JP2005009402A - Hydraulically driven device for work unit - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a hydraulically driven device for a work unit to enable execution of control equal to control in increasing horse power and decreasing horse power of a hydraulic pump effected by conventional speed sensing control irrespective of the output characteristics of an engine, and enable improvement of control precision by further accurately grasping an engine load than conventional speed sensing control. <P>SOLUTION: A working unit controller 18 has the function of each of an engine load rate computing part 18a; a target engine load rate setting part 18b; and a pump torque control part 18 by load rate control. The engine load rate computing part 18a calculates an engine load rate, being a rate of a load applied on an engine 1, from a control lack position. The target engine load rate setting part 18b sets a target engine load rate based on a mode signal. The pump torque control part 18c effects increase or decrease horse power control of the hydraulic pump 2 such that an engine load rate is held at a target engine load rate. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a hydraulic drive device for a working machine that includes a diesel engine as a prime mover, drives a variable displacement hydraulic pump by the engine, and drives an actuator.
[0002]
[Prior art]
A working machine such as a hydraulic excavator generally includes a diesel engine (hereinafter, referred to as an engine as appropriate) as a prime mover, and the engine operates by driving a variable displacement hydraulic pump and driving an actuator. The engine control is generally performed by using a fuel injection device such as a mechanical governor or an electronic governor and controlling the fuel injection amount with this fuel injection device. The control of the hydraulic pump is generally performed by combining the capacity control based on the required flow rate and the torque control (horsepower control) based on the pump discharge pressure. The torque control of the hydraulic pump is to prevent the engine overload by controlling the absorption torque of the hydraulic pump not to exceed the set value (maximum absorption torque) by reducing the capacity of the hydraulic pump as the pump discharge pressure increases. To do.
[0003]
As a technique for effectively using the output horsepower of an engine by torque control of a hydraulic pump, for example, speed sensing control described in JP-A-57-65822 and JP-A-11-101183 is known. This speed sensing control converts the deviation between the target engine speed and the actual engine speed into a torque correction value, and adds or subtracts this torque correction value to or from the pump base torque to correct the pump base torque and absorb this. The torque of the hydraulic pump is controlled as a torque setting value (maximum absorption torque), and the horsepower control or horsepower reduction control of the hydraulic pump is performed.
[0004]
Further, an engine equipped with a conventional mechanical governor or electronic governor has a characteristic (droop characteristic) in which the engine speed increases as the engine output torque (engine load) decreases in the governor area (partial load area). Yes. On the other hand, in recent years, as described in, for example, JP-A-10-89111 and JP-A-10-159599, an engine equipped with a fuel injection control device capable of controlling the governor region to isochronous characteristics (hereinafter referred to as appropriate) , An engine that performs isochronous control) has been proposed. The isochronous characteristic of engine control is a characteristic that keeps the engine speed constant in the governor region regardless of the weight of the engine load, that is, regardless of the decrease in engine output torque. This eliminates fluctuations in engine speed due to load fluctuations in the governor region as compared with an engine having droop characteristics, thereby realizing low fuel consumption and low noise.
[0005]
[Patent Document 1]
JP-A-57-65822
[Patent Document 2]
JP-A-11-101183
[Patent Document 3]
JP-A-10-89111
[Patent Document 4]
Japanese Patent Laid-Open No. 10-159599
[0006]
[Problems to be solved by the invention]
However, the above prior art has the following problems.
[0007]
The output torque characteristic of a diesel engine is divided into a governor area (partial load area) characteristic and a full load area characteristic. As described above, an engine equipped with a conventional mechanical governor has a characteristic that the engine speed increases as the engine output torque (engine load) decreases, that is, a droop characteristic in the governor region. This characteristic is caused by the inertia of the flywheel included in the mechanical governor, and even in an engine that uses an electronic governor as a fuel injection device, the same characteristic (droop characteristic) is used in order to enable the same usage pattern as the mechanical governor. Often set.
[0008]
The conventional speed sensing control described in Japanese Patent Application Laid-Open No. 57-65822 and Japanese Patent Application Laid-Open No. 11-101183 is applied to an engine having such droop characteristics. Not only can the horsepower reduction be controlled on the entire load range where the engine speed (rated speed) falls, but the engine speed will be higher than the target speed on the governor side. As a result, the following effects can be obtained.
[0009]
1. When there is a surplus in engine output, that is, when the engine speed becomes higher than the rated speed, the engine output can be used more effectively by increasing the pump absorption horsepower (horsepower control).
[0010]
2. When the engine load suddenly increases due to a sharp increase in excavation resistance or the like, the pump absorption horsepower can be reduced to prevent an excessive load torque from being applied to the engine and to prevent engine stall (horse reduction control). As a result, the maximum absorption torque (set value) of the hydraulic pump can be set close to the maximum output torque of the engine, and the engine output can be effectively used.
[0011]
3. Even if there are variations in engine performance (for example, manufacturing variations with respect to design performance or performance degradation after long hours of operation), the pump absorption horsepower is controlled accordingly to optimize engine performance, and engine output is effective. Can be used (variation correction control by increasing / decreasing horsepower control).
[0012]
By the way, the speed sensing control is a control for adjusting the pump absorption torque, which is the engine load, to the engine output, and paying attention to the feature that the increase / decrease in the engine speed almost coincides with the increase / decrease in the engine load. The deviation is detected by the engine speed deviation. However, the increase / decrease in engine speed (actual speed) involves factors other than engine load, such as mechanical loss, inertia, and increase / decrease in pump absorption torque other than the main pump. Since it is not necessarily a value that accurately reflects the deviation between the engine load and the engine output, there is room for further improvement in terms of control accuracy.
[0013]
In addition, in an engine that performs isochronous control as described in JP-A-10-89111, JP-A-10-159599, etc., fluctuations in engine speed due to load fluctuations in the governor region are eliminated, and low fuel consumption and low noise are realized. There are advantages you can do. However, since the engine speed does not change in the governor region, the engine speed deviation cannot be obtained even if the engine load increases or decreases. For this reason, even if speed sensing control is applied to a working machine having such an engine, although the horsepower reduction control can be performed on the full load region side where the engine rotational speed falls below the target rotational speed (rated rotational speed), the governor region side In this case, the horsepower control by the rotational speed deviation cannot be performed, and the effect of the horsepower control of 1 and the effect of the variation correction control of 3 cannot be obtained.
[0014]
The first object of the present invention is to enable control equivalent to the horsepower increase control and horsepower reduction control of a hydraulic pump performed by the conventional speed sensing control regardless of the engine output characteristics, and to effectively use the engine output, It is an object of the present invention to provide a hydraulic drive device for a working machine that can improve the working efficiency and can accurately grasp the engine load and improve the control accuracy as compared with the conventional speed sensing control.
[0015]
The second object of the present invention is to facilitate control with reduced fuel consumption by making it possible to arbitrarily change the maximum absorption torque of the hydraulic pump when performing control equivalent to conventional speed sensing control. It is providing the hydraulic drive device of the working machine which can be performed.
[0016]
[Means for Solving the Problems]
(1) In order to achieve the above object, the present invention provides an engine, a fuel injection device that controls the rotational speed and output of the engine, a fuel injection device controller that controls the fuel injection device, and the engine. In a hydraulic drive device for a working machine including at least one variable displacement hydraulic pump that is driven to drive an actuator, first means for calculating a load factor of the engine, and the load factor is maintained at a target value The second means for controlling the maximum absorption torque of the hydraulic pump and the third means for changing the target value of the load factor are provided.
[0017]
By calculating the engine load factor (hereinafter referred to as the engine load factor) in this way and controlling the maximum absorption torque of the hydraulic pump so that the engine load factor is maintained at the target value, the engine load factor is less than the target value. When it is small, the torque increase control of the hydraulic pump is performed, and when the engine load factor becomes larger than the target value, the torque reduction control of the hydraulic pump is performed.As a result, the conventional torque control is performed regardless of the output characteristics of the engine. Control equivalent to the horsepower control and horsepower control of the hydraulic pump that is performed by speed sensing control is possible, and the engine output can be effectively used to improve work efficiency.
[0018]
In addition, since the engine load factor directly related to the engine load is detected and controlled so that the engine load factor is maintained at the target value, the engine load can be grasped more accurately than the conventional speed sensing control. Thus, the control accuracy can be improved.
[0019]
Furthermore, by making it possible to change the target value of the engine load factor, the maximum absorption torque (maximum engine load) of the hydraulic pump can be changed arbitrarily, and the target value of the engine load factor is set to a smaller value. By reducing the maximum absorption torque, it is possible to easily perform control while suppressing fuel consumption.
[0020]
(2) In the above (1), preferably, when the deviation between the load factor and the target value is not zero, the second means controls the maximum absorption torque of the hydraulic pump so that the deviation approaches zero. When the deviation between the load factor and the target value becomes zero, the constant value control is performed so that the maximum absorption torque of the hydraulic pump at that time is maintained and the deviation is kept at zero.
[0021]
This makes it possible to set the maximum absorption torque of the hydraulic pump according to the target value of the engine load factor, and setting the target value of the engine load factor to a smaller value also reduces the maximum absorption torque of the hydraulic pump. Can be easily controlled.
[0022]
(3) In the above (1), preferably, the second means calculates a deviation between the load factor and a target value, and converts the deviation into an increment of a target constant torque value. Means for calculating a new target constant torque value by adding the increment to the previous target constant torque value, and controlling the maximum absorption torque of the hydraulic pump using the new target constant torque value.
[0023]
As a result, the constant value control described in the above (2) becomes possible.
[0024]
(4) Further, in the above (1), preferably, the first means detects a control rack position of the fuel injection device,
Engine load factor (%) = [(current control rack position-control rack position at no load) / (control rack position at full load-control rack position at no load)] x 100
In the above formula, the control rack position at full load and the control rack position at no load are values obtained and stored in advance.
The load factor is calculated by the following formula.
[0025]
As a result, the engine load factor can be obtained using a relatively easy-to-detect state quantity such as the control rack position, and a practical load factor detecting means can be provided.
[0026]
(5) In the above (1), preferably, the third means is operated by an operator, and selectively sets a normal load factor target value and a load factor target value for energy saving smaller than that. Means.
[0027]
As a result, the maximum absorption torque of the hydraulic pump is also controlled to be small, so that control with reduced fuel consumption can be performed.
[0028]
(6) In order to achieve the above object, the present invention provides an engine, a fuel injection device that controls the rotational speed and output of the engine, a fuel injection device controller that controls the fuel injection device, In a hydraulic drive device for a work machine including at least one variable displacement hydraulic pump driven by an engine and driving an actuator, first means for calculating a load factor of the engine, and the load factor is maintained at a target value. And a second means for controlling the maximum absorption torque of the hydraulic pump so that, when the deviation between the load factor and the target value is not zero, the hydraulic pressure is adjusted so that the deviation approaches zero. The maximum absorption torque of the pump is controlled, and when the deviation between the load factor and the target value becomes 0, the constant value control is performed so that the maximum absorption torque of the hydraulic pump at that time is maintained and the deviation is kept at 0. It assumed to be a means for performing.
[0029]
As a result, as described in (1) and (2) above, it is possible to perform the same control as the horsepower control and horsepower control of the hydraulic pump, which is performed by the conventional speed sensing control, regardless of the output characteristics of the engine. The engine output can be effectively utilized to improve the working efficiency, and the control accuracy can be improved by accurately grasping the engine load as compared with the conventional speed sensing control.
[0030]
Further, when performing control equivalent to conventional speed sensing control, the maximum absorption torque of the hydraulic pump can be arbitrarily changed, and control with reduced fuel consumption can be easily performed.
[0031]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
[0032]
FIG. 1 is a diagram showing an entire system of a hydraulic drive device for a working machine according to an embodiment of the present invention.
[0033]
The hydraulic drive apparatus according to the present embodiment is provided in a working machine, for example, a hydraulic excavator. As shown in FIG. 1, an engine 1 and an electronic governor 12 constituting a fuel injection control device of the engine 1 are provided. The engine controller 13 and the throttle dial 14 are provided. The electronic governor 12 and the engine controller 13 are capable of controlling the governor region to isochronous characteristics, that is, to perform isochronous control for maintaining the engine 1 at the rated speed regardless of the increase or decrease of the engine load in the governor region. The electronic governor 12 is controlled by the engine controller 13 and controls the output and the rotational speed of the engine 1 by injecting fuel into the engine 1. The engine controller 13 inputs a target rotation speed instruction signal from the throttle dial 14 and controls the electronic governor 12 in accordance with the instruction signal.
[0034]
The hydraulic drive device according to the present embodiment controls, for example, a swash plate type variable displacement hydraulic pump 2 driven by the engine 1 and a displacement volume (tilt angle of the swash plate) of the hydraulic pump 2. Regulator 16, a plurality of hydraulic actuators such as a hydraulic cylinder 3 driven by pressure oil discharged from the hydraulic pump 2, a hydraulic motor 4, and hydraulic cylinders 5 and 6, and a flow of pressure oil supplied to these hydraulic actuators Directional control valves 7 to 10 for controlling the operation, the main relief valve 11, an operating lever device 50 for generating pilot pressure for switching the directional control valves 7 to 10 (only one is shown), a hydraulic pump 2 detects a discharge pressure of 2 and outputs a pump pressure signal P, and detects a tilt angle (a displacement volume) of a swash plate of the hydraulic pump 2 and outputs a tilt angle signal θ. Tilt angle detector 25, rack position detector 26 for detecting the control rack position of the electronic governor 12 and outputting the control rack position signal L, and selecting the normal mode or the energy saving mode and outputting the corresponding mode signal F A signal control valve 53 having a combination of a possible mode selection switch 27, a shuttle valve that inputs a pilot pressure from the operation lever device 50,..., Selects and outputs one of the pilot pressures, and outputs from the signal control valve 53 Pressure detector 55 that detects the pilot pressure and outputs pilot pressure signal D, pump pressure signal P output from pressure detector 24, tilt angle signal θ output from tilt angle detector 25, rack A control rack position signal L output from the position detector 26, a mode signal F output from the mode selection switch 27, Enter the pilot pressure signal D outputted from the force detector 55, and a working machine controller 18 for outputting a control current signal R for controlling the volume except push regulator 16.
[0035]
FIG. 2 shows an external appearance of a hydraulic excavator in which the hydraulic drive device according to the present embodiment is mounted.
[0036]
The hydraulic excavator includes a lower traveling body 102, an upper swing body 103, and a front work machine 104. The upper swing body 103 is turnably mounted on the upper part of the lower travel body 102. It is attached to the front part so that it can move up and down. The upper swing body 103 is provided with an engine room 105 and an operator cab 106. The front work machine 104 has an articulated structure having a boom 108, an arm 109, and a bucket 110. The lower traveling body 102, the upper swing body 103, and the front work machine 104 have left and right traveling motors 111 (only one shown), a swing motor 112, a boom cylinder 113, an arm cylinder 114, and a bucket cylinder 115 as actuators. The traveling body 102 travels by the rotation of the left and right traveling motors 111, the upper swing body 103 rotates by the rotation of the swing motor 112, and the boom 108 of the front work machine 104 rotates vertically by the expansion and contraction of the boom cylinder 113, The arm cylinder 109 rotates up and down and front and rear as the arm cylinder 114 expands and contracts, and the bucket 110 rotates up and down and front and rear as the bucket cylinder 115 extends and contracts.
[0037]
The hydraulic cylinders 3, 5 and 6 and the hydraulic motor 4 shown in FIG. 1 represent the above-described actuators. For example, the hydraulic cylinders 3, 5 and 6 are a boom cylinder 113, an arm cylinder 114, and a bucket cylinder 115. The motor 4 is a turning motor 112.
[0038]
Further, the operation lever device 50,... And the mode selection switch 27 are disposed in the cab 106, and the engine 1 and the hydraulic pump 2 are disposed in the engine room 105. Hydraulic equipment and electronic equipment such as the direction control valves 7 to 10, the engine controller 13, and the work machine controller 18 are installed at appropriate positions of the upper swing body 103.
[0039]
FIG. 3 shows the relationship between the rotational speed N of the engine 1 and the output torque Te by the fuel injection control device (electronic governor 12 and engine controller 13) that performs isochronous control.
[0040]
As shown in FIG. 3, the output torque characteristic of the engine 1 includes a characteristic of the governor region 33 (isochronous characteristic) represented by a straight line 32 and a characteristic of the entire load region represented by a curve 30. The governor region 33 is an output region where the governor opening of the electronic governor 12 is 100% or less, and the full load region is an output region where the governor opening is 100%. The governor opening 100% of the electronic governor 12 corresponds to the maximum position of the control rack of the electronic governor 12. In the figure, a broken line 31 indicates a characteristic (droop characteristic) in a governor region of a conventional mechanical governor type engine for comparison.
[0041]
Since the mechanical governor has a structure in which the fuel injection amount is adjusted by the balance between the flywheel and the spring, the governor region of the mechanical governor type engine is the engine as the engine output torque (engine load) Te decreases as indicated by the broken line 31. It has a droop characteristic in which the rotation speed N increases. On the other hand, in the engine 1 according to the present embodiment, as indicated by the straight line 32, the isochronous control in which the engine speed N is kept constant at the rated speed N0 by the electronic governor 12 regardless of the decrease in the engine output torque Te in the governor region. It has isochronous characteristics to implement. By this isochronous control, it is possible to realize low fuel consumption and low noise compared to a working machine having a mechanical governor type engine.
[0042]
FIG. 4 shows the operating characteristics of the regulator 16. 4, the horizontal axis represents the control current signal R output from the work machine controller 18, and the vertical axis represents the tilt angle of the swash plate of the hydraulic pump 2 (hereinafter simply referred to as the tilt angle of the hydraulic pump 2 or the pump tilt as appropriate). ).
[0043]
When the control current signal R is equal to or less than R1, the tilt angle of the hydraulic pump 2 is at the minimum position, and the discharge flow rate of the hydraulic pump 2 is minimum. When the control current signal R becomes larger than R1, the tilt angle of the hydraulic pump 2 increases accordingly, the discharge flow rate of the hydraulic pump 2 increases, and when the control current signal R decreases, the pump 2 tilts accordingly. The turning angle becomes smaller and the discharge flow rate of the hydraulic pump 2 decreases. The increase amount and the decrease amount of the discharge flow rate of the hydraulic pump 2 are proportional to the increase amount and the decrease amount of the control current signal R.
[0044]
FIG. 5 is a block diagram showing the overall control flow of the hydraulic drive system and the processing functions of the work machine controller 18. The work machine controller 18 has functions of an engine load factor calculation unit 18a, a target engine load factor setting unit 18b, and a pump torque control unit 18c based on load factor control.
[0045]
The engine load factor calculation unit 18a receives a control rack position signal L (hereinafter, referred to as a control rack position L as appropriate), calculates an engine load factor that is a ratio of a load applied to the engine 1, and an engine load factor signal LF ( Hereinafter, the output is appropriately referred to as an engine load factor LF).
[0046]
Here, the engine load factor will be described. As described with reference to FIG. 3, the output torque characteristic of the engine 1 having the electronic governor 12 is composed of the characteristic 32 of the governor region 33 and the characteristic 30 of the full load region. The governor region 33 is the opening of the governor of the electronic governor 12. This is a partial load region where the degree is 100% or less, and the full load region is an output region where the governor opening is 100%. When the governor opening degree of the electronic governor 12 is 100%, the fuel injection amount (fuel flow rate) is also 100%.
[0047]
In the present invention, based on the current fuel flow rate, the fuel flow rate in a fully loaded state and the fuel flow rate in a no-load state, a value obtained by the following equation is defined as an engine load factor.
[0048]
Engine load factor (%) = [(current fuel flow rate−fuel flow rate at no load) / (fuel flow rate at full load−fuel flow rate at no load)] × 100
FIG. 6 shows the relationship between the fuel flow rate defined by the above equation and the engine load factor. In FIG. 6, the horizontal axis represents the fuel flow rate, that is, the fuel injection amount of the electronic governor 12, and the vertical axis represents the engine load factor. The fuel flow rate (fuel injection amount of the electronic governor 12) in the engine full load state (for example, point A in FIG. 3) is the maximum, and the engine load factor at this time is 100%. The fuel flow rate in the no-load state (for example, point B in FIG. 3) is the minimum, and the engine load factor at this time is 0%. In the load state in the meantime (an arbitrary point on the characteristic 32 of the governor region 33 in FIG. 3), the engine load increases linearly from 0% to 100% as the engine load increases and the fuel flow rate increases. To do.
[0049]
The engine load factor calculation unit 18a calculates the engine load factor using the control rack position of the electronic governor 12 that is inexpensive and easy to install instead of the fuel flow rate. This is based on the fact that in the engine 1 having the electronic governor 12, the control rack position in the electronic governor 12 and the fuel injection amount are in a substantially linear proportional relationship. Is detected as an engine load factor of 100%, a control rack position at no load is detected and set as a load factor of 0%, and the engine load factor is calculated from the control rack position in the meantime by the following formula.
[0050]
Engine load factor (%) = [(current control rack position−control rack position at no load) / (control rack position at full load−control rack position at no load)] × 100
In the above formula, the control rack position in the full load state and the control rack position in the no load state are values obtained and stored in advance, and the current control rack position is a value obtained from the control rack position signal L.
[0051]
FIG. 7 shows the relationship between the control rack position defined by the above equation and the engine load factor. In FIG. 7, the horizontal axis is the control rack position, and the vertical axis is the engine load factor. The control rack position is maximum when the engine is fully loaded (for example, point A in FIG. 3), and the engine load factor at this time is 100%. The control rack position in the no-load state (for example, point B in FIG. 3) is the minimum, and the engine load factor at this time is 0%. In the load state in the meantime (an arbitrary point on the characteristic 32 of the governor region 33 in FIG. 3), the engine load increases linearly from 0% to 100% as the engine load increases and the control rack position increases. Increase. Instead of obtaining the load factor from the above relational expression, the engine load factor calculating unit 18a may store the relationship of FIG. 7 in a memory and refer to the control rack position signal L to obtain the engine load factor.
[0052]
As another example of the engine load factor calculation method, there is the following method using the target fuel flow rate calculated by the engine controller 13.
[0053]
Engine load factor (%) = [(current target fuel flow rate−no load fuel flow rate) / (fuel load at full load−no load fuel flow rate)] × 100
In the above relational expression, the fuel flow rate in the full load state and the unloaded fuel flow rate are values obtained and stored in advance.
[0054]
Further, the engine load factor can be calculated in the same manner by actually detecting the fuel flow rate and replacing the target fuel flow rate with the actual fuel flow rate.
[0055]
The target engine load factor setting unit 18b sets a target engine load factor based on the mode signal F and outputs a target engine load factor signal LFr (hereinafter referred to as target engine load factor LFr as appropriate), as shown in FIG. In addition, a normal target load factor storage unit 118a, a target load factor storage unit 118b for energy saving mode, and a switching unit 118c are provided. The normal target load factor storage unit 118a stores a normal value (for example, 90%) as the target engine load factor, and the target load factor storage unit 118b for the energy saving mode stores an energy saving value as the target engine load factor. (For example, 70%) is stored, and the switching unit 118c selects a normal value when the mode signal F for selecting the normal mode is input, and saves energy when the mode signal F for selecting the energy saving mode is input. Are selected and output as target engine load factor signals LFr, respectively.
[0056]
Here, the normal value and the value for energy saving are set as follows.
[0057]
For example, when the engine load factor is measured by controlling the hydraulic pump 2 so that the pump absorption torque becomes the base torque TB in a standard specification engine in which the atmospheric pressure and the fuel quality are in a standard state, it is 90%. (However, the engine speed is fixed at the rated speed NO). The base torque TB is a design specification value determined by specifications required for the work machine. In this case, 90% is set as a normal value of the target engine load factor. Moreover, as the value for energy saving, a value smaller than 90% of the normal value, for example, 70% is set. How much smaller than the normal value is determined by a balance between a reduction in work amount due to a decrease in engine output and an energy saving effect.
[0058]
The pump torque control unit 18c controls the discharge flow rate of the hydraulic pump 2 according to the required flow rate, and controls the increase / decrease horsepower of the hydraulic pump 2 so that the engine load factor LF is maintained at the target engine load factor LFr. is there.
[0059]
FIG. 9 shows details of the pump torque control unit 18c. The pump torque control unit 18c includes a first target pump tilt angle calculation unit 218a that calculates a target pump tilt θD for positive control, and a second target pump that calculates a target pump tilt θT for pump torque limitation based on load factor control. Tilt angle calculator 218b, minimum value selector 218c for calculating target pump tilt angle θc subjected to torque limit control, subtractor 218d for calculating tilt angle deviation Δθ of feedback control, and tilt angle deviation Δθ Each function of the control current calculation unit 218e is converted to a control current signal R.
[0060]
The first target pump tilt angle calculation unit 218a receives the pilot pressure signal D from the pressure detector 55, refers to the table stored in the memory, and corresponds to the pilot pressure indicated by the signal D at that time. The first target tilt θD of the hydraulic pump 2 is calculated. This first target tilt θD is a positive control target pump tilt according to the lever operation amount (required flow rate) of the operation lever device 50 (see FIG. 1), and the pilot pressure increases in the memory table. Accordingly, the relationship between the pilot pressure and the first target pump tilt is set so that the first target pump tilt θD increases.
[0061]
The second target pump tilt angle calculation unit 218b calculates a pump torque-limited target pump tilt θT (second target pump tilt θT) for maintaining the engine load factor LF at the target engine load factor LFr. , Subtractor 218g for calculating engine load factor deviation, pump increase / decrease torque converter 218h, integral control adder 218i, constant torque curve limiter calculator 218j, primary delay element 218k, target displacement calculator 218m, pump tilt angle Each function of the conversion unit 218n is provided.
[0062]
The subtraction unit 218g calculates a deviation ΔLF between the target engine load factor LFr and the engine load factor LF. The pump increase / decrease torque conversion unit 218h calculates a target increase / decrease torque increment ΔT by multiplying the engine load factor deviation ΔLF by a gain. The adding unit 218i calculates a new target constant torque value Tp2 by adding the increment ΔT to the previously calculated target constant torque value Tp1. The constant torque curve limiter calculation unit 218j obtains a target constant torque value Tp3 by multiplying the new target constant torque value Tp2 by an upper limiter (boosting force) of Tpmax and a lower limiter (decreasing force) of Tpmin. The target displacement calculating unit 218m rewrites the target constant torque value Tp3 to Tp1, creates a new pq diagram, and refers to the pump pressure signal P from the pressure detector 24 in this pq diagram. The target displacement volume q is calculated. The pump tilt angle conversion unit 218n converts the target displacement and calculates the second target pump tilt θT of the hydraulic pump 2.
[0063]
FIG. 10 is a diagram illustrating another example of the pump increase / decrease torque conversion unit 218h. As shown in FIG. 10, the gain of the pump increase / decrease torque conversion unit 218h may be set such that the gain on the minus side of ΔLF is greater than that on the plus side, thereby increasing the effect of the torque reduction (horse reduction force).
[0064]
The minimum value selection unit 218c selects a smaller value of the first target pump tilt θD and the second target pump tilt θT to calculate the target pump tilt angle θc subjected to torque limit control, and the subtraction unit 218d By subtracting the tilt angle signal θ from the tilt angle detector 25 from the target pump tilt angle θc, the tilt angle deviation Δθ of feedback control is calculated, and the control current calculator 218e calculates the tilt angle deviation Δθ. It is converted into a control current signal R and output to the regulator 16.
[0065]
In the above, the subtraction unit 218g, the pump increase / decrease torque conversion unit 218h, the integral control addition unit 218i, the primary delay element 218k, and the target displacement calculation unit 218m of the second target pump tilt angle calculation unit 218b are the engine load factor LF and When the deviation ΔLF from the target value (target engine load factor) LFr is not 0, the maximum absorption torque of the hydraulic pump 2 is controlled so that the deviation approaches 0, and the deviation between the engine load factor LF and the target value LFr. When ΔLF becomes 0, constant value control is performed so that the maximum absorption torque of the hydraulic pump 2 at that time is maintained and the deviation is kept at 0.
[0066]
Details of the fixed value control will be described with reference to FIGS. FIGS. 11 to 13 show details of a pq diagram created by the target displacement calculation unit 218m, the horizontal axis indicates the pump pressure P, and the vertical axis indicates the target displacement volume q.
[0067]
In FIG. 11, reference numeral 41 denotes a reference maximum pump absorption torque curve, which is set by Tp3 (target constant torque value Tp1 after rewriting) calculated when the load factor deviation ΔLF calculated by the subtraction unit 218g becomes zero. It is a maximum pump absorption torque curve. When the engine load factor LF is larger than the target engine load factor LFr and ΔLF <0, the pump increase / decrease torque conversion unit 218h converts the deviation ΔLF into the target increase / decrease torque increment ΔT (<0), and the integration control addition unit 218i. Then, the increment ΔT is added to the previous target constant torque value Tp1 to calculate a new target constant torque value Tp2, and the constant torque curve limiter calculation unit 218j calculates the target constant torque value Tp3 using the lower limit limiter as the lower limit. The target constant torque value Tp1 of the target displacement volume calculating unit 218m is rewritten. At this time, since ΔLF <0 and ΔT <0, Tp2 and Tp3 decrease, and the target constant torque value Tp1 rewritten by the target displacement calculating unit 218m also decreases, so the maximum pump absorption torque curve changes from 41 to 42. And move to the reduced torque side (reduced torque control). As a result, since the engine load decreases, the engine load factor LF also decreases, and ΔLF approaches zero. When the engine load factor matches the target engine load factor and ΔLF = 0, the increment ΔT also becomes 0, and the previous target constant torque value Tp1 (maximum pump absorption torque curve 42) is maintained.
[0068]
When the engine load factor LF is smaller than the target engine load factor LFr and ΔLF> 0, on the contrary, since ΔT> 0, Tp2, Tp3, Tp1 increase, and the maximum pump absorption torque curve increases from 41 to 43. Move to the torque side (torque increase control). As a result, since the engine load increases, the engine load factor LF also increases, ΔLF approaches 0, the engine load factor matches the target engine load factor, and when ΔLF = 0, the increment ΔT also becomes 0, and the previous target The constant torque value Tp1 (maximum pump absorption torque curve 42) is maintained.
[0069]
Further, in such constant value control, even when the target load factor LFr is the same, when the engine output is lower than the standard specification, the target constant torque value Tp1 when the load factor deviation ΔLF is kept at 0 decreases. The reference maximum pump absorption torque curve 41 moves to the reduced torque side. Conversely, when the engine output is higher than the standard specification, the target constant torque value Tp1 when the load factor deviation ΔLF is kept at 0 increases. The reference maximum pump absorption torque curve 41 is controlled to move to the increased torque side (variation correction control). Further, the target constant torque value Tp1 when the load factor deviation ΔLF is maintained at 0 is controlled to be a value corresponding to the target engine load factor LFr.
[0070]
For example, assume that the reference maximum pump absorption torque curve 41 shown in FIG. 11 is set when the target engine load factor is set to 90% and the engine output is in the standard specification. In this case, even if the target engine load factor is the same 90%, when the engine output is lower than the standard specification, if the reference maximum pump absorption torque curve is set to 41, ΔLF <0 and ΔT < Since it becomes 0, Tp2, Tp3, and Tp1 decrease, and the maximum pump absorption torque curve moves from 41 to 42 toward the reduced torque side (reduced torque control). When the engine load factor matches the target engine load factor and ΔLF = 0, the increment ΔT is also zero, and the target constant torque value Tp1 (maximum pump absorption torque curve 42) at that time is maintained. On the contrary, when the engine output is higher than the standard specification, if the target constant torque value Tp1 is set to the maximum pump absorption torque curve 41, ΔLF> 0 and ΔT> 0, so Tp2, Tp3, Tp1 are The maximum pump absorption torque curve increases from 41 to 43 toward the increasing torque side (increasing torque control). When the engine load factor matches the target engine load factor and ΔLF = 0, the increment ΔT is also zero, and the target constant torque value Tp1 (maximum pump absorption torque curve 42) at that time is maintained.
[0071]
FIG. 12 is a diagram showing a change in the maximum pump absorption torque curve as a reference when the target engine load factor is constant and the engine output is different from the standard specification. When the engine output is lower than the standard specification, the reference maximum pump absorption torque curve moves from 41 to 46 toward the reduced torque side. When the engine output is higher than the standard specification, the reference maximum pump absorption torque curve is The second target pump tilt angle calculation unit 218b performs torque reduction control or torque increase control based on the curves 46 and 47 with respect to increase / decrease in the engine load factor. (Variation correction control).
[0072]
Further, when the target engine load factor is switched from 90% to 70%, when the engine load factor is larger than 70% and ΔLF <0, ΔT <0 and Tp2, Tp3, Tp1 decrease as described above, The maximum pump absorption torque curve moves from 41 to 42 on the reduced torque side (reduced torque control). When the engine load factor matches the target engine load factor and ΔLF = 0, the increment ΔT is also zero, and the target constant torque value Tp1 (maximum pump absorption torque curve 42) at that time is maintained.
[0073]
FIG. 13 is a diagram showing a change in the maximum pump absorption torque curve as a reference when the target engine load factor is changed. When the target engine load factor is switched from 90% to 70%, the reference maximum pump absorption torque curve moves from 41 to 48 toward the reduced torque side, and the second target pump tilt angle calculation unit 218b receives the target engine load factor. With respect to the increase / decrease of the engine load factor with respect to (70%), the torque reduction control or the torque increase control is performed based on the curve 48.
[0074]
The operation in the present embodiment configured as described above will be described below.
[0075]
When the operator switches the mode selection switch 27 to the normal mode position, the target engine load factor setting unit 19b in FIG. 8 selects a normal value of 90% as the target engine load factor LFr. As described above, this normal value is obtained by controlling the hydraulic pump 2 so that the pump absorption torque becomes the base torque TB in the standard engine having the atmospheric pressure and the fuel quality in the standard state. This is the value when the rate is measured. The base torque TB is a design specification value determined by the specifications required for the work implement, and the target constant torque value Tp1 when the maximum absorption torque curve 41 in FIG. 11 is set corresponds to the base torque TB.
[0076]
Then, for example, if an unloading operation with a light load is performed and the engine load factor LF calculated by the engine load factor calculation unit 18a is LF = 50%, ΔLF = 90−50 = 40 in the subtraction unit 218g in FIG. (%) Is calculated, a new target constant torque value Tp2 is obtained from the ΔLF by the pump increase / decrease torque conversion unit 218h and the addition unit 218i of FIG. 9, and the constant torque curve limiter calculation unit 218j sets Tpmax as an upper limit. A constant torque value Tp3 is calculated. Here, the target constant torque value Tp3 is a value that changes in the increasing direction as Tp3> Tp1 with respect to the previous target constant torque value Tp1. As a result, the target displacement calculation unit 218m rewrites the target constant torque value Tp3 to Tp1, and the maximum pump absorption torque curve changes from 41 to 43 in FIG. 11 toward the increased torque side, for example, and the increased horsepower control is performed. As a result, the pump tilt (pump flow rate) has been reduced until the pump pressure is higher than PA, but the pump flow rate does not decrease unless the pressure is higher than PB (> PA), and the pump pressure P is lower than PB. In an operation in which the pressure increases from the above pressure, the operating speed of the actuator of the work implement does not decrease and the work efficiency is improved. The engine output is increased by the action of the horsepower control, and as a result, the engine load factor LF is increased. However, if ΔLF is not negative, the horsepower control is continued. When the upper limit of the horsepower is reached, no further horsepower is increased.
[0077]
Next, for example, when the pump pressure is high due to heavy work such as loading work + boom raising operation, if the output torque of the engine 1 (hereinafter referred to as engine torque as appropriate) suddenly drops due to some reason, the engine torque is sufficient. There is a possibility that the engine will stall. In this case, the engine load factor LF calculated by the engine load factor calculating unit 18a is 90% or more, for example, 95%, and ΔLF = 90−95 = −5 (%). A new target constant torque value Tp2 is obtained by the pump increase / decrease torque conversion unit 218h and the addition unit 218i, and the target constant torque value Tp3 is calculated by the constant torque curve limiter calculation unit 218j using Tpmin as the lower limit. Here, the target constant torque value Tp3 is a value that changes in a decreasing direction as Tp3 <Tp1 with respect to the previous target constant torque value Tp1. As a result, the target displacement calculation unit 218m rewrites the target constant torque value Tp3 to Tp1, and the maximum pump absorption torque curve changes from 41 to 42 in FIG. 11 on the reduced torque side, for example, and horsepower reduction control is performed. As a result, the horsepower reduction control is performed until the engine load factor reaches 90%, so that no engine stall occurs.
[0078]
On the other hand, in the case of a working machine operating at a high altitude, the engine output is lower than the standard specification. In this case, if the maximum absorption torque of the hydraulic pump 2 remains the base torque TB (the maximum absorption torque curve 41 in FIG. 11), the engine torque is not sufficient in heavy load work where the pump torque reaches the maximum absorption torque and the engine stalls. there's a possibility that. At this time, the engine load factor LF calculated by the engine load factor calculating unit 18a is 90% or more, for example, 95%, and ΔLF = 90−95 = −5 (%). Similarly to the time, the maximum displacement capacity calculation unit 218m moves the maximum pump absorption torque curve from 41 in FIG. 11 to the reduced torque side, and the horsepower reduction control is performed until the engine load factor becomes 90%. As a result, as shown in FIG. 12, the reference maximum pump absorption torque curve moves from 41 to 46 toward the reduced torque side so that the horsepower control or horsepower control is performed based on the curve 46. Therefore, even if the engine output falls below the standard specification, engine stall does not occur.
[0079]
Similarly, when the performance of the engine 1 is lower than the design performance, the horsepower reduction control is performed, and the reference maximum pump absorption torque curve moves from 41 to 46 in FIG. No (variation correction control).
[0080]
Conversely, when the performance of the engine 1 is higher than the design performance, the engine output increases from the standard specification. In this case, if the maximum absorption torque of the hydraulic pump 2 remains the base torque TB (the maximum absorption torque curve 41 in FIG. 11), an excessive margin is generated in the engine torque during heavy load work in which the pump torque reaches the maximum absorption torque. The engine output cannot be used effectively. At this time, the engine load factor LF calculated by the engine load factor calculation unit 18a is a value smaller than 90%, for example, 85%, and ΔLF = 90−85 = 5 (%). Similarly, in the target displacement calculating section 218m, the maximum pump absorption torque curve moves from 41 in FIG. 11 to the increased torque side, and the horsepower control is performed until the engine load factor reaches 90%. As a result, as shown in FIG. 12, the reference maximum pump absorption torque curve moves from 41 to 47 toward the increasing torque side, so that the horsepower increase control or horsepower reduction control based on the curve 47 is performed. Thus, the engine output can be effectively utilized for the engine whose output is higher than the standard specification (variation correction control).
[0081]
Next, a case where the operator operates the mode selection switch 27 to switch to the energy saving mode position will be described.
[0082]
When the operator switches the mode selection switch 27 to the energy saving mode position, the target engine load factor LFr is switched from 90% to 70% in the target engine load factor setting unit 19b of FIG. In such a setting state, if the engine load factor LF is greater than 70%, for example, 90%, ΔLF = 70−90 = −20 (%), so that the maximum displacement absorption by the target displacement volume calculation unit 218m. The horsepower reduction control is performed until the torque curve moves from 41 in FIG. 11 to the reduced torque side and the engine load factor becomes 70%. As a result, as shown in FIG. 13, the reference maximum pump absorption torque curve moves from 41 to 48 toward the reduced torque side, so that the horsepower increase control or the horsepower reduction control based on the curve 48 is performed. Thus, the maximum load applied to the engine 1 is reduced, and fuel consumption is suppressed.
[0083]
As described above, according to the present embodiment, the engine load factor is calculated, and the maximum absorption torque of the hydraulic pump 2 is controlled so that the engine load factor is maintained at the target value. When it is small, the torque increase control of the hydraulic pump 2 is performed, and when the engine load factor becomes larger than the target value, the torque reduction control of the hydraulic pump 2 is performed. As a result, regardless of the output characteristics of the engine, In other words, even with the engine 1 that performs isochronous control, it is possible to perform control equivalent to the horsepower increase control, horsepower reduction control, and variation correction control of the hydraulic pump performed by the conventional speed sensing control, and the engine output can be effectively utilized. Work efficiency can be improved.
[0084]
Further, since the engine load factor directly related to the load of the engine 1 is detected and control is performed so that the engine load factor is maintained at the target value, the engine load can be grasped more accurately than the conventional speed sensing control. As a result, control accuracy can be improved.
[0085]
Furthermore, by making it possible to change the target value of the engine load factor, the maximum absorption torque (maximum engine load) of the hydraulic pump 2 can be arbitrarily changed, and the target value of the engine load factor can be reduced by the mode selection switch 27. The energy saving mode can be set by reducing the maximum absorption torque of the hydraulic pump, and control with reduced fuel consumption can be easily performed.
[0086]
The second target pump tilt angle calculation unit 218b includes a subtraction unit 218g, a pump increase / decrease torque conversion unit 218h, an integral control addition unit 218i, a first-order lag element 218k, and a target displacement calculation unit 218m. When the deviation ΔLF from the target value (target engine load factor) LFr is not 0, the maximum absorption torque of the hydraulic pump 2 is controlled so that the deviation approaches 0, and the deviation between the engine load factor LF and the target value LFr. When ΔLF becomes 0, constant value control is performed so that the maximum absorption torque of the hydraulic pump 2 at that time is maintained and the deviation is kept at 0, so that the above-described horsepower reduction control and horsepower control can be performed stably, and the engine output When the engine is different from the standard specification or when the target engine load factor is changed, the maximum pump absorption torque curve can be set as a reference accordingly. It becomes.
[0087]
In addition, since the engine load factor is obtained using a relatively easy-to-detect state quantity such as a control rack position, a practical load factor detecting means can be provided.
[0088]
In the above embodiment, the present invention is applied to a work machine having an engine that performs isochronous control. However, the present invention is applied to a work machine having an engine having a mechanical governor having a droop characteristic or an electronic governor. Of course, you may do.
[0089]
Further, in the above embodiment, when the engine load factor with respect to the standard maximum absorption torque (base torque TB) of the engine in the standard state and the standard specification is obtained, the target engine speed is fixed to the rated engine speed NO. Varies depending on the engine speed, so that even if the same reference maximum absorption torque (base torque TB) is input to the engine, the engine load factor takes a different value if the target engine speed is changed. Considering this, measure the engine load factor when changing the target speed, create a map of "engine speed" and "engine load ratio when adding the standard maximum absorption torque", engine target The target engine load factor may be changed in accordance with the engine speed so that the hydraulic pump maintains the same reference maximum absorption torque even when the speed is changed by the dial.
[0090]
Further, for example, since the energy saving mode is a mode in which fuel efficiency is given the highest priority, the target engine load factor may be reduced from 90% to 70%, for example, and the target engine speed may be set lower than the rated speed NO. . Thereby, further fuel consumption reduction can be aimed at.
[0091]
Further, a heavy excavation mode, a standard excavation mode, an energy saving mode, and the like may be provided, and a target engine load factor may be set for each mode, and the target engine load factor may be changed when the mode is switched. For example, if the heavy excavation mode is set to 90%, the standard excavation mode is set to 80%, and the energy saving mode is set to 70%, both work efficiency and fuel consumption can be achieved according to the work mode. Further, not only the target engine load factor but also the target engine speed may be set for each mode, and both the target engine load factor and the target engine speed may be changed when the mode is switched.
[0092]
Further, when the engine output increases due to secular change or the like, the measurement for obtaining the engine load factor with respect to the reference maximum absorption torque in the standard state may be performed again to reset the target engine load factor (for example, 85%). Thereby, excessive performance can be prevented. The same applies when the engine output decreases.
[0093]
In the above embodiment, the mode selection switch 27 can be switched between the normal mode position and the energy saving mode position. However, the mode selection switch 27 can be continuously operated and the target engine load factor can be continuously changed. It may be.
[0094]
Further, in the above embodiment, the increase / decrease amount ΔT of the target increase / decrease torque is calculated by multiplying the deviation ΔLF between the engine load factor LF and its target value (target engine load factor) LFr by the proportional gain, but the calculation of ΔT from ΔLF. PID control may be used in the method. By adopting PID control, response characteristics can be improved in the case of an unstable control system.
[0095]
【The invention's effect】
According to the present invention, regardless of the output characteristics of the engine, it is possible to perform the same control as the horsepower increase control and the horsepower decrease control of the hydraulic pump that is performed by the conventional speed sensing control. Can be improved.
[0096]
In addition, the engine load factor directly related to the engine load is detected, and control is performed so that the engine load factor is maintained at the target value. Therefore, the engine load can be accurately grasped compared to the conventional speed sensing control, Control accuracy can be improved.
[0097]
Furthermore, by making it possible to change the target value of the engine load factor, the maximum absorption torque (maximum engine load) of the hydraulic pump can be arbitrarily changed, and control with reduced fuel consumption can be easily performed. .
[0098]
When the deviation between the engine load factor and the target value is not zero, the maximum pump absorption torque is controlled so that the deviation approaches zero, and when the deviation between the engine load factor and the target value becomes zero, Because constant value control is performed to maintain the maximum pump absorption torque and keep the deviation zero, stable horsepower control and horsepower control can be performed, and when the engine output is different from the standard specification or the target engine load factor is changed When this is done, it is possible to set a maximum pump absorption torque curve as a reference corresponding thereto.
[Brief description of the drawings]
FIG. 1 is a diagram showing an entire system of a hydraulic drive device for a working machine according to a first embodiment of the present invention.
FIG. 2 is an external view of a hydraulic excavator in which a hydraulic drive device according to the present embodiment is mounted.
FIG. 3 is a characteristic diagram showing the relationship between the rotational speed and output torque of an engine having an electronic governor that performs isochronous control.
FIG. 4 is a diagram showing a relationship between a control current signal given to an electromagnetic proportional pressure reducing valve of a regulator and a tilt angle of a hydraulic pump.
FIG. 5 is a block diagram illustrating a control flow of the entire system of the hydraulic drive device and a processing function of the work machine controller.
FIG. 6 is a diagram showing a relationship between a fuel flow rate and an engine load factor.
FIG. 7 is a diagram showing a relationship between a control rack position and an engine load factor.
FIG. 8 is a diagram showing a configuration of a target engine load factor setting unit.
FIG. 9 is a diagram illustrating a configuration of a pump torque control unit.
FIG. 10 is a diagram illustrating another example of a pump increase / decrease torque conversion unit.
FIG. 11 is a diagram illustrating a pq diagram created by a target displacement calculating unit.
FIG. 12 is a diagram showing changes in the reference maximum absorption torque when the engine output is lower and higher than the standard specification.
FIG. 13 is a diagram showing a change in the reference maximum absorption torque when the mode selection switch is operated.
[Explanation of symbols]
1 engine
2 Hydraulic pump
3-6 Hydraulic actuator
7-10 Directional control valve
11 Main relief valve
12 Electronic governor
13 Engine controller
16 Regulator
18 Work machine controller
18a Engine load factor calculation unit
18b Target engine load factor setting unit
18c Pump torque control unit by load factor control
24 Pressure detector
25 Tilt angle detector
26 Rack position detector
27 Mode selection switch
30 Characteristics of full load range
32 Governor characteristics
50 Control lever device
53 Signal control valve
55 Pressure detector
118a Normal target load factor storage unit
118b Target load factor storage unit for energy saving mode
118c switching unit
218a First target pump tilt angle calculation unit
218b Second target pump tilt angle calculation unit
218c Minimum value selection unit
218d subtraction unit
218e Control current calculation unit
218g subtraction part
218h Pump increase / decrease torque converter
218i Integration control adder
218j Constant torque curve limiter calculator
218k first order lag element
218m Target displacement calculation unit
218n Pump tilt angle converter

Claims (6)

Engine,
A fuel injection device that controls the rotational speed and output of the engine;
A fuel injector controller for controlling the fuel injector;
A hydraulic drive device for a working machine comprising at least one variable displacement hydraulic pump driven by the engine and driving an actuator;
First means for calculating a load factor of the engine;
A second means for controlling a maximum absorption torque of the hydraulic pump so that the load factor is maintained at a target value;
A hydraulic drive device for a working machine, comprising: a third means that can change the target value of the load factor. In the hydraulic drive device of the working machine according to claim 1,
When the deviation between the load factor and the target value is not zero, the second means controls the maximum absorption torque of the hydraulic pump so that the deviation approaches zero, and the deviation between the load factor and the target value is zero. Then, the hydraulic drive device for a working machine is means for performing constant value control so as to maintain the maximum absorption torque of the hydraulic pump at that time and keep the deviation at zero. In the hydraulic drive device of the working machine according to claim 1,
The second means includes a means for calculating a deviation between the load factor and a target value, a means for converting the deviation into an increment of a target constant torque value, and adding the increment to the previous target constant torque value. And a means for calculating a desired target constant torque value, and the maximum absorption torque of the hydraulic pump is controlled using the new target constant torque value. In the hydraulic drive device of the working machine according to claim 1,
The first means detects a control rack position of the fuel injection device;
Engine load factor (%) = [(current control rack position-control rack position at no load) / (control rack position at full load-control rack position at no load)] x 100
In the above formula, the control rack position at full load and the control rack position at no load are values obtained and stored in advance.
A hydraulic drive device for a working machine, wherein the load factor is calculated by the following formula. In the hydraulic drive device of the working machine according to claim 1,
The hydraulic drive device for a working machine, wherein the third means is a means operated by an operator to selectively set a normal load factor target value and a load factor target value for energy saving smaller than the normal load factor target value. . Engine,
A fuel injection device that controls the rotational speed and output of the engine;
A fuel injector controller for controlling the fuel injector;
A hydraulic drive device for a working machine comprising at least one variable displacement hydraulic pump driven by the engine and driving an actuator;
First means for calculating a load factor of the engine;
Second means for controlling the maximum absorption torque of the hydraulic pump so that the load factor is maintained at a target value;
When the deviation between the load factor and the target value is not zero, the second means controls the maximum absorption torque of the hydraulic pump so that the deviation approaches zero, and the deviation between the load factor and the target value is zero. Then, the hydraulic drive device for a working machine is means for performing constant value control so as to maintain the maximum absorption torque of the hydraulic pump at that time and keep the deviation at zero.

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