΔp is larger than the input limit target discharge amount QT, the hydraulic pump 1
is held at the input limit target discharge amount QT, and the corrected rotation speed Nn
The pump discharge amount is increased by controlling the engine 63 based on s, and the saturated state is eliminated.

On the other hand, when the pedal of the pilot pressure reducing valve 69 to 58 is released and the operation amount X becomes smaller while the engine speed increase is controlled by the corrected rotation speed Nns, the flow rate passing through the flow rate control valve 4 decreases. A value larger than the differential pressure target discharge amount QΔp is calculated as the input restriction target flow rate QT, and the discharge amount of the hydraulic pump 1 is equal to the differential pressure target discharge amount QΔ
The discharge amount is controlled based on p. However, in the structure of the embodiments up to now, the value of the corrected rotation speed Nns also decreases at the same time, and the rotation speed of the engine 63 also decreases by 1.
Decrease.

In the process, when the discharge amount of the hydraulic pump 1 becomes smaller than the flow rate passing through the flow rate control valve 4, the differential pressure target discharge amount QΔp increases again and exceeds the input limit target discharge amount QT, and the input torque of the hydraulic pump ■ increases. As the restriction control is carried out, the corrected rotational speed Nns increases again, and the engine rotational speed increases. The above-mentioned process is repeated, and as a result, the load sensing control and the engine speed increase control may interfere with each other, resulting in a hunting phenomenon.

This embodiment has been developed to avoid the above-described hunting phenomenon.

In FIG. 21, block 300 is a block that determines whether or not engine 70 rotational speed increase control is being performed, and if so, sets a flag FNns. The determination is made based on the corrected rotation speed Qns,
If Nns is less than the predetermined value Nnsa near O, engine speed increase control is not performed and the predetermined value Nnsa
In the above case, it is determined that engine speed increase control is being performed. At this time, the flag FNns is set to 1 during engine speed increase control, and set to O when no control is being performed.
Block 204A is the minimum value selection block and is the 12th
Similar to block 204 in the figure, the magnitude of the input limit target discharge amount QT and the differential pressure target discharge amount QΔp is determined, and after selecting the smaller value, it is output as the discharge amount target Qor.

Block 301 is a discharge amount target value selection switch for hydraulic bomb l. Here, the flag FN of the corrected rotation speed Nns
ns, if FNns is O, select the discharge amount target value Qor selected by the minimum value selection block 204A, and select FNns.
When Nns is 1, the input limit target flow rate QT is selected and outputted as the discharge rate target value Qo.

Other blocks are the same as those shown in FIG.

Next, the operation of this embodiment will be explained. When the required flow rate of the flow rate control valves 4 and 5 is less than the input restriction target discharge amount QT, the differential pressure target discharge amount QΔp is smaller than QT, and in block 204A, the differential pressure target discharge amount QΔp is set as the selected discharge amount target value Qor. Select. At the same time, the corrected rotation speed NrI
t becomes 0. At that time, the flag FNns becomes 0,
The discharge amount target value selection switch 301 selects the discharge amount target value Qor as the discharge amount target value Qo. As a result, the hydraulic pump ↓ is controlled to the differential pressure target discharge amount QΔp.

Due to an increase in the amount of depression of the pilot pressure reducing valve 58, etc., the required flow rate of the flow control valves 4 and 5 becomes the input limit target discharge amount QT.
When it becomes larger, the differential pressure target discharge amount QΔp becomes larger than QT, and in block 204A, QT is selected as the discharge amount target value Qor.

At the same time, the target discharge amount deviation ΔQ becomes ten, and the correction rotation speed increases. At this time, if the corrected rotational speed N++s exceeds the predetermined 72 value Nnsa, the flag FNns becomes OFF, and the discharge amount target value selection switch 301 selects the input limit target discharge amount QT as the discharge amount target value Qo. As a result, the hydraulic pump 1 is controlled to the input limited target discharge amount QT. Further, the rotational speed of the engine 63 is controlled to increase by the corrected rotational speed Nns, and the saturated state is eliminated.

Up to this point, the operation is the same as that of the embodiment shown in FIG. 12.

After that, when the required flow rate of the flow rate control valves 4 and 5 decreases due to a decrease in the amount of depression of the pilot pressure reducing valve 58, etc., the differential pressure target discharge amount QΔp decreases and the input limit target discharge amount QT
less. Then, in block 204A, QΔp is selected as the discharge amount target value Qor. At that time, the target discharge amount deviation ΔQ becomes one, but the corrected rotation speed Nn
Since s gradually decreases, it remains at the value of 10, and the flag FNns is set while the corrected rotation speed Nns is equal to or higher than the predetermined value Nnsa.
is held at 1. Therefore, the discharge amount target value selection switch 301 selects the input limit target 73 discharge amount QT as the discharge amount target value Qo, and the hydraulic pump 1 maintains the state controlled to QT. This state continues until the corrected rotational speed Nns decreases to the predetermined value Nnsa, the rotational speed of the engine 63 decreases accordingly, and the flow rate passing through the flow rate control valves 4 and 5 matches QT. This prevents the hydraulic pump 1 as described above from being controlled to the differential pressure target discharge amount and interfering with the engine speed increase control.

When the corrected rotation speed Nns reaches the predetermined value Nnsa, the flag FN
When ns is switched to O, the discharge amount target value selection switch 30
1 selects the differential pressure target discharge amount QΔp as the discharge amount target value Qo. After that, the differential pressure target discharge amount QΔp is determined by the flow rate control valve 4.
, 5 is controlled to match the required flow rate.

According to this embodiment, in addition to the effects of the first embodiment, even when the operation amount of the operating means is reduced from the engine speed increase control state and the required flow rate is reduced, the engine speed increase control and the hydraulic pump can be controlled. Prevents interference with load sensing control,
More stable control can be performed.

74 Sixth Embodiment A sixth embodiment of the present invention will be described with reference to FIG. 22. In this embodiment, the calculation of the input limited target discharge amount in the embodiment shown in FIG. 21 is changed from the proportional type to the integral type.

In FIG. 22, block 500 is a target discharge pressure calculation block, which inputs the previous discharge amount target value Qo-1 and calculates the currently allowable target discharge pressure Pr from the preset input limit torque of the hydraulic pump 1. do. The target discharge pressure Pr is sent to the differential pressure calculation block 501, where the target discharge pressure P+ and the current discharge pressure P are compared to calculate the differential pressure ΔP.

The differential pressure ΔP is the input limit target discharge amount increment calculation block 502
is multiplied by the integral gain Klp to calculate an increment ΔQII8 of the input limit target discharge amount for one control cycle time.

The input limit target discharge amount increment ΔQp8 is selected from the discharge amount increment minimum value selection block 204 together with the differential pressure target discharge amount increment ΔQΔp.
Here, the magnitude of the two is compared, and the smaller one is outputted as the target discharge amount increment ΔQor.

In the discharge amount increment selection switch 301A, block 300
Flag FN++s based on the corrected rotation speed Nns output by
If FNns is 0, the target discharge amount increment ΔQor selected by the minimum value selection block 204B for discharge amount increment is selected, and if FNns is 1, the input limit target discharge amount increment ΔQps is selected, and the discharge amount increment is Output as ΔQo.

The discharge amount increment ΔQo selected by the discharge amount increment selection switch 301A is added to the discharge amount target value QO-1 calculated in the previous control cycle in block 503 to calculate the current discharge amount target value Qo.

The input limit target discharge amount increment ΔQps and the differential pressure target discharge amount increment ΔQΔp are also determined by the discharge amount increment minimum value selection block 204
A target discharge amount deviation ΔQ, which is a difference signal between the two, is calculated.

Other blocks are the same as those in FIG. 21.

In FIG. 22, blocks 201, 202, 204B, 30
The flow of 1A, 503 is the load 76 in Figure 21. Blocks 201, 202, 20 in sensing control.
This is the same flow as the calculation flow of the differential pressure target discharge amount in No. 3, 204A, and 301. On the other hand, blocks 500, 501, 50
The flow of steps 2, 204B, 301A, and 503 replaces the flow of calculating the input limit target discharge amount of blocks 200, 204A, and 301 in FIG.

In FIG. 21, a proportional type control is performed in which the input limit target discharge amount QT is calculated directly from the discharge pressure P of the hydraulic pump 1, whereas in FIG. 22 of this embodiment, the input limit torque of the hydraulic pump 1 is The input limit target value is calculated by integral type control in which a discharge amount increment ΔQps is calculated to control the target discharge pressure Pr calculated from , and the value is added to the previous discharge amount target value. However, in the block diagram of FIG. 22, the minimum value selection 2 0 4 B and the selection switch 301A operate on the discharge amount increment.

This is due to the following reasons.

In this embodiment, if the target discharge amount is calculated as shown in FIG. 21, 77 QT = Qo-1 + ΔQ ps − (5
)QΔp=QD-1+ΔQp...(6)
Here, Qo=Select (Min(QT, QΔp)
, QT), so (5). , (Substituting formula 61, Qo = Qo-1 10 Select (Min(ΔQp8sΔQΔp), Δ
Qpa), and FIGS. 21 and 22 perform the same function.

That is, in the load sensing control shown in Fig. 22, the increment of the differential pressure target discharge amount calculated from the differential pressure control and the increment of the input limit target discharge amount calculated from the limit torque are constantly compared, and the minimum of them is By adding the value to the current pump discharge amount, it is always determined which direction the pump discharge amount should be controlled.

Moreover, if the target discharge amount is used like block 205 in FIG. 21 instead of the target discharge amount deviation calculation block 205A, then ΔQ=QΔp −QT, where (5). By substituting equation (6), 78 ΔQ=(Qo−1 +ΔQΔp) (Qo−1+ΔQ ps) =ΔQΔp−ΔQps, and block 205A in FIG. 22 becomes equivalent to block 205 in FIG. 21. Block 206 and below are the 21st
The operation is exactly the same as the one shown in the figure.

According to this embodiment, the basic function is the same as that of the embodiment shown in FIG. Nns controls the engine speed to increase and eliminates the saturated state of the hydraulic pump. Also, this engine 6
Hydraulic pump 1 is in a state where the rotation speed is controlled to increase.
The same applies to controlling the differential pressure to the target discharge amount to prevent interference with engine speed increase control.

However, in this embodiment, since the integral type is used to calculate the input-limited target discharge rate, when the hydraulic pump is controlled from the differential pressure target discharge rate to the input-limited target discharge rate, or When the opposite is true, the new target discharge amount Q
Since o is always calculated from the previous target 79 discharge amount QO-1, the transition can be made smoothly. Therefore, the hydraulic pump does not move suddenly when control is transferred, and more stable control can be performed.

Seventh Embodiment A seventh embodiment of the present invention will be described with reference to FIG. 23. In the figure, the same members as those shown in FIG. 12 are given the same reference numerals. This embodiment differs from the previous embodiments in the configuration of the calculation part for the corrected rotational speed Nns.

That is, block 601 is a half-wave rectifier, and adder 20
Input the differential pressure deviation ΔP' = ΔPo - ΔP calculated in step 1, and when ΔP' ≧0, output ΔP''20,
When P'<0, ΔP' = ΔP'' is output. The output ΔP' of the half-wave rectifier 601 and the pressure deviation ΔP' are input to the signal selection switch 602. The signal selection switch 6
02 receives the output ΔQ of the adder 2o5, and when ΔQ is positive, that is, the differential pressure target discharge amount QΔp≧input limit target discharge amount QT
When ΔP′ is selected, and when ΔQ is negative, that is, QΔ8
When 0 p<QT, the value ΔP' is selected and output as the intermediate value increment ΔN'ns. This value ΔN I ns is added to the output Nns-1 from the previous control cycle in an adder 207 to obtain an intermediate value N' ns. This value N I nS
is sent to limiter 208. The limiter 208 has a value of N'
This is to prevent ns from exceeding the maximum value,
It is output as the corrected rotation speed Nns.

With this configuration, when the differential pressure target discharge amount QΔp is larger than the input limit target discharge amount QT and engine speed increase control is required, the signal selection switch 602 selects the intermediate value N'.
ΔP'(>O) is selected as ns, and positive ΔP'
The rotation speed of the engine 63 is controlled to increase using the corrected rotation speed Nn8 obtained from the above. On the other hand, when QΔp<QT and there is no need to control the increase in engine speed, even if the differential pressure ΔP decreases due to a response delay in the load sensing control of the hydraulic pump, the positive part is cut by the half-wave rectifier 601. The signal selection switch 602 selects the intermediate value increment ΔN.
'IIS selected as N' ns = N ns =
810, and the rotational speed of the engine 63 is not controlled to increase.

On the other hand, when the engine 63 rotational speed is controlled to increase, for example, when the pilot pressure reducing valve 58 is released and the hydraulic pump 1 is controlled to the differential pressure target discharge amount QΔp, the differential pressure ΔP increases. , the differential pressure deviation ΔP' becomes negative, and its value is not cut by the half-wave rectifier 601 and becomes negative ΔP'.
The engine 63 is controlled in the direction of canceling the rotation speed increase control using the reduced corrected rotation speed Nns obtained from P'.

In this way, the present embodiment also provides the same functions as the first embodiment.

Note that in this embodiment, the adder 207 and the limiter 2
Although the calculation is performed using the integral control method in 08, the calculation may be performed using the proportional control method.

Other Embodiments In the above embodiments, the target rotation speed is calculated as the corrected rotation speed Nns, and when Nns≧Nθ with respect to the target rotation speed Nθ set by the rotation speed setting device 68, the correction rotation speed Nns is used instead of Nθ. Although N++s is the final target rotation speed Ntθ, instead of calculating the corrected rotation speed Nns82 as the target rotation speed, calculate the correction rotation speed as an additional value and add this to the target rotation speed Nθ. The final target rotational speed Nrθ may be created by the following, and the rotational speed increase control may be performed.

Furthermore, in the above embodiments, a swing motor and a boom cylinder have been described as the hydraulic actuators 2 and 3, but it goes without saying that the actuators 2 and 3 may be other actuators.

In summary, the embodiments described above can be modified in various ways within the spirit of the present invention.

〔Effect of the invention〕

According to the first aspect of the present invention, the rotational speed of the prime mover automatically increases when the discharge amount of the hydraulic pump is saturated, so that the target rotational speed Nθ can be set to a low rotational speed, and fuel consumption and noise can be reduced. At the same time, the durability of the equipment can be improved. Further, even if the discharge amount of the hydraulic pump is saturated, a desired actuator speed can be obtained and a reliable operation can be performed in a compound operation.

According to the second invention, when the corrected rotational speed Nns is calculated 83 and the operation amount of the operating means exceeds a predetermined value, control is performed to increase the rotational speed of the prime mover. In addition to obtaining the effect, even if the corrected rotation speed Nns is calculated, the motor is controlled based on the target rotation speed Nθ as long as the operation amount of the operating means is below the predetermined value.
Even when a large load is applied to the actuator and the pump discharge amount is saturated, the rotational speed of the prime mover does not increase, and fluctuations in the rotational speed of the prime mover during fine operation can be prevented, improving operability. In addition, since there is no fluctuation in the rotational speed of the prime mover during slight operation, it is not unpleasant to the user's ears, and furthermore, it is possible to prevent deterioration of fuel efficiency due to frequent fluctuations in the rotational speed of the prime mover. Furthermore, when the manipulated variable exceeds a predetermined value, the rotational speed of the prime mover increases, so that the driving state of the prime mover matches the operator's intention, providing a good operating feeling.

According to the third aspect of the present invention, when the corrected rotational speed Nns is calculated and the operation amount of the operating means exceeds a predetermined value, control is performed to increase the rotational speed of the prime mover, and the corrected rotational speed Nns is calculated. Also the operation of the operating means 84? is in the range below a predetermined value, and when the target rotational speed No. is smaller than the target rotational speed NI, the prime mover is controlled based on the target rotational speed No. Therefore, the same effect as the second invention can be obtained. . (2) Furthermore, when the corrected rotational speed Nns is not calculated, or even if the corrected rotational speed Nns is calculated, the operation amount of the operating means is within a predetermined value or less,
When the target rotation speed Nθ is larger than the target rotation speed NI,
Since the rotation speed of the prime mover is controlled based on the target rotation speed Nl, the rotation speed of the prime mover is controlled in conjunction with the operation amount of the operating means that controls the flow rate control valve, and the target rotation speed Nθ is controlled by the rotation speed setting means. You can set it low to improve fuel efficiency, and
A powerful operation feeling proportional to the amount of operation of the operation means can be obtained, improving operability.

[Brief explanation of drawings]

FIG. 1 is a configuration diagram showing a load sensing control hydraulic circuit and its control device according to an embodiment of the present invention, and FIG. 2 is a side view of a wheeled hydraulic excavator to which the embodiment is applied. FIG. 3 is a schematic diagram showing the structure of the discharge amount control chamber of the control device 85.
The figure is a schematic diagram showing the configuration of a control unit forming the main body of the control device, FIG. 5 is a flowchart showing a control procedure program executed by the control unit, and FIG. FIG. 7 is a block diagram showing the procedure for determining the differential pressure target discharge amount from the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure, and FIG. 8 is a diagram showing the input torque limiting function used. FIG. 9 is a block diagram showing the procedure for calculating the corrected rotation speed from the target discharge amount deviation;
The figure is a block diagram showing the procedure for calculating the total consumable flow correction value from the target discharge amount deviation, and FIG. 10 is the procedure for controlling the discharge amount control device from the discharge amount target value, the tilt angle signal, and the actual rotation speed signal FIG. 11 is a flowchart showing the procedure for controlling the pulse motor from the target rotation speed and actual rotation speed signals, and FIG. 12 is a control block diagram showing the entire control procedure as a block diagram. , FIG. 13 is a flow chart similar to FIG. 5 showing a modification of the above embodiment, FIG. 14 is a flow chart similar to FIG. 5 showing another modification of the above 86 embodiment, and FIG. 16 is a block diagram showing a load sensing control hydraulic circuit and its control device according to a second embodiment of the present invention, FIG. 16 is a flowchart showing the control procedure of only the engine control system of the control device, and FIG. FIG. 18 is a control block diagram showing the control procedure in block diagram form, FIG. 18 is a flowchart showing the control procedure of only the engine control system of the control device according to the third embodiment of the present invention, and FIG. 20 is a control block diagram showing a control procedure in a block diagram, FIG. 20 is a partial control block diagram of a control device according to a fourth embodiment of the present invention, and FIG. 21 is a control block diagram according to a fifth embodiment of the present invention. FIG. 22 is a partial control block diagram of the control device according to the sixth embodiment of the present invention, and FIG. 23 is a partial control block diagram of the control device according to the sixth embodiment of the present invention.
FIG. 2 is a partial control block diagram of a control device according to an embodiment of the present invention. Explanation of symbols 1... Hydraulic pumps 2, 3... Hydraulic actuator 87 4, 5... Flow rate control valves 6, 7... Pressure compensation valve l4... Pressure detector (second detection means) 16... Discharge amount control device 43... Differential pressure gauge (first detection means) 40... Control unit 62... Pressure detector (third detection means) 68... Rotation speed setting device 120 , 201-203...first means 200...
Second means 204...Third means 130-132, 205-208...Fourth means 21
3,214...Fifth means 213,214,220,221-...Sixth means 22
3... Seventh means 213, 214, 220, 221, 224... Eighth means

Claims (14)

[Claims] (1) A prime mover, a variable displacement hydraulic pump driven by the prime mover, at least one hydraulic actuator driven by pressure oil discharged from the hydraulic pump, and a connection between the hydraulic pump and each hydraulic actuator. and a flow rate control valve that controls the flow rate of pressure oil supplied to the hydraulic actuator in accordance with an operation signal from an operation means, in a control device for a load sensing control hydraulic circuit, in which a target rotational speed Nθ of the prime mover is set. Rotation speed setting means;
a first detection means for detecting the differential pressure between the discharge pressure of the hydraulic pump and the load pressure of the actuator; a second detection means for detecting the discharge pressure of the hydraulic pump; and a difference between the first detection means. a first means for calculating a differential pressure target discharge amount QΔp of a hydraulic pump that maintains the differential pressure constant from a pressure signal; and at least a pressure signal of the second detection means and a preset hydraulic pump input limiting function. a second means for calculating an input-limiting target discharge amount QT of the hydraulic pump from the above; , a third means for controlling the discharge amount of the hydraulic pump so as not to exceed the input limit target discharge amount QT; and when the input limit target discharge amount QT is selected by the third means, at least the difference is controlled. a fourth means for calculating a corrected rotation speed Nns of the prime mover based on a pressure target discharge amount QΔp and an input limited target discharge amount QT; and when the corrected rotation speed Nns is not calculated by the fourth means, the rotation speed is The rotation speed of the prime mover is controlled based on the target rotation speed Nθ set by the number setting means, and when the corrected rotation speed Nns is calculated, the rotation speed of the prime mover is controlled based on the target rotation speed Nθ and the correction rotation speed Nns. A control device for a load sensing control hydraulic circuit, comprising: fifth means for controlling the number of revolutions of the hydraulic circuit to increase. (2) In the load sensing control hydraulic circuit control device according to claim 1, the fifth means selects the larger value of the target rotation speed Nθ and the corrected rotation speed Nns, and based on this value, the fifth means selects the larger value of the target rotation speed Nθ and the corrected rotation speed Nns. A control device comprising: large value selection means for controlling the rotational speed of a prime mover. (3) A prime mover, a variable displacement hydraulic pump driven by the prime mover, at least one hydraulic actuator driven by pressure oil discharged from the hydraulic pump, and a connection between the hydraulic pump and each hydraulic actuator. and a flow rate control valve that controls the flow rate of pressure oil supplied to the hydraulic actuator in accordance with an operation signal from an operation means, in a control device for a load sensing control hydraulic circuit, in which a target rotational speed Nθ of the prime mover is set. Rotation speed setting means;
a first detection means for detecting a pressure difference between the discharge pressure of the hydraulic pump and a load pressure of the actuator; a second detection means for detecting the discharge pressure of the hydraulic pump; and a detection amount of operation of the operation means. a third detection means for calculating a differential pressure target discharge amount QΔp of a hydraulic pump that maintains the differential pressure constant from a differential pressure signal of the first detection means; a second means for calculating an input limited target discharge amount QT of the hydraulic pump from the pressure signal of the detection means and a preset input limiting function of the hydraulic pump; a third means for selecting either one as the discharge amount target value Qo of the hydraulic pump and controlling the discharge amount of the hydraulic pump so as not to exceed the input limit target discharge amount QT; a fourth means for calculating a corrected rotational speed Nns of the prime mover based on at least the differential pressure target discharge amount QΔp and the input limited target discharge amount QT when the input limited target discharge amount QT is selected; When the correction rotation speed Nns is not calculated by the means, or when the operation amount of the operation means detected by the third detection means is less than or equal to a predetermined value, the target rotation speed set by the rotation speed setting means The rotational speed of the prime mover is controlled based on Nθ, and when the corrected rotational speed Nns is calculated by the fourth means and the operation amount of the operating means exceeds a predetermined value, the target rotational speed Nθ and the fourth and sixth means for controlling the rotational speed of the prime mover to increase based on the corrected rotational speed Nns calculated by the calculation means. (4) In the control device for a load sensing control hydraulic circuit according to claim 3, the sixth means is configured to detect when the operation amount of the operation means detected by the third detection means exceeds a predetermined value. A control device comprising a large value selection means for selecting the larger value of the target rotational speed Nθ and the corrected rotational speed Nns, and controlling the rotational speed of the prime mover based on this value. (5) A prime mover, a variable displacement hydraulic pump driven by the prime mover, at least one hydraulic actuator driven by pressure oil discharged from the hydraulic pump, and a connection between the hydraulic pump and each hydraulic actuator. and a flow rate control valve that controls the flow rate of pressure oil supplied to the hydraulic actuator in accordance with an operation signal from an operation means, in a control device for a load sensing control hydraulic circuit, in which a target rotational speed Nθ of the prime mover is set. Rotation speed setting means;
a first detection means for detecting a pressure difference between the discharge pressure of the hydraulic pump and a load pressure of the actuator; a second detection means for detecting the discharge pressure of the hydraulic pump; and a detection amount of operation of the operation means. a third detection means for calculating a differential pressure target discharge amount QΔp of a hydraulic pump that maintains the differential pressure constant from a differential pressure signal of the first detection means; a second means for calculating an input limited target discharge amount QT of the hydraulic pump from the pressure signal of the detection means and a preset input limiting function of the hydraulic pump; a third means for selecting either one as the discharge amount target value Qo of the hydraulic pump and controlling the discharge amount of the hydraulic pump so as not to exceed the input limit target discharge amount QT; a fourth means for calculating a corrected rotational speed Nns of the prime mover based on at least the differential pressure target discharge amount QΔp and the input limited target discharge amount QT when the input limited target discharge amount QT is selected; seventh means for calculating the target rotational speed Nl of the prime mover according to the operation amount of the operating means detected by the detection means; and when the corrected rotational speed Nns is not calculated by the fourth means; When the operation amount of the operation means detected by the third detection means is less than or equal to a predetermined value, the target rotation speed Nθ set by the target rotation speed setting means and the seventh
Selecting the larger value of the target rotational speed Nl calculated by the means, and controlling the rotational speed of the prime mover based on this value,
When the correction rotation speed Nns is calculated by the fourth means and the operation amount of the operation means exceeds a predetermined value, the correction rotation speed Nns is calculated by the seventh means with the target rotation speed Nθ set by the target rotation speed setting means. and eighth means for controlling the rotational speed of the prime mover to increase based on the target rotational speed Nl and the corrected rotational speed Nns calculated by the fourth calculation means. Sensing control hydraulic circuit control device. (6) In the control device for a load sensing control hydraulic circuit according to claim 5, the eighth means is configured to detect when the operation amount of the operation means detected by the third detection means exceeds a predetermined value. Selecting the maximum value of the target rotation speed Nθ set by the target rotation speed setting means, the corrected rotation speed Nns calculated by the fourth calculation means, and the target rotation speed Nl calculated by the seventh calculation means. , a control device for a load sensing control hydraulic circuit, comprising maximum value selection means for controlling the rotation speed of the prime mover based on this value. (7) In the control device for a load sensing control hydraulic circuit according to claim 1, 3 or 5, the third means sets the smaller of the differential pressure target discharge amount QΔp and the input limit target discharge amount QT to the hydraulic pump. A control device characterized in that the discharge amount target value Qo is selected as the discharge amount target value Qo. (8) In the control device for a load sensing control hydraulic circuit according to claim 1, 3 or 5, the third means is configured to control the differential pressure target discharge amount QΔp when the corrected rotation speed Nns is less than a predetermined value Nnsa. is selected as the discharge amount target value Qo of the hydraulic pump, and the corrected rotation speed Nns is set to a predetermined value Nnsa.
In the above case, the control device is characterized in that the input limit target discharge amount QT is selected as the discharge amount target value Qo of the hydraulic pump. (9) In the control device for a load sensing control hydraulic circuit according to claim 1, 3 or 5, the fourth means includes a target discharge amount deviation as a deviation between the differential pressure target discharge amount QΔp and the input restriction target discharge amount QT. A control device comprising an addition means for determining ΔQ, and calculating the corrected rotation speed Nns using at least this target discharge amount deviation ΔQ. (10) In the control device for a load sensing control hydraulic circuit according to claim 9, the fourth means further includes an increment value ΔNns of the corrected rotation speed Nns for reducing the deviation from the target discharge amount deviation ΔQ to zero. is calculated, and this value is added to the previously calculated corrected rotational speed Nns-1 to obtain the corrected rotational speed N.
A control device characterized by having an integral type calculation means for determining ns. (11) The control device for a load sensing control hydraulic circuit according to claim 9, wherein the first means detects a differential pressure deviation ΔP' between a differential pressure signal of the first detecting means and a preset target differential pressure. the fourth means further comprises a filter that sets the output ΔP'' to zero when the differential pressure deviation ΔP' is positive, and sets the value ΔP' as the output ΔP' when the differential pressure deviation ΔP' is negative; and the target discharge amount deviation Δ
selection means for selecting the output ΔP' of the filter means when Q is negative and selecting the output ΔP' of the addition means when the target discharge amount deviation ΔQ is positive; and a value selected by the selection means. A control device comprising a calculation means for calculating the corrected rotation speed Nns from ΔP' or ΔP'. (12) In the control device for a load sensing control hydraulic circuit according to claim 1, 3 or 5, the first means is configured to maintain the differential pressure constant based on the differential pressure signal of the first detection means. An increment value ΔQΔp of the differential pressure target discharge amount QΔp is calculated, and this value is calculated as the previously calculated differential pressure target discharge amount Qo−1.
is an integral type calculation means for calculating a differential pressure target discharge amount QΔp by adding the pressure signal of the second detection means to the target discharge pressure obtained from the input limiting function of the hydraulic pump. P
Incremental value Δ of input limit target discharge amount QT for controlling to r
Qps and adds this value to the previously calculated input limit target discharge amount QT-1 to obtain the input limit target discharge amount QT; Means for selecting either one of the incremental value ΔQΔp of the target discharge amount QΔp and the incremental value ΔQps of the input limited target discharge amount QT to select either the differential pressure target discharge amount QΔp or the input limited target discharge amount QT. A control device characterized by: (13) In the control device for a load sensing hydraulic drive circuit according to claim 1, 3 or 5, the input limiting function of the second means sets either the discharge pressure or the input limiting target discharge amount of the hydraulic pump as a parameter. an input torque limiting function, and the second means calculates an input limiting target discharge amount QT of the hydraulic pump from the pressure signal of the second detecting means and the input torque limiting function. control device. (14) The control device for a load sensing control hydraulic circuit according to claim 1, 3 or 5, further comprising a fourth detection means for determining a deviation between a target rotational speed and an actual rotational speed of the prime mover, The input limiting function of the means is an input torque limiting function using either one of the discharge pressure and the input limited target discharge amount of the hydraulic pump and the rotational speed deviation of the prime mover as parameters, and this second means A control device for calculating an input limit target value QT of the hydraulic pump from the pressure signal of the second detecting means, the rotation speed deviation signal of the fourth detecting means, and the input torque limit function.