HYDRAULIC DRIVE APPARATUS

Abstract: Differential pressures across flow control valves (6a, 6b, 6c) are controlled by pressure compensating valves (7a, 7b, 7c) to a differential pressure ΔPLS having the same value, and the differential pressure ΔPLS is maintained at a target differential pressure ΔPLSref by a pump capacity control device (5). To modify the target differential pressure ΔPLSref depending upon a change in revolution speed of an engine (1), a flow rate detecting valve (31) is provided on discharge passages (30a, 30b) in a fixed capacity hydraulic pump (30) so that a differential pressure ΔPp across a variable restriction (31a) in the flow rate detecting valve (31) is conducted to a setting modifying device (32) to modify the target differential pressure ΔPLSref. The flow rate detecting valve (31) acts to change an opening area of the variable restriction (31a) depending upon the differential pressure ΔPp across the variable restriction (31a) itself, and to change the differential pressure ΔPp in accordance with the revolution speed of the engine (1). Accordingly, saturation phenomenon is improved in accordance with the engine revolution speed, and a favorable minute operability is obtained in the case where the engine revolution speed is set low.

Description: TECHNICAL FIELD The present invention relates to a hydraulic drive system including a variable displacement hydraulic pump, and more particularly to a hydraulic drive system operating under load sensing control to control the displacement of the hydraulic pump so that a differential pressure between a delivery pressure of the hydraulic pump and a maximum load pressure among a plurality of actuators is maintained at a setting value. BACKGROUND ART As to the load sensing control technique for controlling the displacement of a hydraulic pump so that a differential pressure between a delivery pressure of the hydraulic pump and a maximum load pressure among a plurality of actuators is maintained at a setting value, there are known a pump displacement control system disclosed in JP, A, 5-99126 and a hydraulic drive system disclosed in JP, A, 60-11706. The pump displacement control system disclosed in JP, A, 5-99126 comprises a servo piston for tilting a swash plate of a variable displacement hydraulic pump, and a tilting control unit for supplying a pump delivery pressure to the servo piston in accordance with a differential pressure ΔPLS between a delivery pressure Ps of the hydraulic pump and a load pressure PLS of an actuator driven by the hydraulic pump so as to maintain the differential pressure ΔPLS at a setting value ΔPLSref, thereby controlling the pump displacement. The disclosed pump displacement control system further comprises a fixed displacement hydraulic pump driven by an engine along with the variable displacement hydraulic pump, a throttle disposed in a delivery line of the fixed displacement hydraulic pump, and setting modifying means for modifying the setting value ΔPLSref of the tilting control unit in accordance with a differential pressure ΔPp across the throttle. The setting value ΔPLSref of the tilting control unit is modified by detecting an engine rotational speed based on change in the differential pressure across the throttle disposed in the delivery line of the fixed displacement hydraulic pump. The hydraulic drive system disclosed in JP, A, 60-11706 comprises a variable displacement hydraulic pump, a plurality of actuators driven by a hydraulic fluid delivered from the hydraulic pump, a plurality of flow control valves for controlling flow rates of the hydraulic fluid supplied from the hydraulic pump to the plurality of actuators, a plurality of pressure compensating valves controlling differential pressures across the plurality of flow control valves to become equal to each other, and a pump displacement control unit for controlling the displacement of the hydraulic pump so that a differential pressure ΔPLS between a delivery pressure Ps of the hydraulic pump and a maximum load pressure PLS among the plurality of actuators is maintained at a setting value ΔPLSref. The pressure compensating valves are installed upstream of the flow control valves, respectively. Each pressure compensating valve is arranged to receive the differential pressure across the flow control valve acting in the valve-closing direction and the differential pressure ΔPLS between the delivery pressure Ps of the hydraulic pump and the maximum load pressure PLS among the plurality of actuators in the valve-opening direction, for thereby controlling the differential pressure across the flow control valve with the differential pressure ΔPLS as a target differential pressure for pressure compensation. As a result, the differential pressures across the plurality of flow control valves are controlled to become equal to each other. DISCLOSURE OF THE INVENTION Consider now, as a comparative example, a system in which the pump displacement control system disclosed in JP, A, 5-99126 is used as a pump displacement control system for the hydraulic drive system disclosed in JP, A, 60-11706. In such a system, the target differential pressure across the flow control valve controlled by the pressure compensating valve is coincident with the setting value ΔPLSref of the differential pressure ΔPLS between the delivery pressure Ps of the hydraulic pump controlled by the pump displacement control means and the maximum load pressure PLS. The setting value ΔPLSref in the tilting control unit is therefore controlled in proportion to the engine rotational speed, and so is the target differential pressure (= ΔPLSref) across the flow control valve. In this case, setting is usually made such that a flow rate demanded by each of the actuators in the sole operation thereof does not exceed a maximum delivery rate of the hydraulic pump. As a result, in the sole operation of any one of the actuators, the hydraulic fluid is supplied to each actuator at a flow rate proportional to the amount of stroke by which the flow control valve is shifted, regardless of the engine rotational speed, thus ensuring good operability. On the other hand, when the maximum delivery rate of the hydraulic pump does not reach a flow rate demanded by all of the flow control valves in, e.g., the combined operation during which a plurality of actuators are driven simultaneously, there occurs a condition where the flow rate supplied to the actuators is insufficient (referred to as saturation hereinafter). Further, in the combined operation, if the engine rotational speed is set lower than the speed in ordinary work, the flow rate demanded by all of the flow control valves also lowers in proportion to the engine rotational speed because the target differential pressure ΔPLSref across each flow control valve is reduced in proportion to the engine rotational speed by the cooperation of the above-mentioned two conventional systems even in a combination of the same shift strokes of the flow control valves. However, since the maximum delivery rate of the hydraulic pump is also reduced in proportion to the engine rotational speed, a shortage of the flow rate occurs at the same proportion (see Fig. 4). Accordingly, when the shift stroke of the flow control valve enters the saturation region, the operation of the actuator in proportion to the shift stroke is no longer ensured, making an operator feel awkward. In practice, since excavation work carried out at the ordinary engine rotational speed requires response rather than operability in fine operation, the saturation phenomenon does not lead to a considerable problem. However, when the engine rotational speed is lowered for the purpose of carrying out fine operation, saturation occurs depending on the amount of stroke by which the flow control valve is shifted, thus giving the operator an awkward feeling. An object of the present invention is to provide a hydraulic drive system wherein good operability in fine operation can be obtained when an engine rotational speed is set to a low value, by improving a saturation phenomenon in consideration of the engine rotational speed. Features of the present invention to achieve the above object and other associated features are as follows. (1) To begin with, according to the present invention, there is provided a hydraulic drive system comprising an engine, a variable displacement hydraulic pump driven by the engine, a plurality of actuators driven by a hydraulic fluid delivered from the hydraulic pump, a plurality of flow control valves for controlling flow rates of the hydraulic fluid supplied from the hydraulic pump to a plurality of actuators, and pump displacement control means for controlling the displacement of the hydraulic pump so that a differential pressure ΔPLS between a delivery pressure Ps of the hydraulic pump and a maximum load pressure PLS among the plurality of actuators is maintained at a setting value ΔPLSref, the pump displacement control means being able to modify the setting value ΔPLSref depending on a rotational speed of the engine, characterized in that the hydraulic drive system further comprises: a plurality of pressure compensating valves for controlling respective differential pressures across the plurality of flow control valves to the same value as the differential pressure ΔPLS, and setting modifying means for detecting the rotational speed of the engine and, when the detected engine rotational speed is in a region including the lowest rotational speed of the engine, for modifying the setting value ΔPLSref of the pump displacement control means so that a total maximum flow rate Qvtotal of the plurality of flow control valves having respective flow rates expressed by the products of the differential pressure ΔPLS and respective opening areas of the plurality of flow control valves is smaller than a maximum delivery rate Qsmax of the hydraulic pump corresponding to the engine rotational speed at that time. By providing the setting modifying means to adjust the relationship between the total maximum demanded flow rate Qvtotal of the plurality of flow control valves and the maximum delivery rate Qsmax of the hydraulic pump, the total maximum demanded flow rate of the plurality of flow control valves is greater than the maximum delivery rate of the hydraulic pump and the system is under a condition giving rise to saturation when the engine rotational speed is set to the rated rotational speed suitable for ordinary work, but when the engine rotational speed is set to a low value, the total maximum demanded flow rate of the plurality of flow control valves is reduced to become smaller than the maximum delivery rate of the hydraulic pump and hence no saturation occurs. Accordingly, a change gradient of the flow rate passing through the plurality of flow control valves with respect to a total lever input amount applied to those the flow control valves is so reduced as to ensure a wide metering effective area, and good operability can be realized by using the wide metering effective area. (2) In the above (1), preferably, the setting modifying means comprises a fixed displacement hydraulic pump driven by the engine along with the variable displacement hydraulic pump, a flow rate detecting valve disposed in a delivery line of the fixed displacement hydraulic pump, and an operation driver for modifying the setting value ΔPLSref depending on a differential pressure ΔPp across the flow rate detecting valve, the flow rate detecting valve being constructed to have a larger opening area when the engine rotational speed is in the region including the rated rotational speed than when the engine rotational speed is in a region including the lowest rotational speed. With that feature, the setting modifying means can realize the function of the above (1) (i.e., the function of detecting the rotational speed of the engine and, when the detected engine rotational speed is in the region including the lowest rotational speed of the engine, modifying the setting value ΔPLSref of the pump displacement control means so that the total maximum flow rate Qvtotal of the flow control valves is smaller than the maximum delivery rate Qsmax of the hydraulic pump) by using hydraulic arrangement. (3) In the above (2), preferably, the flow rate detecting valve comprises a valve apparatus including a variable throttle, and throttle adjusting means for adjusting an opening area of the variable throttle to become smaller as the rotational speed of the engine lowers. With that feature, the flow rate detecting valve is constructed to have a larger opening area when the engine rotational speed is in the region including the rated rotational speed than when it is in the region including the lowest rotational speed, as set forth in the above (2). (4) In the above (2), alternatively, the flow rate detecting valve may comprise a valve apparatus including a fixed throttle, and throttle adjusting means for making the fixed throttle effective when the engine rotational speed is in the region including the lowest rotational speed, and controlling the fixed throttle to reduce an increase rate of the differential pressure across the flow rate detecting valve when the engine rotational speed rises to a certain setting rotational speed lower than the rated rotational speed. With that feature, the flow rate detecting valve is also constructed to have a larger opening area when the engine rotational speed is in the region including the rated rotational speed than when it is in the region including the lowest rotational speed, as set forth in the above (2). In addition, the flow rate detecting valve is constructed by using a fixed throttle and therefore it can be manufacture more easily. (5) In the above (3) or (4), preferably, the throttle adjusting means adjusts a position of the valve apparatus depending on the differential pressure ΔPp across the flow rate detecting valve itself. With that feature, the flow rate detecting valve can detect the engine rotational speed in a hydraulic manner and adjust the opening area of the variable throttle or the throttling condition of the fixed throttle depending on the engine rotational speed. (6) In the above (2), preferably, the setting modifying means further comprises a pressure control valve for generating a signal pressure corresponding to the differential pressure ΔPp across the flow rate detecting valve, the operation driver modifying the setting value ΔPLSref in accordance with a signal pressure from the pressure control valve. With that feature, since the signal pressure can be introduced via a single pilot line, therefore the circuit configuration is simplified. In addition, since the signal pressure is produced at a lower level, the pilot line can be formed of a hose or the like adapted for relatively low pressures, resulting in a reduced cost. (7) In the above (2), preferably, the pump displacement control means comprises a servo piston for operating a displacement varying mechanism of the variable displacement hydraulic pump, and a tilting control unit for driving the servo piston depending on the differential pressure ΔPLS between the delivery pressure Ps of the hydraulic pump and the load pressure PLS of the actuators, thereby maintaining the differential pressure ΔPLS at the setting value ΔPLSref, the tilting control unit including a spring for setting a basic value of the setting value ΔPLSref, the operation driver cooperating with the spring to variably set the setting value ΔPLSref. With that feature, the operation driver can modify the setting value ΔPLSref depending on the differential pressure across the flow rate detecting valve. BRIEF DESCRIPTION OF THE DRAWINGS Fig. 1 is a hydraulic circuit diagram showing the configuration of a hydraulic drive system and a pump displacement control system according to a first embodiment of the present invention. Fig. 2 is a diagram showing details of a flow rate detecting valve shown in Fig. 1. Figs. 3A to 3E are graphs showing the operation of the flow rate detecting valve in the first embodiment and the operation of a conventional valve for comparison between them. Fig. 4 is a graph showing the relationships of an engine rotational speed versus a maximum demanded flow rate of flow control valves and a maximum pump delivery rate in a conventional system. Fig. 5 is a graph showing the relationships of an engine rotational speed versus a maximum demanded flow rate of flow control valves and a maximum pump delivery rate as resulted from the provision of the flow rate detecting valve in the first embodiment. Fig. 6 is a graph showing the relationship between a total lever input amount and a flow rate passing through the flow control valves as resulted from the provision of the flow rate detecting valve in the first embodiment. Fig. 7 is a graph showing the relationships of an engine rotational speed versus a maximum demanded flow rate of flow control valves and a maximum pump delivery rate as resulted from the provision of the flow rate detecting valve in the first embodiment. Fig. 8 is a graph showing the relationship between a total lever input amount and a flow rate passing through the flow control valves as resulted from the provision of the flow rate detecting valve in the first embodiment. Fig. 9 is a hydraulic circuit diagram showing the configuration of a hydraulic drive system and a pump displacement control system according to a second embodiment of the present invention. Fig. 10 is a hydraulic circuit diagram showing the configuration of a hydraulic drive system and a pump displacement control system according to a third embodiment of the present invention. Fig. 11 is a diagram showing details of a flow rate detecting valve shown in Fig. 10. Figs. 12A to 12C are graphs showing the operation of the flow rate detecting valve in the third embodiment. Fig. 13 is a graph showing the relationships of an engine rotational speed versus a maximum demanded flow rate of flow control valves and a maximum pump delivery rate as resulted from the provision of the flow rate detecting valve in the third embodiment. BEST MODE FOR CARRYING OUT THE INVENTION Hereunder, embodiments of the present invention will be described with reference to the drawings. Fig. 1 shows a hydraulic drive system according to a first embodiment of the present invention. The hydraulic drive system comprises an engine 1, a variable displacement hydraulic pump 2 driven by the engine 1, a plurality of actuators 3a, 3b, 3c driven by a hydraulic fluid delivered from the hydraulic pump 2, a valve apparatus 4 including a plurality of directional control valves 4a, 4b, 4c connected to a delivery line 100 of the hydraulic pump 2 for controlling flow rates and directions at and in which the hydraulic fluid is supplied from the hydraulic pump 2 to the respective actuators 3a, 3b, 3c, and a pump displacement control system 5 for controlling the displacement of the hydraulic pump 2. The plurality of directional control valves 4a, 4b, 4c are made up of respectively a plurality of flow control valves 6a, 6b, 6c and a plurality of pressure compensating valves 7a, 7b, 7c for controlling differential pressures across the plurality of flow control valves 6a, 6b, 6c to become equal to each other. The plurality of pressure compensating valves 7a, 7b, 7c are of the pre-stage type installed upstream of the flow control valves 6a, 6b, 6c, respectively. The pressure compensating valve 7a has two pairs of opposing control pressure chambers 70a, 70b; 70c, 70d. Pressures upstream and downstream of the flow control valve 6a are introduced respectively to the control pressure chambers 70a, 70b, and a delivery pressure Ps of the hydraulic pump 2 and a maximum load pressure PLS among the plurality of actuators 3a, 3b, 3c are introduced respectively to the control pressure chambers 70c, 70d, whereby the differential pressure across the flow control valve 6a acts in the valve-closing direction and a differential pressure ΔPLS between the delivery pressure Ps of the hydraulic pump 2 and the maximum load pressure PLS among the plurality of actuators 3a, 3b, 3c acts in the valve-opening direction. Thus the pressure compensating valve 7a controls the differential pressure across the flow control valve 6a with the differential pressure ΔPLS as a target differential pressure for pressure compensation. The pressure compensating valves 7b, 7c are also of the same construction. Since the pressure compensating valves 7a, 7b, 7c control the respective differential pressures across the flow control valves 6a, 6b, 6c with the same differential pressure ΔPLS as a target differential pressure, the differential pressures across the flow control valves 6a, 6b, 6c are all controlled to become equal to the differential pressure ΔPLS and respective flow rates demanded by the flow control valves 6a, 6b, 6c are expressed by the products of the differential pressure ΔPLS and opening areas of those valves. The plurality of flow control valves 6a, 6b, 6c are provided with load ports 60a, 60b, 60c, respectively, through which load pressures of the actuators 3a, 3b, 3c are taken out during the operation of the actuators 3a, 3b, 3c. Maximum one of the load pressures taken out through the load ports 60a, 60b, 60c is detected by a signal line 10 via load lines 8a, 8b, 8c, 8d and shuttle valves 9a, 9b, the detected pressure being applied as the maximum load pressure PLS to the pressure compensating valves 7a, 7b, 7c. The hydraulic pump 2 is a swash plate pump wherein a delivery rate is increased by increasing a tilting angle of a swash plate 2a. The pump displacement control system 5 comprises a servo piston 20 for tilting the swash plate 2a of the hydraulic pump 2, and a tilting control unit 21 for driving the servo piston 20 to control the tilting angle of the swash plate 2a, thereby controlling the displacement of the hydraulic pump 2. The serve piston 20 is operated in accordance with a pressure introduced from the delivery line 100 (the delivery pressure Ps of the hydraulic pump 2) and a command pressure from the tilting control unit 21. The tilting control unit 21 includes a first tilting control valve 22 and a second tilting control valve 23. The first tilting control valve 22 is a horsepower control valve for reducing the delivery rate of the hydraulic pump 2 as the pressure introduced from the delivery line 100 (the delivery pressure Ps of the hydraulic pump 2) rises. The first tilting control valve 22 receives the delivery pressure Ps of the hydraulic pump 2, as an original pressure, and if the delivery pressure Ps of the hydraulic pump 2 is lower than a predetermined level set by a spring 22a, a spool 22b is moved to the right on the drawing, causing the delivery pressure Ps of the hydraulic pump 2 to be output as it is. At this time, if the output pressure is directly applied as a command pressure to the servo piston 20, the servo piston 20 is moved to the left on the drawing due to an area difference thereof between the opposite sides, whereupon the tilting angle of the swash plate 2a is increased to increase the delivery rate of the hydraulic pump 2. As a result, the delivery pressure Ps of the hydraulic pump 2 rises. When the delivery pressure Ps of the hydraulic pump 2 exceeds the predetermined level set by the spring 22a, the spool 22b is moved to the left on the drawing to reduce the delivery pressure Ps and a resulting reduced pressure is output as a command pressure. Accordingly, the servo piston 20 is moved to the right on the drawing, whereupon the tilting angle of the swash plate 2a is diminished to reduce the delivery rate Ps of the hydraulic pump 2. The second tilting control valve 23 is a load sensing control valve for controlling the differential pressure ΔPLS between the delivery pressure Ps of the hydraulic pump 2 and the maximum load pressure PLS among the actuators 3a, 3b, 3c to be maintained at the target differential pressure ΔPLSref. The second tilting control valve 23 comprises a spring 23a for setting a basic value of the target differential pressure ΔPLSref, a spool 23b, and a first operation driver 24 operated in accordance with the pressure introduced from the delivery line 100 (the delivery pressure Ps of the hydraulic pump 2) and the maximum load pressure PLS among the actuators 3a, 3b, 3c, for thereby moving the spool 23b. The first operation driver 24 comprises a piston 24a acting on the spool 23b and two hydraulic pressure chambers 24b, 24c divided by the piston 24a. The delivery pressure Ps of the hydraulic pump 2 is introduced to the hydraulic pressure chamber 24b, and the maximum load pressure PLS is introduced to the hydraulic pressure chamber 24c with the spring 23a built in the hydraulic pressure chamber 24c. Further, the second tilting control valve 23 receives the output pressure of the first tilting control valve 22, as an original pressure. When the differential pressure ΔPLS is lower than the target differential pressure ΔPLSref, the spool 23b is moved by the first operation driver 24 to the left on the drawing, causing the output pressure of the first tilting control valve 22 to be output as it is. At this time, if the output pressure of the first tilting control valve 22 is given by the delivery pressure Ps of the hydraulic pump 2, the delivery pressure Ps is applied as a command pressure to the servo piston 20. The servo piston 20 is therefore moved to the left on the drawing due to the area difference thereof between the opposite sides, whereupon the tilting angle of the swash plate 2a is increased to increase the delivery rate of the hydraulic pump 2. As a result, the delivery pressure Ps of the hydraulic pump 2 rises and the differential pressure ΔPLS also rises. On the other hand, when the differential pressure ΔPLS is higher than the target differential pressure ΔPLSref, the spool 23b is moved by the first operation driver 24 to the right on the drawing to reduce the output pressure of the first tilting control valve 22 and a resulting reduced pressure is output as a command pressure. Accordingly, the servo piston 20 is moved to the right on the drawing, whereupon the tilting angle of the swash plate 2a is diminished to reduce the delivery rate of the hydraulic pump 2. As a result, the differential pressure ΔPLS is maintained at the target differential pressure ΔPLSref. Here, the differential pressures across the flow control valves 6a, 6b, 6c are controlled respectively by the pressure compensating valves 7a, 7b, 7c so as to become the same value, i.e., the differential pressure ΔPLS. Therefore, maintaining the differential pressure ΔPLS at the target differential pressure ΔPLSref, as explained above, eventually results in that the differential pressures across the flow control valves 6a, 6b, 6c are maintained at the target differential pressure ΔPLSref. The pump displacement control system 5 further comprises setting modifying means 38 for modifying the target differential pressure ΔPLSref applied to the second tilting control valve 23 depending on change in rotational speed of the engine 1. The setting modifying means 38 is made up of a fixed displacement hydraulic pump 30 driven by the engine 1 along with the variable displacement hydraulic pump 2, a flow rate detecting valve 31 disposed to intermediate between delivery lines 30a, 30b of the fixed displacement hydraulic pump 30 and having a variable throttle 31a of which opening area is continuously adjustable, and a second operation driver 32 for modifying the target differential pressure ΔPLSref depending on a differential pressure ΔPp across the variable throttle 31a of the flow rate detecting valve 31. The fixed displacement hydraulic pump 30 is one that is usually provided to serve as a pilot hydraulic fluid source. A relief valve 33 for specifying an original pressure supplied from the pilot hydraulic fluid source is connected to the delivery line 30b, and the delivery line 30b is further connected to a remote control valve (not shown) for producing a pilot pressure used to shift the flow control valves 6a, 6b, 6c, for example. The second operation driver 32 is an additional operation driver integrated with the first operation driver 24 of the second tilting control valve 23, and comprises a piston 32a acting on the piston 24a of the first operation driver 24 and two hydraulic pressure chambers 32b, 32c divided by the piston 32a. A pressure upstream of the flow rate detecting valve (variable throttle 31a) is introduced to the hydraulic pressure chamber 32b via a pilot line 34a and a pressure downstream of the flow rate detecting valve (variable throttle 31a) is introduced to the hydraulic pressure chamber 32c via a pilot line 34b, causing the piston 32a to urge the piston 24a to the left on the drawing by a force corresponding to the differential pressure ΔPp across the variable throttle 31a of the flow rate detecting valve 31. The target differential pressure ΔPLSref provided by the second tilting control valve 23 is set in accordance with the basic value provided by the spring 23a and the urging force of the piston 32a. As the differential pressure ΔPp across the variable throttle 31a of the flow rate detecting valve 31 becomes smaller, the piston 32a pushes the piston 24a by a smaller force to reduce the target differential pressure ΔPLSref. As the differential pressure ΔPp becomes larger, the piston 32a pushes the piston 24a by a larger force to increase the target differential pressure ΔPLSref. Here, the differential pressure ΔPp across the variable throttle 31a of the flow rate detecting valve 31 varies depending on the rotational speed of the engine 1 (as described later). The second operation driver 32 thus modifies the target differential pressure ΔPLSref provided by the second tilting control valve 23 depending on the engine rotational speed. The flow rate detecting valve 31 is constructed such that the opening area of the variable throttle 31a is changed depending on the differential pressure ΔPp across the variable throttle 31a itself. More specifically, the flow rate detecting valve 31 comprises a valve body 31b, a spring 31c acting on the valve body 31b in the direction to reduce the opening area of the variable throttle 31a, a control pressure chamber 31d acting on the valve body 31b in the direction to increase the opening area of the variable throttle 31a, and a control pressure chamber 31e acting on the valve body 31b in the direction to reduce the opening area of the variable throttle 31a. The pressure upstream of the variable throttle 31a is introduced to the control pressure chamber 31d via a pilot line 35a and the pressure downstream of the variable throttle 31a is introduced to the control pressure chamber 31e via a pilot line 35b. The opening area of the variable throttle 31a is determined by balance among a force of the spring 31c and urging forces of the control pressure chambers 31d, 31e. As the differential pressure ΔPp across the variable throttle 31a becomes smaller, the valve body 31b is moved to the right on the drawing to reduce the opening area of the variable throttle 31a. As the differential pressure ΔPp becomes larger, the valve body 31b is moved to the left on the drawing to increase the opening area of the variable throttle 31a. Then, the differential pressure ΔPp across the variable throttle 31a varies depending on the rotational speed of the engine 1. Specifically, as the rotational speed of the engine 1 lowers, the delivery rate of the hydraulic pump 30 is reduced and the differential pressure ΔPp across the variable throttle 31a is also reduced. The control pressure chambers 31d, 31e and the spring 31c, therefore, function as throttle adjusting means for adjusting the opening area of the variable throttle 31a to become smaller as the rotational speed of the engine 1 lowers. Fig. 2 shows an internal structure of the flow rate detecting valve 31. In Fig. 2, a piston serving as the valve body 31b moves within a casing 31f and the area of a gap defined therebetween provides an opening area Ap of the variable throttle 31a. The piston 31b is supported by the spring 31c, and a resilient force F of the spring 31c acts on the piston 31b in the direction to reduce the opening area of the variable throttle 31a. Due to a flow of the hydraulic fluid in the casing 31f, the differential pressure ΔPp across the variable throttle 31a produces a force acting on the piston 31b in the direction to increase the opening area Ap of the variable throttle 31a. The piston 31b comes to a standstill in a position x where the above two forces are balanced. Since the resilient force F is proportional to a displacement x of the piston 31b with a spring constant K of the spring 31c as a constant of proportionality ( F = Kx), the differential pressure ΔPp across the variable throttle 31a is eventually proportional to the displacement x of the piston 31b (ΔPp α x). The relationship between the displacement x of the piston 31b and the opening area Ap of the variable throttle 31a depends on a shape of the casing 31f. In this embodiment, the casing 31f has a parabolic shape symmetrical with respect to the direction of displacement of the piston 31b. The operation and resulting effect of the setting modifying means 38 including the flow rate detecting valve 31, constructed as explained above, will now be described below. The fixed displacement hydraulic pump 30 delivers the hydraulic fluid at a flow rate Qp expressed by the product of a rotational speed N of the engine 1 and a pump displacement Cm. Qp = CmN Given the opening area of the variable throttle 31a of the flow rate detecting valve 31 being Ap, the rotational speed N of the engine 1 and the differential pressure ΔPp across the variable throttle 31a are related to each other by the following formula: Qp = cAp (2/ρ)ΔPp ΔPp = (ρ/2)(Qp/cAp) 2 = (ρ/2)(CmN/cAp) 2 Assuming now that the opening area Ap of the variable throttle 31a is not changed and remains constant (this case will be referred to as a comparative example hereinafter), the differential pressure ΔPp across the variable throttle 31a increases following a curve of secondary degree with respect to the delivery rate Qp of the hydraulic pump 30 or the rotational speed N of the engine 1 based on the formula (3), as shown in Fig. 3A. Also, since the relationship of ΔPLSref α ΔPp holds by virtue of the second operation driver 32, the load sensing setting differential pressure ΔPLSref also increases following a curve of secondary degree with respect to the delivery rate Qp of the hydraulic pump 30 or the rotational speed N of the engine 1, as shown in Fig. 3A. Further, supposing the case where the differential pressure ΔPLS across one of the flow control valves 6a, 6b, 6c, e.g., the flow control valve 6a, is controlled to the target differential pressure ΔPLSref, a flow rate Qv demanded by the flow control valve 6a is expressed by the following formula given an opening area of the flow control valve 6a being Av: Qv = CAv (2/ρ)ΔPlSref Thus the demanded flow rate Qv increases following a curve of secondary degree with respect to the target differential pressure ΔPLSref, as shown in Fig. 3C. Here, the target differential-pressure ΔPLSref across the flow control valve 6a is given by the differential pressure ΔPp across the variable throttle 31a of the flow rate detecting valve 31 (ΔPLSref α ΔPp). Based on the formula (3), therefore, the demanded flow rate Qv can be related to the rotational speed N of the engine 1 by the following formula: Qv α (Av/Ap)CmN Stated otherwise, as a combined result of the relationship between the flow rate Qp and the differential pressure ΔPp across the variable throttle 31a expressed by a curve of secondary degree (formula (3)) shown in Fig. 3A and the relationship between the differential pressure ΔPLS across the flow control valve 6a and the demanded flow rate Qv thereof expressed by a curve of secondary degree (formula (4)) shown in Fig. 3C, the demanded flow rate Qv increases almost linearly with respect to the rotational speed N of the engine 1, as shown in Fig. 3D. The above explanation is made for one flow control valve 6a. When driving a plurality of, e.g., two or three, actuators, the relationship of Fig. 3D is obtained for each of the flow control valves 6a, 6b or 6a, 6b, 6c, and the relationship between the rotational speed N of the engine 1 and a total of respective demanded rates Qv is given as one resulted from simply adding the relationship of Fig. 3D two or three times. Fig. 4 shows the relationships of the rotational speed N of the engine 1 versus a total maximum demanded flow rate Qvtotal of any two of the flow control valves 6a, 6b, 6c, e.g., the flow control valves 6a, 6b, (i.e., total of the flow rates Qv demanded by the flow control valves 6a, 6b at maximum opening areas thereof) and a maximum delivery rate Qsmax of the variable displacement hydraulic pump 2. Fig. 4 represents an example in which the opening area Ap of the variable throttle 31a of the flow rate detecting valve 31 is constant as stated above. When the actuators 3a, 3b are driven at the same time, a ratio of the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b to the maximum delivery rate Qsmax of the hydraulic pump 2 does not change regardless of change in the rotational speed N of the engine 1; hence a shortage of the flow rate accompanying with a saturation phenomenon during the combined operation occurs at the same proportion over an entire range of the rotational speed N of the engine 1. By contrast, the present invention constructed such that the opening area Ap of the variable throttle 31a of the flow rate detecting valve 31 is changed depending on the differential pressure across the variable throttle 31a. Supposing here that the casing 31f of the flow rate detecting valve 31 shown Fig. 2 has a parabolic shape symmetrical with respect to the direction of displacement of the piston 31b as stated above, the relationship between the opening area Ap of the variable throttle 31a and the differential pressure ΔPp across the variable throttle 31a is expressed by the following formula: Ap = a ΔPp From the formula (2), the relationship between the delivery rate Qp of the fixed displacement hydraulic pump 30 and the differential pressure ΔPp across the variable throttle 31a is expressed by the following formula (7): ΔPp = (1/Ca) (ρ/2)Qp =(Cm/Ca) V(ρ/2) · N Thus the differential pressure ΔPp across the variable throttle 31a increases linearly with respect to the delivery rate Qp of the hydraulic pump 30 or the rotational speed N of the engine 1, as shown in Fig. 3B. Also, from the relationship of ΔPLSref α ΔPp, the relationship between the demanded flow rate Qv of the flow control valve ta and the rotational speed N of the engine 1 is expressed by the following formula (8) similarly to the formula (5): Qv α cAv (Cm/Ca) (2/ρ) 1/2 · N Stated otherwise, as a combined result of the relationship between the flow rate Qp and the differential pressure ΔPp across the variable throttle 31a expressed by linear proportion (formula (7)) shown in Fig. 3B and the relationship between the differential pressure ΔPLS across the flow control valve 6a and the demanded flow rate Qv thereof expressed by a curve of secondary degree (formula (4)) shown in Fig. 3C, the demanded flow rate Qv increases following a curve of secondary degree with respect to the rotational speed N of the engine 1, as shown in Fig. 3E. Also, in this case, when driving a plurality of, e.g., two or three, actuators, the relationship of Fig. 3E is obtained for each of the flow control valves 6a, 6b or 6a, 6b, 6c, and the relationship between the rotational speed N of the engine 1 and a total of respective demanded rates Qv is given as one resulted from simply adding the relationship of Fig. 3E two or three times. Fig. 5 shows the relationships of the rotational speed N of the engine 1 versus a total maximum demanded flow rate Qvtotal of any two of the flow control valves 6a, 6b, 6c, e.g., the flow control valves 6a, 6b, (i.e., total of the flow rates Qv demanded by the flow control valves 6a, 6b at maximum opening areas thereof) and a maximum delivery rate Qsmax of the variable displacement hydraulic pump 2, the relationships being resulted based on Fig. 3E or the formula (8). In Fig. 5, at setting 1 where the rotational speed N of the engine 1 is set to be suitable for carrying out ordinary work, the system is under a condition giving rise to saturation because the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b when driving the plural actuators 3a, 3b is greater than the maximum delivery rate of the variable displacement hydraulic pump 2. On the other hand, at setting 2 where the rotational speed N of the engine 1 is set to a low value, the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b is reduced to become smaller than the maximum delivery rate of the hydraulic pump 2 and hence no saturation occurs. Here, the setting 2 represents an engine rotational speed suitable for fine operation. Specifically, since it is generally said that a rotational speed lower than the middle between the rated rotational speed and the lowest rotational speed is suitable for fine operation, the setting 2 corresponds to a rotational speed lower than the middle rotational speed. Assuming, for example, that the rated rotational speed of the engine 1 is 2,200 rpm and the lowest rotational speed (idling rotational speed) is 1,000 rpm, the middle rotational speed is 1,600 rpm and the setting 2 represents a rotational speed lower than 1,600 rpm. In the illustrated example, the setting 2 represents 1,200 rpm. Additionally, in the illustrated example, "the setting 1" represents the rated rotational speed of 2,200 rpm. As explained above, the flow rate detecting valve 31 is constructed to have a larger opening area when the engine rotational speed is in a region including the lowest rotational speed than when it is in a region including the rated rotational speed. The setting modifying means 38 made up of the flow rate detecting valve 31, the fixed displacement hydraulic pump 30 and the second operation driver 32 detects a rotational speed of the engine 1, and when the detected engine rotational speed is in the region including the lowest rotational speed, the means 38 modifies the setting value ΔPLSref of the pump displacement control system 5 so that the total maximum demanded flow rate Qvtotal of the plural flow control valves 6a, 6b, which is expressed based on the products of the differential pressure ΔPLS and the respective opening areas of the plural flow control valves 6a, 6b, is smaller than the maximum delivery rate Qsmax of the hydraulic pump 2 determined by the engine rotational speed at that time. Fig. 6 shows characteristics of the setting modifying means 38 in terms of the relationship between a total lever input amount applied from an operator to the flow control valves 6a, 6b and the total demanded flow rate of the flow control valves 6a, 6b (total flow rate passing therethrough). In Fig. 6, as the engine rotational speed lowers, the maximum flow rate Qsmax capable of being supplied from the hydraulic pump 2 to the flow control valves is reduced. Concurrently, the total demanded flow rate Qvtotal of the flow control valves 6a, 6b corresponding to the total lever input amount is reduced to become lower than the maximum delivery rate Qsmax of the hydraulic pump 2. Thus a gradient of the line representing change in the flow rate passing through the flow control valves 6a, 6b is so reduced as to ensure a wide metering effective area. In the above-mentioned comparative example, since, the ratio of the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b to the maximum delivery rate Qsmax of the hydraulic pump 2 does not change despite a lowering of the rotational speed N of the engine 1 and a shortage of the flow rate accompanying with a saturation phenomenon occurs at the same proportion as shown in Fig. 4, a gradient of the line representing change in the flow rate passing through the flow control valves 6a, 6b is so large as to narrow the metering effective area, as indicated by a one-dot-chain line in Fig. 6. Consequently, in the present invention, when the operator sets the engine rotational speed to a low value with the intent to carry out slow-speed operation, there occurs no saturation even with combined lever operations which give rise to saturation at the ordinary setting of the engine rotational speed; hence good operability can be realized using the wide metering effective area. Furthermore, in Fig. 7, at setting 3 where the rotational speed N of the engine 1 is set to a value (e.g., around 2,000 rpm) slightly lower than at the ordinary setting (setting 1), the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b is reduced a little from that at the ordinary setting (setting 1), but the amount of change is so small that the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b is held at a higher value than that resulted when providing the setting 3 in the comparative example. In such a condition, a saturation phenomenon tends to easily occur at engine rotational speeds around the setting value (setting 1) suitable for ordinary work. As indicated by a solid line in Fig. 8, however, a gradient of the line representing change in the flow rate passing through the flow control valves 6a, 6b with respect to the total lever input amount is not virtually changed from the gradient resulted at the setting 1. Accordingly, even when the rotational speed of the engine 1 is varied to some extent from the setting suitable for ordinary work, the operating speed of the actuator is kept at the same level and the operation can be performed with good response. In the comparative example, as indicated by a one-dot-chain line in Fig. 8, the gradient of the line representing change in the flow rate passing through the flow control valves 6a, 6b with respect to the total lever input amount is somewhat diminished, whereby the operating speed and response of the actuator are reduced correspondingly. Here, in ordinary work, grater importance is placed on response and powerful movement of the actuator rather than operability having a wider metering effective area from the practical point of view. Consequently, the present invention can provide the operator with a good feeling in the operation. With this embodiment, as stated above, a saturation phenomenon is improved in consideration of the engine rotational speed such that when the engine rotational speed is set to a low value, good operability in fine operation can be achieved, and when the engine rotational speed is set to a high value, a powerful feeling can be realized in the operation with good response. It is thus possible to establish the system setting adapted for the purpose of work intended by the operator based on setting of the engine rotational speed. Further, the relationship between the saturation phenomenon and the total lever input amount during the combined operation is freely adjustable depending on the shape of the casing 31f of the flow rate detecting valve 31. Additionally, in this embodiment, the characteristic of the maximum demanded flow rate Qvtotal, shown in Fig. 5, is obtained by forming the casing 31f of the flow rate detecting valve 31 to have a parabolic shape. However, the shape of the casing 31f may be a quasi-parabolic shape built up by combining a plurality of straight lines so long as when the engine rotational speed is in the region including the lowest rotational speed, the maximum demanded flow rate Qvtotal is smaller than the maximum delivery rate Qsmax of the hydraulic pump 2 determined by the engine rotational speed at that time. In this case, the casing 31f can be manufactured more easily. A second embodiment of the present invention will be described below with reference to Fig. 9. In Fig. 9, equivalent members to those in Fig. 1 are denoted by the same reference numerals and are not described here. Referring to Fig. 9, in a pump displacement control system 5A of this embodiment, setting modifying means 38A includes a pressure control valve 40 for outputting a signal pressure which corresponds to the differential pressure ΔPp across the variable throttle 31a of the flow rate detecting valve 31. The pressure control valve 40 has a pressure control chamber 40b urging a valve body 40a in the direction to increase pressure, and pressure control chambers 40c, 40d urging the valve body 40a in the direction to reduce pressure. The pressure upstream of the variable throttle 31a is introduced to the control pressure chamber 40b, whereas the pressure downstream of the variable throttle 31a and an output pressure of the pressure control valve 40 itself are introduced to the control pressure chambers 40c, 40d, respectively. The signal pressure which corresponds to the differential pressure ΔPp across the variable throttle 31a is produced as an absolute pressure based on balance among the above pressures. The signal pressure is introduced to the hydraulic pressure chamber 32b of the second operation driver 32A via a pilot line 41a, and the hydraulic pressure chamber 32c of the second operation driver 32A is communicated with a reservoir via a pilot line 41b. In this embodiment thus constructed, the second operation driver 32A likewise operates to modify the target differential pressure ΔPLSref depending on the differential pressure ΔPp across the variable throttle 31a of the flow rate detecting valve 31. Accordingly, this embodiment can also provide similar operating advantages as obtainable with the first embodiment. Further, while the embodiment shown in Fig. 1 requires the two pilot lines 34a, 34b for respectively introducing the pressure upstream of the flow rate detecting valve 31 and the pressure downstream thereof to the second operation driver 32, this embodiment requires only one pilot line 41a, resulting in a simpler circuit configuration. In addition, since the pressure control valve 40 detects the differential pressure as an absolute pressure, the signal pressure is produced at a lower level than the case of detecting the individual pressure as they are, resulting in that the pilot lines 41a, 41b can be formed of hoses or the like adapted for relatively low pressures and the circuit configuration can be achieved with a lower cost. A third embodiment of the present invention will be described below with reference to Figs. 10 to 13. In these drawings, equivalent members to those in Figs. 1 and 9 are denoted by the same reference numerals and are not described here. Referring to Fig. 10, in a pump displacement control system 5B of this embodiment, a flow rate detecting valve 31B of setting modifying means 38B has a valve body 31Bb provided with a fixed throttle 31Ba. When a differential pressure ΔPp across the flow rate detecting valve 31B introduced to control pressure chambers 31d, 31e is not larger than a differential pressure corresponding to the resilient force of a spring 31c (referred to as a setting differential pressure hereinafter), the flow rate detecting valve 31B is held in a left-hand position on the drawing where the fixed throttle 31Ba develops its function. When the differential pressure ΔPp across the flow rate detecting valve 31B becomes higher than the setting differential pressure, the flow rate detecting valve 31B is shifted to a right-hand open position on the drawing from the left-hand position on the drawing where the fixed throttle 31Ba develops its function. Fig. 11 shows an internal structure of the flow rate detecting valve 31B. In Fig. 11, a piston serving as the valve body 31Bb moves within a casing 31Bf and the piston 31Ba has a small hole formed therein to serve as the fixed throttle 31Ba. The small hole has an opening area Ap of the fixed throttle 31Ba. Further, the casing 31Bf has a cylindrical shape and a gap having an opening area Af is defined between an outer circumferential surface of the piston 31Bb and an inner circumferential surface of the casing 31Bf. The opening area Af is selected to a large value enough to prevent the gap from serving as a throttle in fact. The piston 31Bb is supported by the spring 31c, and a resilient force F of the spring 31c acts on the piston 31Bb in the direction to close an inlet of the casing 31Bf and to make the function of the fixed throttle 31Ba effective. When the inlet of the casing 31Bf is closed by the piston 31Bb, the differential pressure ΔPp across the fixed throttle 31Ba produces a hydraulic force Fh acting on the piston 31Bb in the direction to open the casing inlet (upward on the drawing) due to a flow of the hydraulic fluid in the casing 31f while passing the fixed throttle 31Ba. When the hydraulic force Fh is smaller than the force F of the spring 31c, the piston 31Bb is held in a state of keeping the inlet of the casing 31Bf closed, allowing the hydraulic fluid to flow just through the fixed throttle 31Ba. In other words, the fixed throttle 31Ba functions effectively. When a flow rate of the hydraulic fluid delivered from the fixed displacement pump 30 increases and the hydraulic force Fh exceeds the force F of the spring 31c, the piston 31Bb is moved upward to open the casing inlet. In this state, the hydraulic fluid is allowed to flow through the gap having the opening area Af and therefore the fixed throttle 31Ba does no longer function. Since the hydraulic force Fh is eliminated upon the fixed throttle 31Ba stopping the function, the piston 31Bb is moved downward to close the casing inlet. However, as soon as the casing inlet is closed, the hydraulic force is generated to open the casing inlet again. As a result of repeating the above up and down movement, the piston 31Bb comes to a standstill in a position x where the two forces F and Fh are balanced. In the standstill position, throttle control is performed so that the differential pressure ΔPp across the flow rate detecting valve 31B is maintained at the differential pressure corresponding to the resilient force of a spring 31c, i.e., the setting differential pressure. Here, the differential pressure ΔPp across the flow rate detecting valve 31B introduced to the control pressure chambers 31d, 31e as explained above varies depending on the rotational speed of the engine 1. Specifically, as the rotational speed of the engine 1 lowers, the delivery rate of the hydraulic pump 30 is reduced and the differential pressure ΔPp across the flow rate detecting valve 31B is also reduced. Accordingly, when the engine rotational speed is lower than an engine rotational speed corresponding to the setting differential pressure specified by the spring 31c (referred to as a setting rotational speed hereinafter), the flow rate detecting valve 31B is held in a position where the fixed throttle 31Ba develops its function (i.e., the left-hand position in Fig. 10), and when the engine rotational speed exceeds the setting rotational speed, the flow rate detecting valve 31B controls a throttle condition so as to maintain the differential pressure ΔPp across the flow rate detecting valve 31B at the setting differential pressure specified by the spring 31c. Stated otherwise, the control pressure chambers 31d, 31e and the spring 31c function as throttle adjusting means for making the fixed throttle 31Ba effective when the engine rotational speed is in the region including the lowest rotational speed, and controlling the fixed throttle 31Ba to reduce an increase rate of the differential pressure ΔPp across the flow rate detecting valve 31B when the engine rotational speed rises to a certain setting rotational speed lower than the rated rotational speed. Also, as a result of the above arrangement, the flow rate detecting valve 31B is constructed to have a larger opening area when the engine rotational speed is in the region including the rated rotational speed than when it is in the region including the lowest rotational speed. The operation and resulting effect of the setting modifying means 38B including the flow rate detecting valve 31B, constructed as explained above, will now be described below. Assuming that the setting rotational speed corresponding to the resilient force of the spring 31c of the flow rate detecting valve 31B is Ns, when the engine rotational speed N is lower than the setting rotational speed Ns, the flow rate detecting valve 31B is held in the left-hand position in Fig. 10 where the fixed throttle 31Ba develops its function, as explained above, and the opening area Ap is constant. Based on the aforesaid formula (3), therefore, the differential pressure ΔPp across the flow rate detecting valve 31B increases following a curve of secondary degree with respect to the delivery rate Qp of the hydraulic pump 30 or the rotational speed N of the engine 1, as shown in Fig. 12A. It to be noted that the opening area Ap of the fixed throttle 31Ba is set smaller than that of the fixed throttle in the comparative example and eventually an increase rate of the differential pressure ΔPp across the fixed throttle is higher than in the comparative example indicated by a dotted line. When the engine rotational speed N exceeds the setting rotational speed Ns, the flow rate detecting valve 31B operates so as to maintain the differential pressure ΔPp across itself at the setting differential pressure specified by the spring 31c. The differential pressure ΔPp across the flow rate detecting valve 31B is therefore kept substantially constant at ΔPpmax, as shown in Fig. 12A. In a like manner as explained above in connection with Fig. 3C, a flow rate Qv demanded by each of the flow control valves 6a, 6b, 6c increases following a curve of secondary degree with respect to the target differential pressure ΔPLSref, as shown in Fig. 12B. As a combined result of the characteristic of Fig. 12A and the characteristic of Fig. 12B, the demanded flow rate Qv varies with respect to the rotational speed N of the engine 1, as shown in Fig. 12C. More specifically, when the engine rotational speed N is lower than the setting rotational speed Ns, the change of ΔPp represented by a curve of secondary degree shown in Fig. 12A and the change of the demanded flow rate Qv represented by a curve of secondary degree shown in Fig. 12B cancel each other. As a result, the demanded flow rate Qv increases almost linearly with respect to the rotational speed N of the engine 1. A gradient of the linear line (change rate) is however greater than in the comparative example indicated by a dotted line. When the engine rotational speed N exceeds the setting rotational speed Ns, ΔPp in Fig. 12A is kept substantially constant at ΔPpmax and therefore the demanded flow rate Qv is also kept substantially constant correspondingly. As stated above, when driving a plurality of, e.g., two or three, actuators, the relationship of Fig. 12C is obtained for each of the flow control valves 6a, 6b or 6a, 6b, 6c, and the relationship between the rotational speed N of the engine 1 and a total of respective demanded rates Qv is given as one resulted from simply adding the relationship of Fig. 12C two or three times. Fig. 13 shows the relationships of the rotational speed N of the engine 1 versus a total maximum demanded flow rate Qvtotal of any two of the flow control valves 6a, 6b, 6c, e.g., the flow control valves 6a, 6b, (i.e., total of the flow rates Qv demanded by the flow control valves 6a, 6b at maximum opening areas thereof) and a maximum delivery rate Qsmax of the variable displacement hydraulic pump 2, the relationships being obtained based on Fig. 12C. As seen from Fig. 13, also in this embodiment, when the engine rotational speed N is lower than the setting rotational speed Ns, the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b is smaller than the maximum delivery rate Qsmax of the hydraulic pump 2 determined by the engine rotational speed at that time. Therefore, at setting 1 where the rotational speed N of the engine 1 is set to be suitable for carrying out ordinary work, the system is under a condition giving rise to saturation because the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b when driving the plural actuators 3a, 3b is greater than the maximum delivery rate of the hydraulic pump 2. On the other hand, at setting 2 where the rotational speed N of the engine 1 is set to a low value, the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b is reduced to become smaller than the maximum delivery rate of the hydraulic pump 2 and hence no saturation occurs. Accordingly, as explained above in connection with the first embodiment by referring to Fig. 6, when the engine rotational speed is lowered, the total demanded flow rate Qvtotal of the flow control valves 6a, 6b corresponding to the total lever input amount is held lower than the maximum delivery rate Qsmax of the hydraulic pump 2 in spite of reduction in the maximum flow rate Qsmax capable of being supplied from the hydraulic pump 2 to the flow control valves. Thus a gradient of the line representing change in the flow rate passing through the flow control valves 6a, 6b is so reduced as to ensure a wide metering effective area. Furthermore, in Fig. 13, at setting 3 where the rotational speed N of the engine 1 is set to a value slightly lower than at the ordinary setting (setting 1), the demanded flow rate Qvtotal of the flow control valves 6a, 6b is reduced a little from that at the ordinary setting (setting 1), but the amount of change is not appreciable and the total maximum demanded flow rate Qvtotal of the flow control valves 6a, 6b is held at a higher value than that resulted when providing the setting 3 in the comparative example. As explained above in connection with the first embodiment by referring to Fig. 8, however, a gradient of the line representing change in the flow rate passing through the flow control valves 6a, 6b with respect to the total lever input amount is not virtually changed from the gradient resulted at the setting 1, thus enabling the operation to be performed with good response. As a result, this embodiment can also provide similar operating advantages as obtainable with the first embodiment in that when the engine rotational speed is set to a low value, good operability in fine operation can be achieved, and when the engine rotational speed is set to a high value, a powerful feeling can be realized in the operation with good response. Further, this embodiment can provide a practical flow rate detecting valve because the casing 31Bf of the flow rate detecting valve 31B has a simple cylindrical shape and hence can be manufactured very easily. It is to be noted that while the above embodiments have been explained as detecting the engine rotational speed and modifying the target differential pressure based on the detected speed in a hydraulic manner, such a process may be performed electrically by, e.g., detecting the engine rotational speed with a sensor and calculating the target differential pressure from a sensor signal. Additionally, while the pressure compensating valves have been described as being of the pre-stage type installed upstream of the flow control valves, the pressure compensating valves may be of the post-stage type installed downstream of the flow control valves to control respective output pressures of all the flow control valves to the same maximum load pressure, thereby controlling respective differential pressures across the flow control valves to the same differential pressure ΔPLS. INDUSTRIAL APPLICABILITY According to the present invention, it is possible to establish the system setting adapted for the purpose of work intended by the operator based on setting of the engine rotational speed and to realize a good feeling in the operation.

Claims: A hydraulic drive system comprising an engine (1), a variable displacement hydraulic pump (2) driven by said engine, a plurality of actuators (3a, 3b) driven by a hydraulic fluid delivered from said hydraulic pump, a plurality of flow control valves (6a, 6b) for controlling flow rates of the hydraulic fluid supplied from said hydraulic pump to a plurality of actuators, and pump displacement control means (5, 5A, 5B) for controlling the displacement of said hydraulic pump so that a differential pressure ΔPLS between a delivery pressure Ps of said hydraulic pump and a maximum load pressure PLS among said plurality of actuators is maintained at a setting value ΔPLSref, said pump displacement control means being able to modify the setting value ΔPLSref depending on a rotational speed of said engine, characterized in that said hydraulic drive system further comprises: a plurality of pressure compensating valves (7a, 7b) for controlling respective differential pressures across said plurality of flow control valves (6a, 6b) to the same value as said differential pressure ΔPLS, and setting modifying means (38, 38A, 38B) for detecting the rotational speed of said engine (1) and, when the detected engine rotational speed is in a region including the lowest rotational speed of said engine, for modifying the setting value ΔPLSref of said pump displacement control means (5, 5A, 5B) so that a total maximum flow rate Qvtotal of said plurality of flow control valves (6a, 6b) having respective flow rates expressed by the products of said differential pressure ΔPLS and respective opening areas of said plurality of flow control valves (6a, 6b) is smaller than a maximum delivery rate Qsmax of said hydraulic pump (2) corresponding to the engine rotational speed at that time. A hydraulic drive system according to Claim 1, wherein said setting modifying means (38) comprises a fixed displacement hydraulic pump (30) driven by said engine (1) along with said variable displacement hydraulic pump (2), a flow rate detecting valve (31, 31B) disposed in a delivery line (30b) of said fixed displacement hydraulic pump, and an operation driver (32, 32A) for modifying said setting value ΔPLSref depending on a differential pressure ΔPp across said flow rate detecting valve, said flow rate detecting valve being constructed to have a larger opening area when the engine rotational speed is in the region including the rated rotational speed than when the engine rotational speed is in a region including the lowest rotational speed. A hydraulic drive system according to Claim 2, wherein said flow rate detecting valve (31) comprises a valve apparatus (31b) including a variable throttle (31a), and throttle adjusting means (31c, 31d, 31e) for adjusting an opening area of said variable throttle (31a) to become smaller as the rotational speed of said engine (1) lowers. A hydraulic drive system according to Claim 2, wherein said flow rate detecting valve (31B) comprises a valve apparatus (31Bb) including a fixed throttle (31Ba), and throttle adjusting means (31c, 31d, 31e) for making said fixed throttle (31Ba) effective when the engine rotational speed is in the region including the lowest rotational speed, and controlling said fixed throttle (31Ba) to reduce an increase rate of the differential pressure across said flow rate detecting valve when the engine rotational speed rises to a certain setting rotational speed lower than the rated rotational speed. A hydraulic drive system according to Claim 3 or 4, wherein said throttle adjusting means (31c, 31d, 31e) adjusts a position of said valve apparatus (31b, 31Bb) depending on the differential pressure ΔPp across said flow rate detecting valve (31, 31B) itself. A hydraulic drive system according to Claim 2, wherein said setting modifying means (38A) further comprises a pressure control valve (40) for generating a signal pressure corresponding to the differential pressure ΔPp across said flow rate detecting valve (31), said operation driver (32A) modifying said setting value ΔPLSref in accordance with a signal pressure from said pressure control valve. A hydraulic drive system according to Claim 2, wherein said pump displacement control means (5, 5A, 5B) comprises a servo piston (20) for operating a displacement varying mechanism (2a) of said variable displacement hydraulic pump (2), and a tilting control unit (21) for driving said servo piston depending on the differential pressure ΔPLS between the delivery pressure Ps of said hydraulic pump (2) and the load pressure PLS of said actuators (3a, 3b), thereby maintaining the differential pressure ΔPLS at said setting value ΔPLSref, said tilting control unit including a spring (23a) for setting a basic value of said setting value ΔPLSref, said operation driver (32, 32A) cooperating with said spring to variably set said setting value ΔPLSref.

Inventors: KANAI TAKASHI, KAWAMOTO JUNYA, TSURUGA YASUTAKA, KANAI, TAKASHI, KAWAMOTO, JUNYA, TSURUGA, YASUTAKA
Applicants: HITACHI CONSTRUCTION MACHINERY, HITACHI CONSTRUCTION MACHINERY CO., LTD.
Application year: 1997
Main classification: F15B11
IPC codes: F15B11, F04B49, E02F9