Engine control system for construction machine

An engine control system includes pressure sensors (73, 74), position sensors (75, 76), pressure sensors (77, 78), a target revolution speed modification value computing unit (90), and a modification value adder (70r). A target revolution speed NR2 for use in control is computed based on changes of status variables such that the target revolution speed NR2 increases from the target revolution speed NR1 applied from an input unit (71), and then moderately returns to the target revolution speed NR1. In accordance with the computed target revolution speed NR2 for use in control, a target fuel injection amount FN1 is computed and a fuel injection amount is controlled. As a result, a drop of an engine revolution speed attributable to an abrupt increase of an engine load can be suppressed without sacrificing the work efficiency, and lowering of durability caused by an excessive increase of the engine revolution speed can be prevented.

TECHNICAL FIELD

The present invention relates to an engine control system for a construction machine, and more particularly to an engine control system for a construction machine in which a variable displacement hydraulic pump is driven by a diesel engine to drive a hydraulic actuator.

BACKGROUND ART

In general, a construction machine such as a hydraulic excavator comprises an engine, at least one variable displacement hydraulic pump driven by the engine, a plurality of hydraulic actuators driven by a hydraulic fluid delivered from the hydraulic pump, a plurality of flow control valves for controlling respective flow rates of the hydraulic fluid supplied from the hydraulic pump to the plurality of hydraulic actuators, and a plurality of control lever devices serving as operating means to operate the plurality of flow control valves. Also, a diesel engine is employed as the engine for driving the hydraulic pump. The diesel engine is equipped with a fuel injector, called a governor, to control an amount of fuel injected, thereby controlling a revolution speed of the engine.

In such a diesel engine equipped with a fuel injector, when a control lever of the control lever device is quickly manipulated for shift of the flow control valve, an input torque (load) of the hydraulic pump is abruptly increased and the engine revolution speed abruptly drops. This abrupt drop of the engine revolution speed leads to problems of not only deteriorating fuel consumption and exhaust gas, but also causing noises.

Techniques for suppressing such a drop of the engine revolution speed are disclosed in, for example, JP,A 2000-154803 and JP,A 2001-173605.

With the technique disclosed in JP,A 2000-154803, the load state of a hydraulic pump is detected, and when it is detected that a load is applied to the hydraulic pump, a limit value for the input torque of the hydraulic pump is reduced to perform torque decrease control. As a result, the absorption torque of the hydraulic pump (i.e., the engine load) is reduced so as to suppress the drop of the engine revolution speed.

With the technique disclosed in JP,A 2001-173605, the operating speed of a control lever is detected, and when the operating speed exceeds a predetermined value, fuel is supplied in an increased amount to an engine in response to a command signal from a controller. As a result, the engine output is increased so as to suppress the drop of the engine revolution speed.

DISCLOSURE OF THE INVENTION

However, the above-described known techniques have problems as follows.

With the technique disclosed in JP,A 2000-154803, because the drop of the engine revolution speed is suppressed by reducing the absorption torque of the hydraulic pump, the delivery rate of the hydraulic pump is also reduced and so is the actuator speed correspondingly. Hence, an amount of feasible work is reduced and the work efficiency is sacrificed.

The technique disclosed in JP,A 2001-173605 is intended to suppress the drop of the engine revolution speed by supplying the fuel in an increased amount to the engine so that the engine output is increased. However, the engine revolution speed cannot be controlled with an increase of the fuel amount, and there is a possibility that the engine revolution speed goes up beyond a required level. In some cases, the engine revolution speed may exceed a critical level in terms of durability.

It is an object of the present invention to provide an engine control system for a construction machine, which can suppress a drop of an engine revolution speed attributable to an abrupt increase of an engine load without sacrificing the work efficiency, and can prevent lowering of durability caused by an excessive increase of the engine revolution speed.(1) To achieve the above object, the present invention provides an engine control system for a construction machine comprising an engine, at least one variable displacement hydraulic pump driven by the engine, a plurality of hydraulic actuators driven by a hydraulic fluid delivered from the hydraulic pump, a plurality of flow control valves for controlling respective flow rates of the hydraulic fluid supplied from the hydraulic pump to the plurality of hydraulic actuators, operating means for operating the plurality of flow control valves, a fuel injector for controlling a revolution speed of the engine, input means for commanding a target revolution speed of the engine, and fuel injection amount control means for computing a target fuel injection amount based on the target revolution speed and controlling the fuel injector, wherein the engine control system comprises status variable detecting means for detecting a status variable related to a load of the hydraulic pump, and target revolution speed modifying means for computing a target revolution speed for use in control based on a change of the status variable such that the target revolution speed for use in control increases from the target revolution speed set in accordance with a command from the input unit, and then moderately returns to the target revolution speed set in accordance with the command from the input unit, the fuel injection amount control means computing the target fuel injection amount based on the target revolution speed for use in control.

Thus, the status variable detecting means and the target revolution speed modifying means are provided, and the target revolution speed for use in control is increased depending on the change of the status variable, whereby an actual revolution speed is also increased correspondingly. It is therefore possible to suppress a drop of the engine revolution speed when an engine load is abruptly increased. Also, since the control process is performed on the basis of engine revolution speed, the absorption torque of the hydraulic pump is not reduced and the work efficiency is not sacrificed. Further, the target revolution speed for use in control is computed based on the change of the status variable so as to increase from the target revolution speed set in accordance with the command from the input means and then moderately return to the target revolution speed set in accordance with the command from the input means, and the engine revolution speed is controlled in accordance with the target revolution speed thus computed. As a result, the engine revolution speed can be avoided from going up beyond a required level, and lowering of durability caused by an excessive increase of the engine revolution speed can be prevented.(2) In above (1), preferably, the target revolution speed modifying means maintains the increased engine revolution speed for a certain time after the change of the status variable has ceased.

With that feature, a drop of the engine revolution speed attributable to an abrupt increase of the engine load can be suppressed with higher certainty.(3) In above (1), preferably, the target revolution speed modifying means computes an increase amount of the target revolution speed as a variable value depending on the target revolution speed set in accordance with the command from the input unit.

With that feature, as the target revolution speed set in accordance with the command from the input means changes, the increase amount of the target revolution speed is also changed correspondingly. Therefore, an optimum increase amount of the target revolution speed can be computed regardless of the target revolution speed.(4) In above (1), preferably, the target revolution speed modifying means includes means for computing, based on the change of the status variable, an engine revolution speed modification value which increases from 0 by a predetermined amount and then moderately returns to 0, and means for adding the engine revolution speed modification value to the target revolution speed set in accordance with the command from the input unit.

With that feature, depending on the change of the status variable, the target revolution speed for use in control increases from the target revolution speed set in accordance with the command from the input means and then moderately returns to the target revolution speed set in accordance with the command from the input means.(5) In above (1), preferably, the status variable detecting means detects, as the status variable related to the load of the hydraulic pump, at least one of operation signals from the operating means, a delivery capacity of the hydraulic pump, and a delivery pressure of the hydraulic pump.

With that feature, the load state of the hydraulic pump can be detected with high accuracy.

BEST MODE FOR CARRYING OUT THE INVENTION

An embodiment of the present invention will be described below with reference to the drawings. In the following embodiment, the present invention is applied to an engine control system for a hydraulic excavator.

A first embodiment of the present invention will be first described with reference toFIGS. 1 to 8.

InFIG. 1, reference numerals1and2denote variable displacement hydraulic pumps of, e.g., swash plate type. Numeral9denotes a fixed displacement pilot pump. The hydraulic pumps1,2and the pilot pump9are connected to an output shaft11of a prime mover10and are driven by the prime mover10for rotation.

A valve unit5, shown inFIG. 2, is connected to delivery lines3,4of the hydraulic pumps1,2. A hydraulic fluid is supplied to a plurality of actuators50-56through the valve unit5, thereby driving the actuators. A pilot relief valve9bfor holding the delivery pressure of the pilot pump9at a certain pressure is connected to a delivery line9aof the pilot pump9.

Details of the valve unit5will be described below.

InFIG. 2, the valve unit5has two valve groups comprising respectively flow control valves5a-5dand flow control valves5e-5i. The flow control valves5a-5dare positioned on a center bypass line5jconnected to the delivery line3of the hydraulic pump1, and the flow control valves5e-5iare positioned on a center bypass line5kconnected to the delivery line4of the hydraulic pump2. A main relief valve5mfor deciding a maximum value of the delivery pressure of the hydraulic pumps1,2is disposed in the delivery lines3,4.

The flow control valves5a-5dand the flow control valves5e-5iare each of the center bypass type. The hydraulic fluid delivered from the hydraulic pumps1,2is supplied to corresponding one or more of the actuators50-56through the associated flow control valves. The actuator50is a hydraulic motor for travel on the right side (i.e., a right travel motor), and the actuator51is a hydraulic cylinder for a bucket (i.e., a bucket cylinder). The actuator52is a hydraulic cylinder for a boom (i.e., a boom cylinder), and the actuator53is a hydraulic motor for swing (i.e., a swing motor). The actuator54is a hydraulic cylinder for an arm (i.e., an arm cylinder), the actuator55is a reserve hydraulic cylinder, and the actuator56is a hydraulic motor for travel on the left side (i.e., a left travel motor). The flow control valve5aserves for the travel on the right side, and the flow control valve5bserves for the bucket. The flow control valve5cserves for a first boom, and the flow control valve5dserves for a second arm. The flow control valve5eserves for the swing, the flow control valve5fserves for a first arm, and the flow control valve5gserves for a second boom. The flow control valve5hserves for reserve, and the flow control valve5iserves for the travel on the left side. Stated another way, two flow control valves5g,5care disposed in association with the boom cylinder52and two flow control valves5d,5fare disposed in association with the arm cylinder54, whereby respective hydraulic fluids from the two hydraulic pumps1,2can be supplied in a joined way to the bottom side of each of the boom cylinder52and the arm cylinder54.

FIG. 3shows an operation pilot system for the flow control valves5a-5i.

The flow control valves5i,5aare operated for position shift by operation pilot pressures TR1, TR2; TR3, TR4produced from operation pilot devices39,38of an operating unit35. The flow control valve5band the flow control valves5c,5gare operated for position shift by operation pilot pressures BKC, BKD; BOD, BOU produced from operation pilot devices40,41of an operating unit36. The flow control valves5d,5fand the flow control valve5eare operated for position shift by operation pilot pressures ARC, ARD; SW1, SW2produced from operation pilot devices42,43of an operating unit37. The flow control valve5his operated for position shift by operation pilot pressures AU1, AU2produced from an operation pilot device44.

The operation pilot devices38-44have pairs of pilot valves (pressure reducing valves)38a,38b-44a,44b, respectively. Further, the operation pilot devices38,39and44have control pedals38c,39cand44c, respectively. The operation pilot devices40,41have a common control lever40c, and the operation pilot devices42,43have a common control lever42c. When any of the control pedals38c,39cand44cand the control levers40c,42cis manipulated, the pilot valve of the associated operation pilot device corresponding to the direction of the manipulation is operated and an operation pilot pressure is produced depending on an input amount by which the control pedal or lever is manipulated.

Shuttle valves61-67, shuttle valves68,69and100, shuttle valves101,102, and a shuttle valve103are connected in a hierarchical arrangement to output lines of the respective pilot valves of the operation pilot devices38-44. The shuttle valves61,63,64,65,68,69and101cooperate to detect a maximum one of the operation pilot pressures from the operation pilot devices38,40,41and42as a control pilot pressure PP1for the hydraulic pump1, whereas the shuttle valves62,64,65,66,67,69,100,102and103cooperate to detect a maximum one of the operation pilot pressures from the operation pilot devices39,41,42,43and44as a control pilot pressure PP2for the hydraulic pump2.

An engine/pump control system including the engine control system of the present invention is applied to a hydraulic drive system thus constructed. Details of the engine/pump control unit will be described below.

InFIG. 1, the hydraulic pumps1,2are provided with regulators7,8, respectively. The regulators7,8regulate tilting positions of swash plates1a,2athat serve as displacement varying mechanisms of the hydraulic pumps1,2, thereby controlling respective pump delivery rates.

The regulators7,8for the hydraulic pumps1,2comprise respectively tilting actuators20A,20B (hereinafter represented by20as appropriate), first servo valves21A,21B (hereinafter represented by21as appropriate) for performing positive tilting control in accordance with the operation pilot pressures from the operation pilot devices38-44shown inFIG. 3, and second servo valves22A,22B (hereinafter represented by22as appropriate) for performing total horsepower control of the hydraulic pumps1,2. Those servo valves21,22control the pressure of a hydraulic fluid supplied from the pilot pump9and acting upon the respective tilting actuators20, thereby controlling the tilting positions of the hydraulic pumps1,2.

Details of the tilting actuators20and the first and second servo valves21,22will be described below.

Each tilting actuator20comprises an working piston20chaving a large-diameter pressure bearing portion20aand a small-diameter pressure bearing portion20bwhich are formed at opposite ends thereof, and a large-diameter pressure bearing chamber20dand a small-diameter pressure bearing chamber20ein which the pressure bearing portions20a,20bare positioned respectively. When the pressures in both the pressure bearing chambers20d,20eare equal to each other, the working piston20cis moved to the right, as viewed inFIG. 1, due to a difference in pressure bearing area, whereupon the tilting of the swash plate1aor2ais reduced to decrease the pump delivery rate. When the pressure in the large-diameter pressure bearing chamber20dlowers, the working piston20cis moved to the left, as viewed inFIG. 1, whereupon the tilting of the swash plate1aor2ais enlarged to increase the pump delivery rate. Further, the large-diameter pressure bearing chamber20dis selectively connected through the first and second servo valves21,22to one of the delivery line9aof the pilot pump9and a return fluid line13leading to a reservoir12. The small-diameter pressure bearing chamber20eis directly connected to the delivery line9aof the pilot pump9.

Each first servo valve21for the positive tilting control is a valve operated by a control pressure from a solenoid control valve30or31to control the tilting position of the hydraulic pump1or2. When the control pressure is low, a valve member21aof the servo valve21is moved to the left, as viewed inFIG. 1, by the force of a spring21b, whereupon the large-diameter pressure bearing chamber20dof the tilting actuator20is communicated with the reservoir12via the return fluid line13to increase the tilting of the hydraulic pump1or2. When the control pressure rises, the valve member21aof the servo valve21is moved to the right, as viewed inFIG. 1, whereupon the pilot pressure from the pilot pump9is introduced to the large-diameter pressure bearing chamber20dto decrease the tilting of the hydraulic pump1or2.

Each second servo valve22for the total horsepower control is a valve operated by both the delivery pressure of the hydraulic pump1or2and a control pressure from a solenoid control valve32to perform the total horsepower control of the hydraulic pump1or2. In other words, the second servo valve22controls a maximum absorption torque of the hydraulic pump1or2in accordance with the control pressure from the solenoid control valve32.

More specifically, the delivery pressures of the hydraulic pumps1,2and the control pressure from the solenoid control valve32are introduced respectively to pressure bearing chambers22a,22band22cof the second servo valve22. When the sum of hydraulic forces of the delivery pressures of the hydraulic pumps1,2is smaller than a setting value that is determined depending on a difference between the force of a spring22dand the hydraulic force of the control pressure introduced to the pressure bearing chamber22c, a valve member22eis moved to the right, as viewed inFIG. 1, whereupon the large-diameter pressure bearing chamber20dof the tilting actuator20is communicated with the reservoir12via the return fluid line13to increase the tilting of the hydraulic pump1or2. As the sum of hydraulic forces of the delivery pressures of the hydraulic pumps1,2increases in excess of the above-mentioned setting value, the valve member22eis moved to the left, as viewed inFIG. 1, whereupon the pilot pressure from the pilot pump9is transmitted to the pressure bearing chamber20dto decrease the tilting of the hydraulic pump1or2. Further, when the control pressure from the solenoid control valve32is low, the above-mentioned setting value is increased so that the tilting of the hydraulic pump1or2starts to decrease from a relatively high delivery pressure of the hydraulic pump1or2. As the control pressure from the solenoid control valve32rises, the above-mentioned setting value is reduced so that the tilting of the hydraulic pump1or2starts to decrease from a lower delivery pressure of the hydraulic pump1or2.

FIG. 4shows characteristics of absorption torque control performed by the second servo valve22. InFIG. 4, the horizontal axis represents an average value of the delivery pressures of the hydraulic pumps1,2, and the vertical axis represents the tilting (displacement) of the hydraulic pump1or2. As the control pressure from the solenoid control valve32rises (i.e., as the setting value determined depending on the difference between the force of the spring22dand the hydraulic force introduced to the pressure bearing chamber22creduces), an absorption torque characteristic of the second servo valve22changes as indicated by A1, A2and A3in this order, and a maximum absorption torque of the hydraulic pump1or2decreases as indicated by T1, T2and T3in this order. Also, as the control pressure from the solenoid control valve32lowers (i.e., as the setting value determined depending on the difference between the force of the spring22dand the hydraulic force introduced to the pressure bearing chamber22cincreases), the absorption torque characteristic of the second servo valve22changes as indicated by A1, A4and A5in this order, and the maximum absorption torque of the hydraulic pump1or2increases as indicated by T1, T4and T5in this order. In other words, by raising the control pressure to reduce the setting value, the maximum absorption torque of the hydraulic pump1or2decreases, and by lowering the control pressure to increase the setting value, the maximum absorption torque of the hydraulic pump1or2increases.

The solenoid control valves30,31and32are proportional pressure reducing valves operated by drive currents SI1, SI2and SI3, respectively. The solenoid control valves30,31and32operate so as to maximize output control pressures when the drive currents SI1, SI2and SI3are minimum, and to lower the output control pressures as the drive currents SI1, SI2and SI3increase. The drive currents SI1, SI2and SI3are outputted from a machine body controller70shown inFIG. 5.

The prime mover10is a diesel engine and includes an electronic fuel injector14operated in response to a signal indicative of a target fuel injection amount FN1. The command signal is outputted from a fuel injector controller80shown inFIG. 5. The electronic fuel injector14controls the revolution speed and output of the prime mover (hereinafter referred to as an “engine”)10.

There is provided a target engine revolution speed input unit71through which the operator manually inputs a target revolution speed NR1for the engine10. An input signal indicative of the target revolution speed NR1is taken into the machine body controller70and the engine fuel injector controller80. The target engine revolution speed input unit71is an electrical input means, such as a potentiometer, and the operator instructs a target revolution speed as a reference (i.e., a target reference revolution speed).

Further, there are provided a revolution speed sensor72for detecting an actual revolution speed NE1of the engine10, pressure sensors73,74(seeFIG. 3) for detecting the respective control pilot pressures PP1, PP2for the hydraulic pumps1,2, pressure sensors75,76for detecting respective tiltings SR1, SR2of the hydraulic pumps1,2, and pressure sensors77,78(seeFIG. 3) for detecting respective delivery pressures DP1, DP2of the hydraulic pumps1,2.

FIG. 5shows input and output relationships of all signals to and from the machine body controller70and the fuel injector controller80.

The machine body controller70receives a signal indicative of the target revolution speed NR1from the target engine revolution speed input unit71, signals indicative of the pump control pilot pressures PP1, PP2from the pressure sensors73,74, signals indicative of the tiltings SR1, SR2from the pressure sensors75,76, and signals indicative of the pump delivery pressures DP1, DP2from the pressure sensors77,78. After executing predetermined arithmetic processing based on those input signals, the machine body controller70outputs the drive currents SI1, SI2and SI3to the solenoid control valves30-32, respectively, and it also outputs the signal indicative of the target revolution speed NR1to the fuel injector controller80. The engine fuel injector controller80receives the signal indicative of the target revolution speed NR1from the machine body controller70and a signal indicative of the actual revolution speed NE1from the revolution speed sensor72. After executing predetermined arithmetic processing based on those input signals, the fuel injector controller80outputs a signal indicative of the target fuel injection amount FN1to the electronic fuel injector14.

FIGS. 6 and 7show the processing functions of the machine body controller70in relation to control of the hydraulic pumps1,2and computation of the target revolution speed NR1.

Referring toFIG. 6, the machine body controller70has various functions executed by pump target tilting computing units70a,70b, solenoid output current computing units70c,70d, an engine load increase amount computing unit70f, an engine revolution speed increase gain computing unit70g, a multiplier70h, an engine revolution speed increment value selector70i, a primary delay element70j, a subtracter70k, a subtracter70m, a gain multiplier70n, an integral adder70p, a primary delay element70q, a modification value adder70r, a base torque computing unit70s, and a solenoid output current computing unit70t.

The pump target tilting computing unit70areceives the signal indicative of the control pilot pressure PP1on the side of the hydraulic pump1and computes a target tilting θR1of the hydraulic pump1corresponding to the control pilot pressure PP1at that time by referring to a table, which is stored in a memory, based on the input signal. The computed target tilting θR1serves as a basis for metering of a reference flow rate in the positive tilting control with respect to the input amounts by which the pilot operation devices38,40,41and42are manipulated. The table stored in the memory sets therein the relationship between PP1and θR1such that, as the control pilot pressure PP1rises, the target tilting θR1is also increased.

The solenoid output current computing unit70cdetermines, on the computed θR1, the drive current SI1for the tilting control of the hydraulic pump1at which that θR1is obtained, and then outputs the determined drive current SI1to the solenoid control valve30.

Also, in the pump target tilting computing unit70band the solenoid output current computing unit70d, the drive current SI2for the tilting control of the hydraulic pump2is computed from the signal indicative of the pump control pilot pressure PP2, and then outputted to the solenoid control valve31in a similar manner.

The engine load increase amount computing unit70f, the engine revolution speed increase gain computing unit70g, the multiplier70h, the engine revolution speed increment value selector70i, the primary delay element70j, the subtracter70k, the subtracter70m, the gain multiplier70n, the integral adder70p, and the primary delay element70qconstitute a means90(hereinafter referred to as a “revolution speed modification value computing unit”) for computing the increase amount of the engine revolution speed, as a revolution speed modification value ΔT3, based on respective change rates of the control pilot pressures PP1, PP2, the pump tiltings SR1, SR2, and the pump delivery pressures DP1, DP2, which are status variables related to the loads of the hydraulic pumps1,2. The modification value adder70radds the revolution speed modification value ΔT3to the target engine revolution speed NR1applied from the input unit71, and then inputs the resulting sum, as a target engine revolution speed NR2for use in the control, to the base torque computing unit70s. These points will be described in more detail below.

The engine load increase amount computing unit70freceives the status variables regarding the load of each hydraulic pump, and computes an engine load increase amount ΔT1.

FIG. 7shows details of the processing functions of the engine load increase amount computing unit70f. The engine load increase amount computing unit70fhas the functions executed by primary delay elements701a,701b,701c,701d,701eand701f, subtracters702a,702b,702c,702d,702eand702f, gain multipliers703a,703b,703c,703d,703eand703f, filtering units704a,704b,704c,704d,704eand704f, adders705a,705band705c, as well as a filtering unit706.

The engine load increase amount computing unit70freceives the signals indicative of the control pilot pressures PP1, PP2, the signals indicative of the pump tiltings SR1, SR2, and the signals indicative of the pump delivery pressures DP1, DP2, and computes respective input speeds of those signals by taking the differences between the previous and current input values in the subtracters702a-702f. The computed input speeds represent change rates of the corresponding status variables. Then, the input speeds are multiplied by respective gains Knn in the gain multipliers703a-703f, and the resulting values are obtained as load increase amounts. Then, the signals are introduced to the filtering units704a-704fto pass through respective filters such that the load increase amounts are made zero when their changes are small. The filtered load increase amounts are totalized by the adders705a-705c. Finally, the filtering unit706allows only a positive value of the totalized load increase amount, which represents the load increasing direction, to pass through it, thereby obtaining the positive value as the load increase amount ΔT1.

Returning toFIG. 6, the engine revolution speed increase gain computing unit70gcomputes a gain KΔT1as a function of the target revolution speed NR1inputted to it. The gain KΔT1is multiplied by the load increase amount ΔT1in the multiplier70hto obtain an engine revolution speed increase amount ΔT2. The engine revolution speed increase gain computing unit70gstores the relationship between NR1and KΔT1set such that the gain KΔT1reduces as the target revolution speed NR1decreases. Accordingly, when the target revolution speed NR1is low, the gain KΔT1is set to a relatively small value and the engine revolution speed increase amount ΔT2is computed as a relatively small value in the multiplier70h.

The subtracter70kcomputes the difference between the current value of the engine revolution speed increase amount ΔT2and the previous value thereof which is supplied from the primary delay element70j, to thereby produce a determination value α. The determination value α takes a positive, negative or zero (0) value depending on the presence or absence of change of the engine revolution speed increase amount ΔT2and the direction of the change. More specifically, the determination value α takes a positive value when the engine revolution speed increase amount ΔT2is changed in the increasing direction, and a negative value when it is changed in the decreasing direction. Also, the determination value α is 0 when the engine revolution speed increase amount ΔT2is not changed (i.e., when it is constant).

The engine revolution speed increment value selector70idetermines whether the determination value α is positive, negative or 0, and it switches over an engine revolution speed increment value ΔT2A, which is applied to the subtracter70m, depending on the determination result. If α≧0 (namely if the engine revolution speed increase amount ΔT2is changed in the increasing direction, or if ΔT2is not changed), the selector70iis held in a state B to select the engine revolution speed increase amount ΔT2so that the engine revolution speed increase amount ΔT2is outputted as the increment value ΔT2A applied to the subtracter70m. If α<0 (namely if the engine revolution speed increase amount ΔT2is changed in the decreasing direction), the selector70itakes a state A to select 0 as the increment value ΔT2A applied to the subtracter70m. At the time of switching from the state B to A, the operation is delayed for a certain time (e.g., 3 seconds) to provide the hold function of maintaining the previous value.

The subtracter70msubtracts a revolution speed modification value ΔT4in the previous cycle from the increment value ΔT2A selected by the engine revolution speed increment value selector70i, thereby obtaining a deviation ΔΔT2.

The gain multiplier70nserves to give the deviation ΔΔT2a primary delay. A primary delay gain is set to 1 when the deviation ΔΔT2is in the increasing direction (i.e., ΔΔT2≧0), and to a value smaller than 1 when the deviation ΔΔT2is in the decreasing direction (i.e., ΔΔT2<0). The gain is multiplied by ΔΔT2to obtain a deviation ΔΔT4.

The integral adder70padds ΔΔT4to the revolution speed modification value ΔT4in the previous cycle which is supplied from the primary delay element70q, thereby obtaining the revolution speed modification value ΔT3in the current cycle.

The revolution speed modification value ΔT3thus computed is applied to the modification value adder70r, and the modification value adder70radds the revolution speed modification value ΔT3to the target revolution speed NR1, thereby obtaining the target revolution speed command NR2for use in the control.

The base torque computing unit70sreceives the target revolution speed command NR2from the modification value adder70r, and computes a pump base torque TR0corresponding to the target revolution speed command NR2at that time by referring to a table, which is stored in a memory, based on the input signal. The solenoid output current computing unit70tdetermines the drive current SI3for the solenoid control valve32at which the maximum absorption torque of the hydraulic pump1,2controlled by the second servo valve22becomes TR0, and then outputs the determined drive current SI3to the solenoid control valve32.

The solenoid control valve32having received the drive current SI3in such a way outputs a control pressure corresponding to the received drive current SI3and controls the setting value in the second servo valve22, thereby controlling the maximum absorption torque of the hydraulic pump1,2to be TR0.

FIG. 8shows the processing functions of the fuel injector controller80.

The fuel injector controller80has the control functions executed by a revolution speed deviation computing unit80a, a fuel injection amount converting unit80b, an integral adder80c, a limiter computing unit80d, and a primary delay element80e.

The revolution speed deviation computing unit80acompares the target revolution speed NR2with the actual revolution speed NE1to obtain a revolution speed deviation ΔN (=NR2−NE1), and the fuel injection amount converting unit80bmultiplies the revolution speed deviation ΔN by a gain KF to compute an increment ΔFN of the target fuel injection amount. The integral adder80cadds the increment ΔFN of the target fuel injection amount to the previous value FN2of the target fuel injection amount FN1which is supplied from the primary delay element80e, thereby computing a new target fuel injection amount FN3. The limiter computing unit80dmultiplies the target fuel injection amount FN3by upper and lower limiters to obtain the target fuel injection amount FN1. This target fuel injection amount FN1is converted to a corresponding control current that is outputted to the electronic fuel injector14for control of the fuel injection amount. With such a process, the target fuel injection amount FN1is computed through the integral operation such that when the actual revolution speed NE1is lower than the target revolution speed NR2(i.e., when the revolution speed deviation ΔN is positive), the target fuel injection amount FN1is increased, and when the actual revolution speed NE1exceeds the target revolution speed NR2(i.e., when the revolution speed deviation ΔN becomes negative), the target fuel injection amount FN1is decreased, i.e., such that the deviation ΔN of the actual revolution speed NE1from the target revolution speed NR2becomes 0. The fuel injection amount is thereby controlled so as to make the actual revolution speed NE1matched with the target revolution speed NR2.

Features in operation of this embodiment having the above-described construction will be described below with reference toFIGS. 9 and 10.

FIG. 9is a time chart showing changes of the engine revolution speed responsive to changes of an operation input in the prior art, andFIG. 10is a time chart showing changes of the engine revolution speed responsive to changes of an operation input in this embodiment. In each ofFIGS. 9 and 10, individual charts indicate the pump control pilot pressure PP1or PP2(represented by PP hereinafter), the pump delivery pressure DP1or DP2(represented by DP hereinafter), the pump tilting SR1or SR2(represented by SR hereinafter), the target revolution speed NR1(FIG. 9) or NR2(FIG. 10), and the actual engine revolution speed NE1in this order from above. The pump control pilot pressure PP is a value corresponding to a lever input amount applied from any of the operation pilot devices38-44shown inFIG. 3. Also, it is assumed that the target revolution speed NR1applied from the input unit71is constant, and that the control lever is slightly manipulated at time t1, quickly manipulated at time t2, and then stopped at time t3. It is further assumed that, during each of periods from the time t1to t2and from the time t2to t3, respective change rates of the pump control pilot pressure PP, the pump delivery pressure DP, and the pump tilting SR are constant.

In the prior art, as shown inFIG. 9, when the control lever is slightly manipulated at the time t1, the engine revolution speed drops in small amount. However, when the control lever is quickly manipulated at the time t2, the pump delivery pressure DP and the pump tilting SR are quickly increased correspondingly, whereupon the actual engine revolution speed NE1drops abruptly. At this time, a drop amount of the actual revolution speed NE1is large.

In contrast, according to this embodiment, when the control lever is quickly manipulated at the time t2, the target revolution speed command NR2is modified by the revolution speed modification value computing unit90such that the target revolution speed increases from the target revolution speed NR1applied from the input unit71, and then it moderately returns to the target revolution speed NR1. Therefore, an abrupt drop of the actual engine revolution speed NE1is avoided and the speed drop amount is reduced. Details of such a process are as follows.

From time t1to t2:

During this period, since the control lever is slightly manipulated, the respective change rates of the pump control pilot pressure PP, the pump delivery pressure DP, and the pump tilting SR are so small that the signals inputted to the filtering units704a-704fof the engine load increase amount computing unit70f, shown inFIG. 7, are processed to become zero through the respective filters. In this case, therefore, the load increase amount ΔT1computed in the engine load increase amount computing unit70fis 0 and the revolution speed modification value ΔT3is also 0, whereby the target revolution speed NR2(=NR1) is constant. As a result, the actual engine revolution speed NE1changes in the same manner as in the prior art.

From time t2to t3:

During this period, since the control lever is quickly manipulated, the load increase amount ΔT1is computed as a value other than 0 in the engine load increase amount computing unit70f, and the multiplier70hmultiplies the load increase amount ΔT1by the gain KΔT1depending on the target revolution speed NR1at that time to compute the engine revolution speed increase amount ΔT2.

In the first cycle of an arithmetic operation process at the time t2, the previous value of the engine revolution speed increase amount ΔT2is zero. Therefore, the subtracter70kcomputes a positive determination value α, and the engine revolution speed increment value selector70itakes the state B so that the engine revolution speed increase amount ΔT2computed by the multiplier70his introduced as the increment value ΔT2A to the subtracter70m. Further, because the previous value of the revolution speed modification value ΔT3is zero, the subtracter70mcomputes the increment value ΔT2(=the engine revolution speed increase amount ΔT2) as the deviation ΔΔT2, and the gain multiplier70noutputs the deviation ΔΔT4(=ΔΔT2) as a value resulting from multiplying the deviation ΔΔT2by the gain of 1. The deviation ΔΔT4is applied to the integral adder70p. At this time, the deviation ΔΔT4is given as the revolution speed modification value ΔT3because the previous value of the revolution speed modification value ΔT3is zero. Thus, as shown inFIG. 10, the target revolution speed NR2is increased by a value corresponding to ΔT3at the time t2.

During the period from the time t2to t3, since the respective change rates of the pump control pilot pressure PP, the pump delivery pressure DP, and the pump tilting SR are constant, the arithmetic operation processes are executed as follows. The input speeds computed in the subtracters702a-702fshown inFIG. 7are provided as the same values as those in the previous cycle. Responsively, the load increase amount ΔT1is computed as the same value, and further the engine revolution speed increase amount ΔT2is computed as the same value. Therefore, the subtracter70kcomputes the determination value α=0, and the engine revolution speed increment value selector70iholds the state B so that the engine revolution speed increase amount ΔT2computed by the multiplier70his introduced as the increment value ΔT2A to the subtracter70m.

Thus, in the second and subsequent cycles of the arithmetic operation process, the previous value of the revolution speed modification value ΔT3is equal to the increment value ΔT2A computed in the current cycle. Accordingly, the subtracter70mcomputes the deviation ΔΔT2=0, and the gain multiplier70nalso computes the deviation ΔΔT4=0, whereby the previous value of the revolution speed modification value ΔT3is maintained. As a result, during the period from the time t2to t3, the target revolution speed NR2is maintained at the increased value as shown inFIG. 10.

From time t3to t4:

When the lever manipulation is stopped at the time t3, the pump control pilot pressure PP, the pump delivery pressure DP, and the pump tilting SR are held constant. Therefore, the input speeds computed in the subtracters702a-702fshown inFIG. 7are provided as negative values. Responsively, the load increase amount ΔT1is computed as a negative value, and further the engine revolution speed increase amount ΔT2is computed as a negative value. Therefore, the subtracter70kcomputes a negative determination value α, and the engine revolution speed increment value selector70iholds the previous value for a certain time (e.g., 3 seconds). Thus, during the holding period of the selector70i, the previous value of the revolution speed modification value ΔT3is maintained as in the above-described period from t2to t3. As a result, for the certain time after t3, the target revolution speed NR2is maintained at the increased value as shown inFIG. 10.

From time t4to t5:

When reaching the time t4after the lapse of the certain time, the engine revolution speed increment value selector70iswitches over from the state B to A, whereupon the increment value ΔT2A is set to 0. Therefore, the subtracter70mcomputes the previous negative value of the revolution speed modification value ΔT3as the deviation ΔΔT2, and the gain multiplier70noutputs the deviation ΔΔT4(<0) as a value resulting from multiplying the deviation ΔΔT2by the gain smaller than 1. The deviation ΔΔT4is applied to the integral adder70p. Accordingly, the revolution speed modification value ΔT3computed by the integral adder70pis smaller than the previous value, and the target revolution speed NR2is also smaller than the previous value. Thus, as shown inFIG. 10, the target revolution speed NR2decreases gradually after the time t4.

After time t5:

When the revolution speed modification value ΔT3reaches 0 (ΔT3=0) at the time t5, the deviation ΔΔT2computed by the subtracter70malso becomes 0, and the revolution speed modification ΔT3is maintained at 0. As a result, the target revolution speed NR2is returned to NR1after the time t5.

With this embodiment, as described above, the engine control system includes status variable detecting means, i.e., the pressure sensors73,74, the position sensors75,76and the pressure sensors77,78, for detecting status variables related to the loads of the hydraulic pumps1,2, and target revolution speed modifying means made up of the target revolution speed modification value computing unit90and the modification value adder70r. The target revolution speed NR2for use in the control is computed based on changes of the status variables such that the target revolution speed NR2for use in the control increases from the target revolution speed NR1applied from the input unit71, and then moderately returns to the target revolution speed NR1. In accordance with the thus-computed target revolution speed NR2for use in the control, the target fuel injection amount FN1is computed and the fuel injection amount is controlled. Therefore, when the engine load is abruptly increased, it is possible to not only suppress a drop of the engine revolution speed, but also to keep the engine revolution speed from going up beyond a required level and to prevent lowering of durability caused by an excessive increase of the engine revolution speed.

Also, the above control process is performed on the basis of engine revolution speed without reducing the absorption torques of the hydraulic pumps1,2. Therefore, the hydraulic pumps1,2can maintain the same maximum delivery rate as that obtained in the case not performing the above-described control, and the work efficiency is not sacrificed.

Further, the control process is performed by computing the target revolution speed NR2for use in the control based on changes of the status variables such that the target revolution speed NR2increases from the target revolution speed NR1applied from the input unit71, is maintained at the increased engine revolution speed for a certain time after detection of the changes of the status variables has ceased, and then moderately returns to the target revolution speed NR1. Therefore, a drop of the engine revolution speed attributable to an abrupt increase of the engine load can be suppressed with certainty.

Moreover, the engine revolution speed increase gain computing unit70gis provided to compute the revolution speed modification value ΔT3, i.e., the increase amount of the target revolution speed, as a variable value depending on the target revolution speed NR1which is set in accordance with a command applied from the input unit71. Therefore, as the target revolution speed NR1set in accordance with the command applied from the input unit71changes, the increase amount of the target revolution speed (i.e., the revolution speed modification value ΔT3) is also changed correspondingly. Hence, an optimum increase amount of the target revolution speed (i.e., the revolution speed modification value ΔT3) can be computed regardless of the target revolution speed NR1, and the control for suppressing a drop of the engine revolution speed can be performed in an appropriate manner without causing an excessive increase of the engine revolution speed.

In addition, since the control pilot pressures PP1, PP2(lever input amounts), the pump tiltings SR1, SR2and the pump delivery pressures DP1, DP2are detected and used in the control as the status variables related to the loads of the hydraulic pumps1,2, the load states of the hydraulic pumps1,2can be confirmed with high accuracy. From this point of view, too, the control for suppressing a drop of the engine revolution speed can be performed in an appropriate manner.

INDUSTRIAL APPLICABILITY

According to the present invention, it is possible to suppress a drop of the engine revolution speed attributable to an abrupt increase of the engine load without sacrificing the work efficiency, and to prevent lowering of durability caused by an excessive increase of the engine revolution speed.