DRIVING CONTROL APPARATUS FOR HYDRAULIC CONSTRUCTION MACHINES

Abstract: The driving control is provided with a motor (1) and at least one hydraulic actuator (3) driven by the oil discharged from the hydraulic pump. A lever (5) moves a linkage (7) for setting the number of revolutions per minute of the motor, which in turn operates an operator (8) to control the hydraulic actuator. The device is further provided with a controller unit (20) connected to the operator and adapted to output a revolution number control signal, by which the set number of revolutions of the motor is increased when the displacement of the operator has exceeded a predetermined level. A revolution number control unit is connected to the controlled which is adapted to make the number of revolutions set by the lever not exceed a predetermined level, and to set the control at the required number of revolutions.

Description: TECHNICAL FIELD The present invention relates to a drive control system for hydraulic construction machines as referred to in the generic part of the main claim. BACKGROUND ART In general, a conventional drive control system for a hydraulic construction machine comprises a prime mover, a hydraulic pump driven by the prime mover, a hydraulic actuator driven by discharge hydraulic fluid from the hydraulic pump, rotational speed setting means including a fuel lever for setting rotational speed of the prime mover, and an operating lever for controlling operation of the hydraulic actuator. Connected between the hydraulic pump and the hydraulic actuator is a control valve for controlling flow rate and direction of the discharge hydraulic fluid from the hydraulic pump. Operation of the operating lever controls the position of the control valve to control operation of the hydraulic actuator. In the above conventional system, rotational speed of the prime mover or an engine is set by displacement of the fuel lever, to vary a horsepower characteristic of the engine in accordance with the set rotational speed. The maximum horsepower of the engine is determined on the basis of the horsepower characteristic. Specific fuel consumption (g/PS · h) of the engine is determined depending upon the set rotational speed and the magnitude of an operational load at that time. If, for example, the rotational speed is set to the maximum value, the specific fuel consumption is brought to the best value, at the heavy load operation in the vicinity of the maximum horsepower of the horsepowwer characteristic obtained by the set rotational speed. On the other hand, at the light load operation which requires only horsepower lower than the maximum horsepower, the engine rotational speed increases to a value higher than the rotational speed at the maximum horsepower of the horsepower characteristic, so that the specific fuel consumption is deteriorated. In general, at the actual operation of the hydraulic excavator, for example, the proportion of the operation which is carried out under the load excellent in the specific fuel consumption is extremely low. For example, in one operational cycle including 1 ○ excavating, 2 ○ boom raising-swing, 3 ○ dumping and 4 ○ boom lowering swing which are repeated in the mentioned order, the operations necessitating the above-mentioned maximum horsepower are only the relief excavating during the operation 1 ○ and the operation at acceleration at the initial stage of the swing during the operation 2 ○. Thus, it is not preferable from the energy-saving point of view to set the rotational speed to the maximum value. In the drive control system of tie kind referred to above, Japanese Patent Application Laid-Open No. 52-53189 has proposed an arrangement in which not only is the rotational speed of the engine set by the fuel lever, but also the engine rotational speed is interlocked with the operating lever for controlling operation of the hydraulic actuator in such a manner that when the operating lever is operated, the engine rotational speed is set also by displacement of the operating lever, to vary the horsepower characteristic, thereby controlling the maximum horsepower. With this arrangement, when displacement of the operating lever is small, the engine rotational speed is set to a low value to give the maximum horsepower required for the light load operation, while as the displacement increases, the engine rotational speed is set to a high value to raise also the maximum horsepower of the engine so as to give the maximum horsepower required for the heavy load operation. Thus, the operation is carried out always in a region excellent in the specific fuel consumption, thereby preventing deterioration of the specific fuel consumption. Further, in a similar drive control system, Japanese Patent Application Laid-Open No. 58-204940 has proposed an arrangement in which only a specific operating lever is interlocked wits the engine rotational speed, and only when the operating lever is operated, the engine rotational speed is set by displacement of the operating lever to vary the horsepower characteristic, thereby controlling the maximum horsepower. In this system, low rotational speed providing the maximum horsepower required for the light load operation is set by the fuel lever and, usually, the operation is carried out with the horsepower characteristic obtained at the low set rotational speed. When the specific operating lever is operated, the rotational speed is set, in interlocked relation to the operation of the operation lever, to a value higher than that set by the fuel lever, so as to give the maximum horsepower required for the heavy load operation with the horsepower characteristic obtained at the set rotational speed, like the above-described conventional system. Thus, the operation is carried out always in a region excellent in the specific fuel consumption, thereby preventing deterioration of the specific fuel consumption. Furthermore, in the drive control system described above, Japanese Patent Application No. 59-129957 has proposed an arrangement comprising, in place of the control valve, a hydraulic pump of variable capacity type and means for varying an angular position of a swash plate of the hydraulic pump, that is, a displacement volume of the hydraulic pump by the operating lever, wherein the engine rotational speed is controlled only by the operating lever, the engine rotational speed is set to a low value when displacement of the operating lever is equal to or lower than a predetermined value, and as the displacement of the operating lever exceeds the predetermined value, the rotational speed is set to a high value in dependence upon the displacement of the operating lever, also in this system, like the above-mentioned conventional system, an attempt can be made to improve the specific fuel consumption, because, in the displacement of the operating lever equal to or larger than the predetermined value, the engine rotational speed is set on the basis of the displacement of the operating lever. In addition to the above-mentioned patents, Japanese Patent Application Laid-Open Nos. 48-53162 and 50-15980, and Japanese Patent Publication No. 60-38561 are listed as being relevant to the arrangement in which the engine rotational speed is interlocked with operation of the operating lever. Moreover, U.S. Patent Serial No. 947,524 (corresponding to EP-A-228787) is listed, which discloses an arrangement in which the engine rotational speed is controlled in response to operation modes or the actuator load. In the system disclosed in Japanese Patent Application Laid-Open 58-204940, output power of an engine is increased only when necessary, by increasing a speed of the engine at operative time of a turning control lever and a boom control lever in the captioned device used for an oil hydraulic work machine of hydraulic shovels or the like. A signal generated in accordance with operation of a throttle lever 6, signal given from a maximum output speed setting means 7, signal given from a turning control level operating means 11, and a signal given from a boom control lever operation detecting means 12 are applied to a fuel injection pump controller 5. When a turning lever 8 is urged further a boom control lever 10 is urged in the direction of lifting a boom, speed of an engine is increased. In this way, the speed of the engine is increased only when a large engine output is requested. However, this system suffers from the following problem. That is, when an operating lever other than the specific operating lever is operated, the rotational speed set by the fuel lever is low and, therefore, it is impossible to carry out the operation necessitating the output power equal to or higher than the maximum horse- power obtained with the horepower characteristic at the set rotational speed. Thus a bad influence is exerted upon the operability. Specifically, for example, in the above-mentioned operational cycle, when an operating lever carrying out the boom raising-swing operation 2 is selected as the specific operating lever the requisite maximum horsepower cannot be obtained at the relief excavating of the operation 1 ○. In other words, it is impossible for the operating lever other than the specific operating lever to effectively utilize the maximum horsepower which tie engine has. Furthermore, in the system disclosed in Japanese Patent Application No. 59-129957, the arrangement is such that the engine rotational speed is set to a constant low value in a region of operation of the operating lever equal to or less than the predetermined displacement. Since, however, the constant value is fixedly determined, the operating lever must be operated with displacement equal to or larger than the predetermined value to set the engine rotational speed to a higher value, at the operation necessitating the maximum horsepower higher than that obtained with the horsepower characteristic of the set low rotational speed. Also in this case, the engine rotational speed frequently fluctuates, giving rise to problems such as deterioration of the specific fuel consumption, smoke emission and generation of noises. For example, in the above-mentioned operational cycle, when the constant rotational speed is set to a low value suitable for the operation 4 ○ lowest in the requisite horsepower, operation of the operating lever causes the engine rotational speed to frequently fluctuate at the operations other than the operation 4 ○. This raises problems such as deterioration of the specific fuel consumption due to acceleration of the flywheel, smoke emission and noises. In addition, when the constant rotational speed is set to a high value, the engine rotational speed is brought to a high value inferior in the specific fuel consumption from the horsepower characteristic point of view, at the operation which necessitates only the horsepower lower than the maximum horsepower obtained with the horsepower characteristic of the constant rotational speed, thereby making it impossible to achieve the original object. That is, in the above-mentioned operational cycle, when the constant rotational speed is set to an intermediate value suitable for the usual excavating operation of 1 ○ and the swing operation subsequent to the initial acceleration of 2 ○, the specific fuel consumption is deteriorated at the operations 3 ○ and 4 ○ which are low in requisite horsepower. Furthermore, since the constant rotational speed is determined in a fixed fashion, even if an operator desires operation in which noises and smoke emission due to fluctuation of the rotational speed are not caused, it is impossible to carry out such desired operation. Thus, there has been a problem regarding the operability. DISCLOSURE OF THE INVENTION It is therefore an object of the invention to provide a drive control system for a hydraulic construction machine, which can improve the specific fuel consumption and can reduce fluctuation in rotational speed of an engine, and which is superior in operability. This object is solved in accordance with the main claim, dependent claims are directed on preferred embodiments of the present invention. BRIEF DESCRIPTION OF THE DRAWINGS Fig. 1 is a diagrammatic view showing the entirety of a drive control system for a hydraulic construction machine according to a first embodiment of the invention; Fig. 2 is a detailed view of an operating device in the drive control system; Fig. 3 is a detailed view of a rotational speed control device in the drive control system; Fig. 4 is a flow chart for explanation of operation of a controller in the drive control system; Figs. 5 and 6 are characteristic graphs showing the relationship between displacement of an operating lever and set rotational speed of an engine in the drive control system; Fig. 7 is a graphical representation of a requisite engine output power in one operational cyclic, for explanation of the operation of the drive control system; Fig. 8 is a graphical representation of characteristics of respective output horsepower, torque and specific fuel consumption when the rotational speed of the engine is varied; Fig. 9 is a diagrammatic view showing a modification of the rotational speed control device; Figs. 10 and 11 are characteristic graphs showing the relationship between displacement of the operating lever and the set rotational speed of the engine when the rotational speed control device illustrated in Fig. 9 is employed; Figs. 12(a), 12(b) and 12(c) are diagrammatic views respectively showing operating positions different from each other, in another modification of the rotational speed control device; Fig. 13 is a characteristic graph showing the relationship between displacement of the operating lever and the set rotational speed of the engine when the rotational speed control device illustrated in Figs. 12(a) through 12(c) is employed; Fig. 14 is a characteristic view when the rotational speed control device is further modified; Fig. 15 is a diagrammatic view showing the entirety of a drive control system according to another embodiment of the invention; Fig. 16 is a graphical representation of the relationship between displacement of the operating lever and a stroke amount of a control valve in an embodiment in which the displacement and the stroke amount are set especially; Fig. 17 is a graphical representation of the relationship between displacement of the operating lever and flow rate of fluid passing through the control valve in the embodiment illustrated in Fig. 16; Fig. 18 is a diagrammatic view showing the entirety of a drive control system according to still another embodiment of the invention; Fig. 19 is a flow chart for explanation of operation of a controller in the drive control system illustrated in Fig. 18; Fig. 20 is a graphical representation of the relationship between displacement of an operating lever and flow rate of fluid passing through a control valve in the drive control system illustrated in Fig. 18; Fig. 21 is a graphical representation of the relationship between engine rotational speed and a pump discharge quantity in the drive control system illustrated in Fig. 18; Fig. 22 is a graphical representation of the relationship between pump discharge pressure and the pump discharge quantity in the drive control system illustrated in Fig. 18; Fig. 23 is a diagrammatic view showing the entirety of a drive control system when the embodiment illustrated in Fig. 18 is arranged electronically; BEST MODE FOR CARRYING OUT THE INVENTION Preferred embodiments of tie invention will be described below with reference to the drawings. Fig. 1 shows a drive control system for a hydraulic construction machine, according to a first embodiment of the invention. The drive control system comprises a prime mover or an engine 1, a hydraulic pump 2 driven by the engine 1, and a hydraulic actuator 3 driven by discharge hydraulic fluid from the hydraulic pump 2. A control valve 4 is connected between the hydraulic pump 2 and the hydraulic actuator 3, for controlling flow rate and direction of the hydraulic fluid supplied from the hydraulic pump 2 to the hydraulic actuator 3. The system comprises a plurality of actuators 3 and control valve 4. The prime mover 1 is preferably a diesel engine which comprises a fuel injection system provided with an all-speed governor. In order to set rotational speed of this engine, a first rotational speed setting device 7 is provided, which is composed of a first operating device or a fuel lever 5, and a governor lever 6 operatively connected to the fuel lever 5. In this first rotational speed setting device 7, as the fuel lever 5 is operated in a direction A, the governor lever is operated in a direction B in response to the operation of tie fuel lever 5, so that the rotational speed is set to a value in compliance with displacement of the fuel lever 5. The operation of the hydraulic actuators 3 is controlled by a respective second operating device 8. As shown in Fig. 2, the second operating device 8 comprises an operating lever 9 and two hydraulic pilot valves 10 and 11. The hydraulic pilot valves 10 and 11 have their respective primary ports which are connected to a pilot pump 12 driven by the engine 1 and to a reservoir 13. Second ports of the respective hydraulic pilot valves 10 and 11 are connected respectively to pilot ports of the control valve 4 through respective pilot lines 14 and 15. The arrangement is such that the pilot valves 10 and 11 are supplied with primary pressure from the pilot pump 12, and secondary pressures in accordance with displacements of the respective pilot valves 10 and 11 are supplied respectively to the pilot ports of the control valve 4. In response to receipt of the secondary pressures, the control valve 4 is controlled in position, that is, in stroke amount and direction, thereby controlling the flow rate and the direction of the hydraulic fluid supplied to the hydraulic actuator 3 to control the operation of the same. The second operating device 8 is also provided with springs 16 and 17 which serve to increase lever operating force when displacement of the operating lever 9, that is, an operating amount thereof exceeds a predetermined value X₀. By the springs, as the operating amount of the lever 9 is brought to a value equal to or higher than X₀, the operating force becomes heavy to thereby inform an operator of the position of the operating lever 9. The second operating devices 8 are respectively connected to a second rotational speed setting device 20 which outputs a rotational speed control signal increasing the set rotational speed of the engine 1 as the sum of displacements of the operating devices exceeds the predetermined value X₀. A rotational speed control device 21 is associated with the second rotational speed setting device 20. The second rotational speed setting device 20 is composed of a plurality of pressure sensors 23 connected to the pilot lines 14 and 15 of the respective actuator through a shuttle valve 22, for detecting the maximum pressure, and a controller 24 formed by a microcomputer or the like. A detecting signal from the sensors 23 is inputted into the controller 24, and the controller 24 executes a predetermined operation processing to obtain the above-mentioned rotational control signal and outputs the same. The controller has inputted beforehand therein a control program as shown by a flow chart in Fig. 4, inclusive of the above predetermined value X₀. As shown in Fig. 3, the rotational speed control device 21 is comprised, for example, of a linear solenoid cylinder 25 which is adapted to extend a piston 26 in accordance with a level of tie rotational control signal from the controller 24, to operate the governor lever 6 in the direction B. The operation of the embodiment will next be described with reference to the flow chart shown in Fig. 4. When the program starts, the detecting signals from the pressure sensors 23 are read into the controller 24 at a step S₁. At a step S₂, it is judged by the controller 24 wheter or not the sum of displacements of the operating levers 9 indicated by the detecting signals exceeds the above-mentioned predetermined value X₀ which is set beforehand. If it is not judged that the sum of displacement exceeds the predetermined value X₀, the program skips a step S₃ and is returned to the start (step S₁). Accordingly, the rotational speed control signal is not outputted from the controller 24, and the linear solenoid cylinder 25 shown in Fig. 3 is not actuated. Thus, the governor lever 6 is operated only of the fuel lever 5, so that the rotational speed set by the fuel lever 5 is validated. On the other hand, if it is judged at the step S₂ that the sum of displacements of the operating levers 9 exceeds the predetermined value X₀, the program proceeds to the step S₃ where the rotational speed control signal of a level corresponding to the read-in detecting signal is outputted. This rotational speed control signal is sent to the linear solenoid cylinder 25, to proportionally control the stroke amount of the piston 26. Thus, the governor lever 6 is operated by the linear solenoid cylinder 25, so that the rotational speed set by the controller 24 is set. With the arrangement described above, when the rotational speed of the engine 1 is set to an idling N i by the fuel lever 5, as shown in Fig, 5, the set rotational speed at the idling N i is maintained until the sum of displacements of the operating levers 9 reaches the predetermined value X₀. As the sum of displacements exceeds X₀, the set rotational speed of the engine increases in proportion to displacement of the operating levers 9, and reaches the maximum value N max at the maximum displacement X max . When the engine rotational speed is set to an intermediate value N₁ by the fuel lever 5, as shown in Fig. 6, the set rotational speed being to increase as the sum of displacements of the operating levers 9 exceeds a value X₁ at which a set rotational speed N₁ is obtained. In this manner, the rotational speed control device 21 is so arranged as to adjust the rotational speed set by the first rotational speed setting device 7 in a first region Z₁ in which the sum of displacements of the second operating devices 8 is at least equal to or less than the above predetermined value X₀, that is, is equal to or less than the predetermined value X₀ or the sun of displacements X₁ larger than the predetermined value. In addition, the rotational speed control device 21 sets a rotational speed higher than the rotational speed set by the first rotational speed setting device 7, modified by the rotational speed control signal from the second rotational speed setting device 20 in a second region Z₂ in which the sum of displacements is larger than the value X₀ or X₁. Particularly in this embodiment, the rotational speed control device 21 is so arranged as to adjust the set rotational speed indicated by the rotational speed control signal from the second rotational speed setting device 21 in the second region Z₂. Advantageous effects of the drive control system constructed as above will next be described. Fig. 7 is a graphical representation of one operational cycle which is a typical example of the operation conducted by the hydraulic excavator, in which 1 ○ excavating, 2 ○ boom raising-swing, 3 ○ dumping and 4 ○ boom lowering-swing are repeated in the mentioned order. Fig. 7 shows the one operational cycle in relation to the engine output power required for each operation. In the figure, N A is a set rotational speed of the engine suitable for giving the output power required for the light load operation, N B is a set rotational speed suitable for giving the output power required for the usual heavy load operation, and N C is a set rotational speed suitable for giving the output power required for especially heavy load operation. Further, Fig. 8 shows an output horsepower characteristic, a torque characteristic and a specific fuel consumption when the engine rotational speed is set to a selected one of the above values N A , N B and N C . In the one operational cycle shown in Fig. 7, when the engine rotational speed is set to a constant value of the highest one N C , the specific fuel consumption is brought to a value g 1c and is excellent, at the relief excavation of the operation 1 ○ and at acceleration at the initial stage of the swing in the operation 2 ○, as shown in the operation 2 ○, as shown in Fig. 8. At other operations, for example, at usual swing in the operation 2 ○, the specific fuel consumption is brought to a value g 2c , and at boom lowering swing in the operation 4 ○, the specific fuel consumption is brought to a value g 3c . Thus, the specific fuel consumption is deteriorated. Accordingly, if the rotational speed is set by the fuel lever to the value N A suitable for the operation 4 ○, and if the engine rotational speed is set to appropriate values in dependence upon the respective operations in interlocked relation to the operating lever, the specific fuel consumption is raised, for example, to g₂ b and g 3g . In this case, however, the engine rotational speed frequently fluctuates in interlocked relation to operation of the operating lever, during almost all of the period of time of the operations other than the boom Iowering-swing operation, so that energy is consumed to accelerate the flywheel of the engine. This is not preferable in the specific fuel consumption. There also exist problems of smoke emission and noises due to fluctuation of the engine rotational speed. In the drive control system of the embodiment, the rotational speed is set to a value of a desirable level in dependence upon displacement of the fuel level 5 in the first region Z₁. By doing so, in the above-mentioned operational example, the engine rotational speed is set to the value N B by the fuel lever 5, whereby the specific fuel consumption in the vicinity of g 2b , is obtained at the usual excavating of 1 ○ and at the usual swing of 2 ○, and the specific fuel consumption in the vicinity of g 3b more excellent than g 3c is obtained at the dumping of 3 ○ and at the boom lowering-swing of 4 ○. On the other hand, since, in the second region Z₂, the rotational speed is set to the higher value by the operating levers 9, the rotational speed of the engine is set by operation of the operating levers 9 at the relief excavating of 1 ○ and at acceleration at the initial stage of the swing of 2 ○, to obtain a higher set rotational speed, whereby the specific fuel consumption of g 1c is obtained. In this manner, it is possible to obtain excellent specific fuel consumption as a whole. Since, in the first region Z₁, setting of the rotational speed by means of the operating levers 9 is not carried out, the rotational speed does not fluctuate even if the operating levers 9 are operated. Thus, fluctuation of the engine rotational speed is reduced as a whole, so that energy consumption due to acceleration of the flywheel can be ignored, and the problems of smoke emission and noises due to fluctuation of the engine rotational speed are diminished. Further, if an operator desires the operation in which noises and smoke emission due to fluctuation of the engine rotational speed are completely eliminated, setting of the engine rotational speed by the fuel lever 5 is brought to the maximum value N C , whereby the aforesaid operation can be realized. Thus, the operability is improved. Additionally, in practice, the above-mentioned predetermined value X₀ is determined in consideration of the following points. The first point is as follows. When the engine rotational speed is set by the fuel lever 5 to a value in the vicinity of the idling N i which is employed in the lightest load operation such as normal plane operation or the like, the discharge quantity of the hydraulic pump 2 is determined by the rotational speed. On the other hand, as the operating levers 9 are operated, the control valve 4 begins to be opened in dependence upon displacement of tie operating lever, and the requisite flow rate required by the control valve and the flow rate of fluid passing through the control valve flowing at the discharge quantity of the hydraulic pump are brought into coincidence with each other at a certain specific opening degree of the control valve. Thus, the first point is to bring the sum of displacements of the operating lever 9 indicating the specific opening degree to X₀. That is, the sun of displacements of the operating levers 9 is brought to a value obtaining the opening degree of the control valves 4 at which the flow rate of fluid passing through the control valves 4 obtained by restricting the discharge quantity of the hydraulic pump 2 is brought into coincidence with the requisite flow rate. By doing so, it is mace possible to secure the first and second regions Z₁ and Z₂ shown in Fig. 5 and 6, at substantially the entire set rotational speed. Thus, the engine rotational speed can be set in interlocked relation to the operating levers in a region equal to or higher than the predetermined value X₀ or X₁. The second point is the sum of displacements of the operating levers 9 which obtains a valve opening degree corresponding to an upper limit of a metering region of the control valve 4 required for the fine or minute operation working. By the displacements, it is made possible to secure the metering characteristic as designed, which is not influenced by rise in the engine rotational speed in a region equal to or lower than the predetermined value X₀. Thus, the desirable fine operation working can be carried out. As another point, there is sum of displacements of the operating levers 9 giving the predetermined value X₀ at which the problems of smoke emission and noises in the region equal to or higher than the predetermined value X₁ are minimized, in consideration of every operational contents. In the above-described embodiment, the sum of displacements of the operating levers 9 and the set rotational speed of the engine are brought to the linear proportional relationship in the second region Z₂ as shown in Figs. 5 and 6, but are not limited only to this relationship. For example, the arrangement may be such that the opening degree of the control valve 4 is calculated on the basis of the sum of displacements of the operating levers 9, and an engine rotational speed control signal is outputted which can obtain discharge quantity of the hydraulic pump 2 corresponding to the requisite flow rate prescribed by the opening degree. In this case, the engine rotational speed increases in predetermined functional relation other than the linear proportion to the displacements of the operating levers. In the above embodiment, it has been described that the rotational speed control signal outputted from the controller 24 is increased proportionally in dependence upon sum of displacements of the operating levers 9, and the linear solenoid cylinder 25 is employed which is operated with a strove amount on the oasis of the signal. As a result, as shown in Figs. 5 and 6, the predetermined value forming the boundary between tee first and second regions Z₁ and Z₂ is changed from X₀ to X₁ in dependence upon the rotational speed set by the fuel lever 5, and the set rotational speed increases in response to sum of displacements of the operating levers 9 in the second region Z₂. However, an arrangement different from the above embodiment in this respect can be employed. That is, the rotational speed control signal outputted from the controller 24 at a sum equal to or larger than the predetermined value X₀ is set to a constant value, and the rotational speed control device 21 is formed, in place of the linear solenoid cylinder 25, by a solenoid cylinder of ON-OFF type which extends to the maximum stroke when the rotational speed control signal reaches the constant value. Moreover, as shown in Fig. 9, a rotational speed control device 32 may be composed of an electromagnetic directional control valve 30 turned on and off in response to the rotational speed control signal, and a hydraulic cylinder 31 movable between ON and OFF positions in accordance with the position of the directional control valve 30. In this case, the relationship between the sum of displacement of the operating levers 9 and the set rotational speed of the engine is brought to one as shown in Fig. 10 of the rotational speed set by the fuel lever 5 is the idling N i , and to one as shown in Fig. 11 if the rotational speed set by the fuel lever 5 is the intermediate rotational speed N₁. That is, the predetermined value forming the boundary between the first and second regions Z₁ and Z₂ is X₀ and is constant independently of the rotational speed set by the fuel lever 5, while the set rotational speed is brought to the maximum value N max in the second region Z₂ independently of displacement of the operating lever 9. With such arrangement, the number of component parts is reduced and the structure is simplified. The arrangement of each of the above-described embodiments is such that the rotational speed control device 21, 32 adjusts the rotational speed set by the rotational speed control signal obtained by the second rotational speed setting device 20 in the second region Z₂. However, an arrangement different from the above arrangement also in this respect can be employed. Figs. 12(a) through 12(c) shows an embodiment having such arrangement, in which the reference numeral 40 denotes a rotational speed control device. The rotational speed control device 40 is so arranged as to add the rotational speed set by the rotational speed control signal to the set rotational speed obtained by the fuel lever 5 in the above-mentioned second region Z₂. That is, as shown in Fig. 12(a) in which the fuel lever 5 is in the OFF-position, the fuel lever 5 is pivotally supported by a console box 41 within an operator's cab, and is connected, through a push-pull cable 43, to one end of a first intermediate lever 42 which is pivotally supported at a predetermined portion of the vehicle. The first intermediate lever 42 has the other end to which a linear solenoid cylinder 44 is fixedly mounted. A second intermediate lever 45 is pivotally supported in coaxial relation to the first intermediate lever 42. Pivotal movement of the first intermediate lever 42 is transmitted to the second intermediate lever 45 through the linear solenoid cylinder 44. The second intermediate lever 45 is connected to the governor lever 6 through a push-pull cable 46. The rotational speed control signal is supplied from the controller 24 of the second control setting means 20 to the linear solenoid cylinder 44, so that a stroke amount corresponding to the magnitude of the signal is obtained at the linear solenoid cylinder 44. An idling position is a position where pivotal movement of the fuel lever 5 in the direction A causes the forward end of the linear solenoid cylinder 25 to be brought into engagement with the second intermediate lever 45. In this case, as indicated by the line 1₁ in Fig. 13, the set rotational speed of the engine 1 is a constant value N i in the first region Z₁ of from zero to the predetermined value X₀ of the sum of displacements of the operating levers 9. As the sum of displacements of the operating levers 9 exceeds the predetermined value X₀, a rotational speed control value increasing in proportion to the displacement is obtained at the second rotational speed setting device 20. A rotational speed control signal corresponding to the rotational speed control valve is sent to the linear solenoid cylinder 44, so that the linear solenoid cylinder 44 extends with a stroke in dependence upon the rotational speed control signal. Thus, the set rotational speed increases in the second region Z₂ as indicated by the line 1₁ in Fig. 13. Also as shown in Fig. 12(b), in case where the engine rotational speed is set by the fuel lever 5 to the interemdiate value N₁, as the sum increases to the maximum value X max beyond the predetermined value X₀, the linear solenoid cylinder 44 extends to the maximum strove amount as shown in Fig. 12(c). Thus, the set rotational speed increases as indicated by the line 1₂ in Fig. 13. Additionally, like the embodiment described with reference to Figs. 9 and 10, the linear solenoid cylinder 44 may be formed by an actuator operable between ON and OFF positions. In this case, the relationship of the set rotational speed of the engine with respect to the sum of displacement of the operating levers 9 is brought to one shown in Fig. 14. The above-described embodiments are directed to an example which employs the controller 24 generating the rotational speed control signal to the second rotational speed setting device 20. However, an arrangement different from the embodiments in this respect may be employed. Fig. 15 shows an embodiment having such arrangement. In the figure, component parts similar to those illustrated in Fig. 1 are designated by the same reference numerals. In this embodiment, a second rotational speed setting device 60 comprises a directional control valve 61 which is controlled in switching as the secondary pressure of the operating device 8 formed by a pilot valve exceeds a predetermined value corresponding to the predetermined value X₀ of the sum of displacement of the operating levers 9, and a rotational speed control device 62 comprises a proportional control hydraulic cylinder 63 which is extended and retracted directly by the secondary pressure of the operating device 8 transmitted through the directional control valve 61. That is, when the secondary pressure of the operating device 8 is equal to or lower than the predetermined value, the directional control valve 61 is in the illustrated position where transmission of the secondary pressure is blocked. As the secondary pressure exceeds the predetermined value, the directional control valve 61 is switched to the other position where the secondary pressure is applied to the hydraulic cylinder 63 as a rotational speed control signal, so that the hydraulic cylinder 63 is extended with a stroke amount corresponding to the pressure. Also in this embodiment, the relationship between the sum of displacements of the operating levers 9 and the set rotational speed of the engine 1 is brought to one shown in Figs. 5 and 6 in dependence upon the rotational speed set by the fuel lever 5. If the hydraulic cylinder 63 is controlled in an ON and OFF manner, the relationship between the aforesaid displacement and the set rotational speed is brought to one shown in Figs. 10 and 11. If the arrangement illustrated in Figs. 12(a) through 12(c) is employed so as to add the value set by the second rotational speed setting device 60, the relationship between displacement of the operating lever and the set rotational speed is brought to one shown in Fig. 13. If control is carried out in an ON and OFF manner, the relationship is brought to one shown in Fig. 14. In the above embodiments, it has been described that one of control of the engine rotational speed by operation of the fuel lever 5 and the operating lever 9 is carried out mechanically by the first rotational speed setting device 7, and the other is carried out electronically or hydraulically by the second rotational speed setting means 20, 60 in a manner separate from the one control. However, an arrangement can be employed in which these controls are put together into a single electronic control system and are carried out thereby. In the above embodiments, it has been described that, the relationship between the sum of displacement of the operating levers 9 and the stroke amount determining the opening degree of the control valve 4 is set such that, the stroke amount of the control valve is so set as not to reach the maximum value when the displacement of the operating lever reaches the predetermined value X₀, but as to be brought to an intermediate stroke amount. This has been referred to above at the second point as the points to be taken into consideration when the predetermined value X₀ is set. As shown in Fig. 16, however, setting can be made such that the stroke amount of the control valve 4 is brought to the maximum value (maximum in opening degree), when the operating levers 9 together are operated up to the predetermined value X₀. If setting is made in this manner, the relationship between the sum of displacement of the operating levers 9 and flow rate of fluid passing through the control valve 4 is brought to one shown in Fig. 17. Thus, in a range within which the sum is less than the predetermined value X₀, the engine rotational speed does not fluctuate due to operation of the operating lever. Accordingly, it is possible to obtain the requisite flow rate in accordance with the stroke amount (opening degree) of the control valve over the entire stroke amount thereof, so that desirable actuator speed can be obtained even at the light load operation. In the above embodiments, it has been described that when the sum exceeds the predetermined value X₀, only the engine rotational speed is controlled on the basis of the displacement. However, an arrangement may be employed in which a displacement volume of the hydraulic pump 2 is also controlled. Fig. 18 shows an embodiment having such arrangement. In the figure, component parts similar to those illustrated in Fig. 1 are designated by the same reference numerals. The arrangement of this embodiment is such that a hydraulic pump 90 of variable capacity type is employed, and a displacement volume of the hydraulic pump is varied by a displacement volume control device 91 which regulates an angle of inclination of a swash plate of the hydraulic pump. Like the embodiment shown in Fig. 1, a controller 92 forms a second rotational speed setting device 93 for outputting a rotational speed control signal to the rotational speed control device 21. The controller 92 also outputs a displacement volume control signal to the displacement volume control device 91. With such arrangement, as the sum of displacements of the operating levers 9 exceeds the predetermined value X₀, the displacement volume (angle of inclination) of the hydraulic pump 90 is reduced correspondingly to an increase in the rotational speed of the engine 1. That is, the controller 92 has stored therein a control program as indicated by a flow chart in Fig. 19. At a step S₁, a detecting signal from the pressure sensor 23 is read into the controller 92. It is judged at a step S₂ whether or not the sum indicated by the detecting signals exceeds the predetermined value X₀. If it is judged that the sum exceeds the predetermined value, the rotational speed control signal increasing the set rotational speed of the engine 1 in proportion to the displacement is outputted at a step S₃. At the same time, the displacement volume control signal reducing the displacement volume (angle of inclination) in dependence upon the increase in the set rotational speed is outputted to the displacement volume control device 91. At this time, preferably, the displacement volume control signal is so determined as to reduce the displacement volume such that the discharge quantity of the hydraulic pump is brought substantially to a constant value with respect to the increase in the engine rotational speed. Further, like the embodiment described with reference to Figs. 21 and 22, the arrangement of the present embodiment is such that the stroke amount of the control valve 4 is brought to the maximum value, that is, the valve opening degree is brought to the maximum value, at the predetermined value X₀ of the sum of operating lever displacement. By doing so, the relationship between the sum of displacement of the operating lever 9 and the flow rate of fluid passing through the control valve 4 is brought to one shown in Fig. 20. That is, since no fluctuation exists in the set rotational speed of the engine till the predetermined value X₀, it is possible to obtain the passing flow rate in compliance with the requisite flow rate determined by the stroke amount (operating degree) of the control valve 4, over the entire stroke amount thereof. In a range exceeding the predetermined value X₀, the control valve passing flow rate is made constant by the above-mentioned control of the set rotational speed and the displacement volume. As a result, as the set rotational speed of the engine 1 is increased and decreased in response to the requisite load, it is made possible to complementarily increase and decrease the absorption horsepower of the hydraulic pump 2 in accordance with the increase and decrease in the engine rotational speed. Thus, it is made possible to effectively utilize the engine horsepower, while the operating speed is maintained constant. The above will now be described with reference to Figs. 21 and 22. In this embodiment, as shown in Fig. 21, the discharge quantity of the hydraulic pump 90 increases in proportion to an increase in the engine rotational speed until the engine rotational speed reaches a value N₀ corresponding to the predetermined value X₀ of the operating lever displacement, because the displacement volume is constant. On and after the value N₀, the discharge quantity is brought to a constant value Q₀ until the discharge quantity reaches the maximum value N max as described above. The relationship between the pump discharge pressure P and the pump discharge quantity Q at this time is brought to one shown in Fig. 22. That is, the relationship indicates a P-Q characteristic as shown by the dot-and-chain line in the running condition of the rotational speed N₀, and indicates a P-Q characteristic as shown by the solid line in the running condition of the rotational speed N max . In a range of from N₀ to N max of the rotational speed, the P-Q characteristic varies continuously between the dot-and-chain line and the solid line in response to variation in the rotational speed. At this time, the region in which tie pump discharge quantity Q is constant at Q₀ increases from P₀ to P₁ of tie pump pressure, and the absorption horsepower also increase correspondingly. Even on and after the engine rotational speed N₀, if the pump discharge quantity is increased in proportion to the engine rotational speed as usual, the P-Q characteristic is brought to one indicated by the broken line in Fig. 22 at N max of the engine rotational speed. In this manner, in the present embodiment since the pump discharge quantity Q₀ is controlled to a constant value in the range equal to or higher than the engine rotational speed N₀, it is made possible to increase the consumptive horsepower correspondingly to an increase in tie engine rotational speed. Thus, the engine horsepower can effectively be utilized while the operating speed is maintained constant. Further, if the control valve 4 is maintained at the maximum opening degree like the present embodiment, the entire pump discharge quantity can be supplied to the hydraulic actuator 3, whereby it is made possible to more effectively utilize the engine horsepower. Additionally, the displacement volume control device 91 may be composed, for example, of a hydraulic cylinder and a linear solenoid valve proportionally controlled by the signal from the controller 92. Fig. 23 shows an embodiment having such arrangement, in which component parts similar to those shown in Fig. 18 are designated by the same reference numerals. That is, a controller 95 has inputted thereinto a signal from the displacement detector 73 for detecting displacement of the fuel lever 5, and signals from the respective detecting devices 75 and 76 for detecting displacements of the respective operating levers 9 and 72. The controller 75 outputs a commmand signal instructing a final set rotational speed to the pulse motor 77, and outputs a displacement volume control signal to a displacement volume control device 96 formed by a linear solenoid cylinder.

Claims: 1. A drive control system for a hydraulic construction machine, comprising a prime mover (1), a hydraulic pump (2; 90) driven by the prime mover (1), a plurality of hydraulic actuators (3) driven by discharge hydraulic fluid from the hydraulic pump (2, 90), first rotational speed setting means (7; 80) including first operating means (5) for setting rotational speed of the prime mover (1); and a plurality of second operating means (8, 71) for respectively controlling operation of the plurality of hydraulic actuators (3), and further comprising: second rotational speed setting means (102) operating in response to the actuation of the second operating means (8, 9); and rotational speed control means (82) associated with at least the second rotational speed setting means (102), for setting the rotational speed set by the first rotational speed setting means (7; 80) in a first region (Z₁) and for setting a rotational speed higher than the rotational speed set by the first rotational speed setting means (7; 80), modified by the rotational speed control signal from the second rotational speed setting means in a second region (Z₂), said regions (Z₁, Z₂) being distinguished by an operating amount of the second operating means; characterized in that said second rotational speed setting means (102) is so arranged as to provide a signal determining the first region (Z₁) when a sum of displacements of a plurality of respective second operating means (8, 81) is smaller than a predetermined value (X₀; X₀) and so as to provide a signal determining the second region (Z₂) when said sum exceeds said value (X₀, X₀), said value (X₀, X₀) being larger than zero. 2. A drive control system according to claim 1, characterized in that the second rotational speed setting means (20, 24; 60, 61; 81; 85; 92; 93) sets the rotational speed control signal so as to increase proportionally the set rotational speed in said second region (Z₂) in dependence upon the sum of displacements of the second operating weans (8). 3. A drive control system according to claim 1, characterized in that the second rotational speed setting means (20, 24; 84; 87) sets said rotational speed control signal so as to bring the set rotational speed in said signal region (Z₂) to a constant value. 4. A drive control system according to claim 1, characterized in that the rotational speed control means (21; 32; 62; 77, 82) is so arranged as to adjust the rotational speed of the rotational speed control signal in said second region (Z₂). 5. A drive control system according to claim 1, characterized in that the rotational speed control means (40; 77, 86) is so arranged as to add the rotational speed of the rotational speed control signal to the set rotational speed in said second region (Z₂). 6. A drive control system according to claim 1, characterized in that the second rotational speed setting means comprises detecting weans (23) for detecting displacements of the second operating means (8), and control means (20, 24; 60, 61; 92; 93) for obtaining said rotational speed control signal on the basis of the displacement detected by the detecting means, and that the rotational speed control means (21; 32; 40; 62; 77, 82; 86) comprises an actuator (26; 31; 44) driven by an output signal from this control means. 7. A drive control system according to claim 1, characterized in that the first rotational speed setting means comprises means (80) having inputted thereinto the displacement signals from the first operating means (5) for obtaining the set rotational speed on the basis of the sum of displacements that the second rotational speed setting means comprises means (81; 84; 85; 87) having inputted thereinto the displacement signals from the second operating means (8) for obtaining the rotational speed control signal on the basis of the sum of displacements and that the rotational speed control means comprises means (82; 86) for obtaining a final rotational speed on the basis of output signals from the respective first and second rotational speed setting means (20, 24; 60, 61; 81; 84; 85; 87; 92, 93), and an actuator (77) driven by an output signal from this means. 8. A drive control system according to claim 1 wherein said hydraulic pump (90) is of variable capacity type, characterized of further comprising: displacement volume control means (91; 96, 97) for controlling said hydraulic pump (90) so as to decrease a displacement volume of the hydraulic pump (90) when the sum of displacements of the second operating means (9) exceeds said predetermined value (X₀; X₀). 9. A drive control system according to claim 8, characterized in that said displacement volume control means is arranged to reduce the displacement volume so as to bring a discharge quantity of the hydraulic pump (90) to a substantially constant value with respect to increasing rotational speed of the prime mover (1). 10. A drive control system according to claim 1 or claim 8 wherein a control valve (4) is connected between said hydraulic pump (2; 90) and the hydraulic actuator (3) for controlling flow rate and direction of discharge hydraulic fluid from the hydraulic pump (2; 90), the control valve being controlled in position by said second operating means (8) to control operation of the hydraulic actuator (3), characterized in that the control valve (4) is so arranged that when the sum of displacements of the second operating means (8) reaches said predetermined value (X₀; X₀), an opening degree of the control valve is brought to a maximum value. 11. A drive control system according to any one of claims 1 and 8, wherein a control valve (4) is connected between said hydraulic pump (2; 90) and each of the hydraulic actuator or actuators (3), for controlling flow rate and direction of discharge hydraulic fluid from the hydraulic pump (2; 90), the control valve being controlled in position by said second operating means (8) to control operation of the hydraulic actuator or actuators (3), characterized in that the second rotational speed setting means (20, 24; 81; 85; 92, 93) calculates an opening degree of said control valve (4) on the basis of a detecting signal indicative of the sum of displacements of the second operating means (8, 71), and calculates the rotational speed control signal capable of obtaining discharge quantity of the hydraulic pump (2; 90) which corresponds to a requisite flow rate prescribed by the opening degree.

Inventors: TATSUMI, AKIRA
Applicants: HITACHI CONSTRUCTION MACHINERY CO., LTD.
Application year: 1987
Main classification: F02D29
IPC codes: F02D29, E02F9