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Inventors
Tsuruga, Yasutaka
Kanai, Takashi
Kawamoto, Junya
Nakatani, Kenichiro
Takahashi, Kiwamu
Hamamoto, Satoshi
Okazaki, Yasuharu
Nagao, Yukiaki
Application #
601518
Filed
Aug-2-2000
Published
Jun-4-2002
Current US Class
060/422
091/446
091/516
International Classes
F15B 013/06
Field of Search
60/420 60/469 60/422 91/516 91/446
Assignee
Hitachi Construction Machinery Co., Ltd. (Tokyo, JP)
Examiners
Look; Edward K.
Attorney, Agent or Firm
Mattingly, Stanger & Malur, P.C.
US Patent References
| 4617854 |
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Multiple consumer... |
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| 5937645 |
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Hydraulic device |
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Referenced by:
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Citation
Cite This Patent
More From Subclass 446
More From Class 091
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Abstract
A hydraulic drive system includes a pump control unit 18 for controlling a pump delivery rate such that a pump delivery pressure is held at a predetermined value higher than a maximum load pressure among actuators 2-6. Pressure compensating valves 12-16 are each constructed to set, as a target compensation differential pressure, a differential pressure between the delivery pressure of a hydraulic pump 1 and the maximum load pressure among the actuators 2-6. The pressure compensating valve 12 is given such a load dependent characteristic that the target compensation differential pressure is reduced when a load pressure rises. A lower limit setting spring 55 for limiting the target compensation differential pressure from becoming smaller than a predetermined value is provided in the pressure compensating valve 12 for the swing section.
Claims
What is claimed is:
1. A hydraulic drive system comprising a hydraulic pump, a plurality of actuators, including a swing motor, which are driven by a hydraulic fluid delivered from said hydraulic pump, a plurality of directional control valves for controlling respective flow rates of the hydraulic fluid supplied from said hydraulic pump to said plurality of actuators, a plurality of pressure compensating valves for controlling respective differential pressures across said plurality of directional control valves, and pump control means for load sensing control to control a pump delivery rate such that a delivery pressure of said hydraulic pump is held a predetermined value higher than a maximum load pressure among said plurality of actuators, wherein said hydraulic drive system further comprises:
first means provided respectively in those of said plurality of pressure compensating valves, which are not for a swing section associated with said swing motor, and setting, as a target compensation differential pressure, a differential pressure between the delivery pressure of said hydraulic pump and the maximum load pressure among said plurality of actuators,
second means provided in the pressure compensating valve for the swing section and setting a target compensation differential pressure of the pressure compensating valve,
third means provided in at least one of said plurality of pressure compensating valves, which is for the swing section, and reducing the target compensation differential pressure set by said second means when a load pressure of said swing motor rises, thereby giving a load dependent characteristic to the pressure compensating valve for the swing section, and
fourth means provided in the pressure compensating valve for the swing section and setting a lower limit of the target compensation differential pressure that is set by said second means and modified by said third means.
2. A hydraulic drive system according to claim 1,
wherein said second means is means for setting, as the target compensation differential pressure, the differential pressure between the delivery pressure of said hydraulic pump and the maximum load pressure among said plurality of actuators as with said first means, and
said fourth means functions as lower limit setting means for limiting both reduction in the target compensation differential pressure itself set by said second means and reduction in the target compensation differential pressure due to the load dependent characteristic given by said third means.
3. A hydraulic drive system according to claim 1,
wherein said fourth means is biasing means for applying a biasing force to a spool of the pressure compensating valve for the swing section in the valve-opening direction when the target compensation differential pressure set by said second means and modified by said third means reaches a predetermined value.
4. A hydraulic drive system according to claim 3,
wherein said biasing means is biasing means for applying a biasing force to a spool of the pressure compensating valve for the swing section in the valve-opening direction when the target compensation differential pressure set by said second means and modified by said third means reaches a predetermined value.
5. A hydraulic drive system according to claim 1,
wherein said pressure compensating valve (12) for the swing section comprises a spool and a weak spring (350) for holding the spool in an initial-position and said fourth means (55) is provided separately from said spring.
6. A hydraulic drive system comprising a hydraulic pump, a plurality of actuators, including a swing motor, which are driven by a hydraulic fluid delivered from said hydraulic pump, a plurality of directional control valves for controlling respective flow rates of the hydraulic fluid supplied from said hydraulic pump to said plurality of actuators, a plurality of pressure compensating valves for controlling respective differential pressures across said plurality of directional control valves, and pump control means for load sensing control to control a pump delivery rate such that a delivery pressure of said hydraulic pump is held a predetermined value higher than a maximum load pressure among said plurality of actuators, wherein said hydraulic drive system further comprises:
first means provided respectively in those of said plurality of pressure compensating valves, which are not for a swing section associated with said swing motor, and setting, as a target compensation differential pressure, a differential pressure between the delivery pressure of said hydraulic pump and the maximum load pressure among said plurality of actuators,
second means provided in the pressure compensating valve for the swing section and setting a target compensation differential pressure of the pressure compensating valve,
third means provided in at least one of said plurality of pressure compensating valves, which is for the swing section, and reducing the target compensation differential pressure set by said second means when a load pressure of said swing motor rises, thereby giving a load dependent characteristic to the pressure compensating valve for the swing section, and
fourth means provided in the pressure compensating valve for the swing section and setting a lower limit of the target compensation differential pressure that is set by said second means and modified by said third means; and
wherein said second means is means for setting, as the target compensation differential pressure, a value not changed depending on the differential pressure between the delivery pressure of said hydraulic pump and the maximum load pressure among said plurality of actuators, and
said fourth means functions as lower limit setting means for limiting reduction in the target compensation differential pressure due to the load dependent characteristic given by said third means.
7. A hydraulic drive system comprising a hydraulic pump, a plurality of actuators, including a swing motor, which are driven by a hydraulic fluid delivered from said hydraulic pump, a plurality of directional control valves for controlling respective flow rates of the hydraulic fluid supplied from said hydraulic pump to said plurality of actuators, a plurality of pressure compensating valves for controlling respective differential pressures across said plurality of directional control valves, and pump control means for load sensing control to control a pump delivery rate such that a delivery pressure of said hydraulic pump is held a predetermined value higher than a maximum load pressure among said plurality of actuators, wherein said hydraulic drive system further comprises:
first means provided respectively in those of said plurality of pressure compensating valves, which are not for a swing section associated with said swing motor, and setting, as a target compensation differential pressure, a differential pressure between the delivery pressure of said hydraulic pump and the maximum load pressure among said plurality of actuators,
second means provided in the pressure compensating valve for the swing section and setting a target compensation differential pressure of the pressure compensating valve,
third means provided in at least one of said plurality of pressure compensating valves, which is for the swing section, and reducing the target compensation differential pressure set by said second means when a load pressure of said swing motor rises, thereby giving a load dependent characteristic to the pressure compensating valve for the swing section, and
fourth means provided in the pressure compensating valve for the swing section and setting a lower limit of the target compensation differential pressure that is set by said second means and modified by said third means; and
wherein said fourth means is biasing means for always adding a supplement value to the target compensation differential pressure that is set by said second means and modified by said third means, and
the directional control valve for the swing section is constructed such that meter-in variable throttles thereof each have an opening area smaller than that in the directional control valves not for the swing section by an amount of the target compensation differential pressure corresponding to the supplement value added by said biasing means.
8. A hydraulic drive system according to claim 7,
wherein said biasing means is a swing priority spring always acting on said spool of the pressure compensating valve for the swing section in the valve-opening direction.
Description
TECHNICAL FIELD
The present invention relates to a hydraulic drive system for a construction machine including a swing control system, such as a hydraulic excavator. More particularly, the present invention relates to a hydraulic drive system wherein, when a hydraulic fluid from a hydraulic pump is supplied to a plurality of actuators, including a swing motor, through respective associated directional control valves, a delivery rate of the hydraulic pump is controlled by a load sensing system and differential pressures across the directional control valves are controlled by respective associated pressure compensating valves.
BACKGROUND ART
JP, A, 60-11706 discloses a hydraulic drive system for controlling a delivery rate of a hydraulic pump by a load sensing system (hereinafter referred to also as an LS system). Also, JP, A, 10-37907 discloses a hydraulic drive system for a construction machine including a swing control system, the hydraulic drive system including an LS system and being intended to realize independence and operability of the swing control system. A 3-pump system mounted on an actual machine is also disclosed as an open-center hydraulic drive system for a construction machine including a swing control system, the hydraulic drive system being intended to realize independence of the swing control system. Further, JP, A, 10-89304 discloses a hydraulic drive system wherein a delivery rate of a hydraulic pump is controlled by an LS system and a pressure compensating valve is given a load dependent characteristic.
In the hydraulic drive system disclosed in JP, A, 60-11706, a plurality of pressure compensating valves each include means for setting, as a target compensation differential pressure, a differential pressure between a delivery pressure of the hydraulic pump and a maximum load pressure among a plurality of actuators. In the combined operation where a plurality of actuators are simultaneously driven, there occurs a saturation state that the delivery rate of the hydraulic pump is not enough to supply flow rates demanded by a plurality of directional control valves. In such a saturation state, the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure is lowered, and correspondingly the target compensation differential pressure of each pressure compensating valve is reduced. As a result, the delivery rate of the hydraulic pump can be redistributed in accordance with a ratio between the respective flow rates demanded by the actuators.
In the hydraulic drive system disclosed in JP, A, 10-37907 and the 3-pump system mounted on an actual machine, an independent open-center circuit using an independent hydraulic pump is constructed for a swing section, which includes a swing motor, separately from a circuit for the other actuators, whereby independence and operability of the swing control system is ensured.
In the hydraulic drive system disclosed in JP, A, 10-89304, a plurality of pressure compensating valves each have hydraulic pressure chambers constructed as follows. A pressure bearing area of a hydraulic pressure chamber, to which an input side pressure of a directional control valve is introduced and which produces a force acting in the valve-closing direction, is set to be greater than a pressure bearing area of a hydraulic pressure chamber, to which an output side pressure of the directional control valve is introduced and which produces a force acting in the valve-opening direction. The pressure compensating valve is thereby given such a load dependent characteristic that, as a load pressure of each associated actuator rises, the target compensation differential pressure of the pressure compensating valve is reduced (i.e., the pressure compensating valve is throttled) to decrease a supply flow rate to the actuator. As a result, the actuators on both the lower and higher load sides can be operated with good operability in a stable manner without hunting.
DISCLOSURE OF THE INVENTION
The conventional hydraulic drive systems described above however have the following problems with the swing control system.
JP, A, 60-11706: problems 1 and 2
JP, A, 10-89304: problems 2 and 3
JP, A, 10-37907: problem 4
Open-center 3-pump system mounted on actual machine: problem 4
1 jerky feel in operation at start-up of swing alone
2 change of the swing speed at shift from operation of swing alone to combined operation including swing and vice versa.
3 extreme drop of the swing speed at start-up of combined operation including swing
4 increase in cost and space and complicated circuit configuration due to provision of a separate circuit
(1) JP, A, 60-11706
When the hydraulic drive system including the LS system, disclosed in JP, A, 60-11706, is applied to the swing control system, it is difficult to keep balance between load sensing control (hereinafter referred to also as LS control) of the hydraulic pump and a flow rate compensating function of the pressure compensating valve due to an inertial load of the swing control system. This is because a difficulty occurs in keeping balance between response of the pressure compensating valve and response in the LS control of the hydraulic pump due to the following reasons when a swing driving pressure is controlled in a stage of shift from swing acceleration to steady rotation.
(1) In a swing start-up and acceleration mode, the pump LS control is performed so as to raise a delivery pressure of the hydraulic pump depending on the swing start-up pressure for holding a constant flow rate.
(2) To hold constant a differential pressure across a throttling element of the directional control valve, the pressure compensating valve is operated in a direction to increase a flow rate passing itself that tends to reduce upon a rise of the load pressure.
(3) When the swing reaches a steady speed, the swing driving pressure is lowered and therefore the pump LS control is not required to control the delivery pressure of the hydraulic pump so high as in the swing start-up and acceleration mode. Hence the pump LS control is performed in a direction to lower the delivery pressure of the hydraulic pump.
(4) Upon a lowering of the swing driving pressure, the pressure compensating valve is operated in a direction to reduce the flow rate passing itself that tends to increase.
Because of quick shift from (1) to (4), the swing operation becomes jerky (above problem 1).
In the combined operation, as described above, there occurs a saturation state that the delivery rate of the hydraulic pump is not enough to supply flow rates demanded by a plurality of directional control valves. Corresponding to such a saturation state, the target compensation differential pressure of each pressure compensating valve is reduced, and the delivery rate of the hydraulic pump is redistributed in accordance with a ratio between the respective flow rates demanded by the actuators. With that function, even in the combined operation, the actuators are operated, although slightly slowed down, by the hydraulic fluid distributed at the ratio depending on the intended operations, whereby a feel in the operation is not impaired.
However, such slowdown likewise occurs in the swing operation, and during the combined operation including swing, the swing speed is also reduced as with one or more other actuators. This slowdown gives rise to change of the swing speed at shift from the swing-combined operation to the swing-alone operation and vice versa, thus causing the operator to feel awkward (above problem 2).
(2) JP, A, 10-89304
In the hydraulic drive system disclosed in JP, A, 10-89304, since the pressure compensating valve is given a load dependent characteristic, the target compensation differential pressure of the pressure compensating valve is reduced in response to a rise of the load pressure of the swing motor at the start-up of swing alone, and when the swing motor shifts to a steady sate, the target compensation differential pressure of the pressure compensating valve is also returned to the original value in response to a lowering of the load pressure of the swing motor. As a result, the swing can be started up without causing a jerky feel in operation. However, when the delivery rate of the hydraulic pump comes into a saturation state in the combined operation, the delivery rate of the hydraulic pump is redistributed in accordance with a ratio between the respective flow rates demanded by the directional control valves, as with the hydraulic drive system disclosed in JP, A, 60-11706. Accordingly, the swing speed is changed at shift from the swing-combined operation to the swing-alone operation and vice versa, thus causing the operator to feel awkward (above problem 2).
Further, since the pressure compensating valve is given a load dependent characteristic, the target compensation differential pressure of the pressure compensating valve for the swing section is reduced depending on the condition of the delivery rate of the hydraulic pump at the start-up of the swing-combined operation. In addition, the target compensation differential pressure is also reduced due to the load dependent characteristic as the load pressure of the swing motor rises up to a relief pressure. Such a reduction in the target compensation differential pressure continues until the swing motor shifts to the steady sate. As a result, the swing speed is extremely lowered as compared with the speeds of other actuators at the start-up of the swing-combined operation, whereby swing operability at the start-up of the swing-combined operation is deteriorated (above problem 3).
(3) Hydraulic Drive System Disclosed in JP, A, 10-37907 and Open-center 3-Pump System Mounted on Actual Machine
In the hydraulic drive system disclosed in JP, A, 10-37907, the swing control system is constructed by a separate open-center circuit to ensure satisfactory swing operability in the LS system. Also, in the open-center 3-pump system mounted on an actual machine, the swing control system is constructed as a separate open-center circuit to ensure satisfactory swing operability.
More specifically, in the open-center system, when the driving pressure rises at the swing start-up, a flow rate of the hydraulic fluid returning to a reservoir through a center bypass fluid line is increased, which reduces a flow rate of the hydraulic fluid passing a throttle of the directional control valve for the swing section. A flow rate of the hydraulic fluid supplied to the swing motor is therefore restricted in the swing start-up and acceleration mode. When the swing speed reaches a steady speed, no restriction is imposed on the supply flow rate to the swing motor because of the driving pressure being not so high as at the swing start-up, and the hydraulic fluid is supplied to the swing motor at a flow rate corresponding to an opening of the throttle of the directional control valve for the swing section. The swing can be thereby smoothly started up without causing a jerky feel in operation for starting up the swing solely unlike the LS control.
Although the above problem 2 occurs in not only the LS system but also the open-center system, change of the swing speed is not caused in the hydraulic drive system and the open-center 3-pump system mounted on an actual machine, which are disclosed in JP, A, 10-37907, because the swing control system is constructed as the separate open-center circuit and independence of the swing control system is realized.
However, in the hydraulic drive system disclosed in JP, A, 10-37907 and the 3-pump system mounted on an actual machine, the swing control system must be constructed as a separate circuit in parallel to the system for the other actuators. Correspondingly, a cost is pushed up and a space required for installation is increased. In addition, a hydraulic pump for the swing control system must be separately provided. In the system disclosed in JP, A, 10-37907, particularly, a signal line is required to keep power balance between the swing control system and the LS system which are arranged in parallel, and hence the circuit configuration is complicated (problem 4).
An object of the present invention is to provide a hydraulic drive system including a swing control system, which enables swing operation to be accelerated for shift to a steady state without causing a jerky feel at the start-up of swing alone and combined operation including swing, which can suppress change of the swing speed at shift from the swing-alone operation to the swing-combined operation and vice versa, which can avoid the swing speed from extremely reducing as compared with the speeds of one or more other actuators at the start-up of the swing-combined operation, thereby ensuring superior swing operability and swing independence, and which is free from problems resulted from providing a separate circuit, such as an increase in cost and space and complication of the circuit configuration.
(1) To achieve the above object, the present invention provides a hydraulic drive system comprising a hydraulic pump, a plurality of actuators, including a swing motor, which are driven by a hydraulic fluid delivered from the hydraulic pump, a plurality of directional control valves for controlling respective flow rates of the hydraulic fluid supplied from the hydraulic pump to the plurality of actuators, a plurality of pressure compensating valves for controlling respective differential pressures across the plurality of directional control valves, and pump control means for load sensing control to control a pump delivery rate such that a delivery pressure of the hydraulic pump is held a predetermined value higher than a maximum load pressure among the plurality of actuators, wherein the hydraulic drive system further comprises first means provided respectively in those of the plurality of pressure compensating valves, which are not for a swing section associated with the swing motor, and setting, as a target compensation differential pressure, a differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure among the plurality of actuators; second means provided in the pressure compensating valve for the swing section and setting a target compensation differential pressure of that pressure compensating valve; third means provided in at least one of the plurality of pressure compensating valves, which is for the swing section, and reducing the target compensation differential pressure set by the second means when a load pressure of the swing motor rises, thereby giving a load dependent characteristic to the pressure compensating valve for the swing section; and fourth means provided in the pressure compensating valve for the swing section and setting a lower limit of the target compensation differential pressure that is set by the second means and modified by the third means.
With the present invention thus constructed, since the third means is provided in the pressure compensating valve for the swing section to give it the load dependent characteristic, the pressure compensating valve for the swing section finely adjusts the flow rate passing the same depending on change in the load pressure of the swing motor at the swing start-up, whereby the swing motor is smoothly accelerated and shifted to the steady state.
Also, the second means for setting the target compensation differential pressure of the pressure compensating valve for the swing section may be means for setting, as the target compensation differential pressure, the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure among the plurality of actuators as with the first means. In this case, by providing the fourth means as set forth above, the fourth means functions as lower limit setting means for limiting both reduction in the target compensation differential pressure itself set by the second means and reduction in the target compensation differential pressure due to the load dependent characteristic given by the third means (see (2) below). With this function, when the target compensation differential pressure of the pressure compensating valve for the swing section is going to reduce upon the delivery rate of the hydraulic pump coming into the saturation state, or when the target compensation differential pressure of the pressure compensating valve for the swing section is going to reduce in accordance with the load dependent characteristic upon a rise of the load pressure of the hydraulic pump, or when both of the above phenomena occur at the same time, the fourth means limits the reduction of the target compensation differential pressure so that the hydraulic fluid is supplied to the swing motor with priority. As a result, change of the swing speed is suppressed at shift from the swing-alone operation to the swing-combined operation, and vice versa. Further, at the start-up of the swing-combined operation, the swing speed is prevented from being extremely slowed down as compared with the speed of another actuator, whereby superior swing operability and swing independence can be ensured.
The second means for setting the target compensation differential pressure of the pressure compensating valve for the swing section may be means for setting, as the target compensation differential pressure, a value not changed depending on the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure among the plurality of actuators. In this case, the fourth means functions as lower limit setting means for limiting reduction in the target compensation differential pressure due to the load dependent characteristic given by the third means (see (3) below). With this function, even when the delivery rate of the hydraulic pump comes into the saturation state, the target compensation differential pressure of the pressure compensating valve for the swing section is not reduced. Also, when the target compensation differential pressure of the pressure compensating valve for the swing section is going to reduce in accordance with the load dependent characteristic upon a rise of the load pressure of the hydraulic pump, the fourth means limits the reduction in the target compensation differential pressure. Thus, even when the reductions in the target compensation differential pressure due to the saturation and the load dependent characteristic occur solely or simultaneously, the hydraulic fluid is supplied to the swing motor with priority. As a result, change of the swing speed is suppressed at shift from the swing-alone operation to the swing-combined operation, and vice versa. Further, at the start-up of the swing-combined operation, the swing speed is prevented from being extremely slowed down as compared with the speed of another actuator, whereby superior swing operability and swing independence can be ensured.
Additionally, since the above-described functions are achieved without providing a separate circuit, such problems as an increase in cost and space and complication of the circuit configuration are avoided.
(2) In the above (1), preferably, the second means is means for setting, as the target compensation differential pressure, the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure among the plurality of actuators as with the first means, and the fourth means functions as lower limit setting means for limiting both reduction in the target compensation differential pressure itself set by the second means and reduction in the target compensation differential pressure due to the load dependent characteristic given by the third means.
With that feature, as set forth in the above (1), when the target compensation differential pressure of the pressure compensating valve for the swing section is going to reduce upon the delivery rate of the hydraulic pump coming into the saturation state, or when the target compensation differential pressure of the pressure compensating valve for the swing section is going to reduce in accordance with the load dependent characteristic upon a rise of the load pressure of the hydraulic pump, or when both of the above phenomena occur at the same time, the fourth means limits the reduction of the target compensation differential pressure so that the hydraulic fluid is supplied to the swing motor with priority, whereby superior swing operability and swing independence can be ensured.
(3) In the above (1), the second means may be means for setting, as the target compensation differential pressure, a value not changed depending on the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure among the plurality of actuators. In this case, the fourth means functions as lower limit setting means for limiting the reduction in the target compensation differential pressure due to the load dependent characteristic given by the third means.
With that feature, as set forth in the above (1), even when the delivery rate of the hydraulic pump comes into the saturation state, the target compensation differential pressure of the pressure compensating valve for the swing section is not reduced. Also, when the target compensation differential pressure of the pressure compensating valve for the swing section is going to reduce in accordance with the load dependent characteristic upon a rise of the load pressure of the hydraulic pump, the fourth means limits the reduction in the target compensation differential pressure. Thus, even when the reductions in the target compensation differential pressure due to the saturation and the load dependent characteristic occur solely or simultaneously, the hydraulic fluid is supplied to the swing motor with priority, whereby superior swing operability and swing independence can be ensured.
(4) In the above (1)-(3), preferably, the fourth means is biasing means for applying a biasing force to a spool of the pressure compensating valve for the swing section in the valve-opening direction when the target compensation differential pressure set by the second means and modified by the third means reaches a predetermined value.
With that feature, the fourth means prevents the target compensation differential pressure of the pressure compensating valve for the swing section from reducing down below a value corresponding to the biasing force applied by the biasing means.
(5) In the above (4), preferably, the biasing means is a lower limit setting spring acting on the spool of the pressure compensating valve for the swing section and biasing the spool in the valve-opening direction when the target compensation differential pressure set by the second means and modified by the third means reaches the predetermined value.
With that feature, the biasing means applies the biasing force to the spool of the pressure compensating valve for the swing section in the valve-opening direction when the target compensation differential pressure of the pressure compensating valve for the swing section reaches the predetermined value.
(6) In the above (1) and (2), preferably, the fourth means is biasing means for always adding a supplement value to the target compensation differential pressure that is set by the second means and modified by the third means, and the directional control valve for the swing section is constructed such that meter-in variable throttles thereof each have an opening area smaller than that in the directional control valves not for the swing section by an amount of the target compensation differential pressure corresponding to the supplement value added by the biasing means.
With that feature, the fourth means restricts the reduction in the target compensation differential pressure of the pressure compensating valve for the swing section by an amount corresponding to the supplement value added by the biasing means, thereby setting a lower limit of the target compensation differential pressure.
(7) In the above (6), preferably, the biasing means is a swing priority spring always acting on the spool of the pressure compensating valve for the swing section in the valve-opening direction.
With that feature, the fourth means always adds the supplement value to the target compensation differential pressure of the pressure compensating valve for the swing section.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a circuit diagram showing a hydraulic drive system according to a first embodiment of the present invention.
FIG. 2 is a sectional view showing details of the structure of a pressure compensating valve for a swing section.
FIG. 3 is a graph showing a load dependent characteristic of the pressure compensating valve for the swing section.
FIG. 4 is a graph showing a function of setting a lower limit of a target compensation differential pressure performed by a swing priority spring in the pressure compensating valve for the swing section.
FIG. 5 shows an appearance of a hydraulic excavator to which the hydraulic drive system of the present invention is applied.
FIG. 6 is a time chart showing change in the target compensation differential pressure of the pressure compensating valve for the swing section during the operation of swing alone.
FIG. 7 is a time chart for explaining the operation of the pressure compensating valve for the swing section when another actuator is started up during steady swing rotation and the degree of saturation is large, F in the figure indicating, for reference, the combined operation not including swing or the combined operation including swing with a spring 55 not provided.
FIG. 8 is a time chart for explaining the operation of the pressure compensating valve for the swing section when another actuator is started up during the steady swing rotation and the degree of saturation is small.
FIG. 9 is a time chart for explaining the operation of the pressure compensating valve for the swing section when the swing is started up simultaneously with another actuator and the degree of saturation is large, F in the figure indicating, for reference, the combined operation not including swing or-the combined operation including swing with the spring 55 not provided.
FIG. 10 is a time chart for explaining the operation of the pressure compensating valve for the swing section when the swing is started up simultaneously with another actuator and the degree of saturation is small.
FIG. 11 is a circuit diagram showing a hydraulic drive system according to a second embodiment of the present invention.
FIG. 12 is a graph showing an opening area characteristic of a directional control valve for a swing section.
FIG. 13 is a sectional view showing details of the structure of a pressure compensating valve for the swing section.
FIG. 14 is a graph showing a priority characteristic of a swing-section flow rate in a saturation state.
FIG. 15 is a circuit diagram showing a hydraulic drive system according to a third embodiment of the present invention.
FIG. 16 is a sectional view showing details of the structure of a pressure compensating valve for a swing section.
BEST MODE FOR CARRYING OUT THE INVENTION
Embodiments of the present invention will be described below 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 a hydraulic pump 1, a plurality of actuators 2-6, including a swing motor 2, which are driven by a hydraulic fluid delivered from the hydraulic pump 1, a plurality of closed-center directional control valves 7-11 for controlling respective flow rates of the hydraulic fluid supplied from the hydraulic pump 1 to the plurality of actuators 2-6, a plurality of pressure compensating valves 12-16 for controlling respective differential pressures across the plurality of directional control valves 7-11, load check valves 17a-17e disposed respectively between the directional control valves 7-11 and the pressure compensating valves 12-16 to prevent reverse flow of the hydraulic fluid, and a pump control delivery rate such that a delivery pressure of the hydraulic pump 1 is held a predetermined value higher than a maximum load pressure among the plurality of actuators 2-6. Overload relief valves 60a, 60b are provided in an actuator line for the swing motor 2. Though not shown, similar overload relief valves are provided in association with the other actuators 3-6.
The plurality of directional control valves 7-11 are provided with lines 20-24 respectively for detecting load pressures of themselves. A maximum one of load pressures detected with the detection lines 20-24 is extracted and introduced to a signal line 37 through signal lines 25-29, shuttle valves 30-33 and signal lines 34-36.
The pump control unit 18 comprises a tilting control actuator 40 coupled to a swash plate 1a which serves as a displacement varying member of the hydraulic pump 1, and a load sensing control valve (hereinafter referred to also as an LS control valve) for selectively controlling connection of a hydraulic pressure chamber 40a of the actuator 40 to a delivery fluid line 1b of the hydraulic pump 1 and a reservoir 19. The delivery pressure of the hydraulic pump 1 and the maximum load pressure in the signal line 37 act, as control pressures, on the LS control valve in opposite directions. When the pump delivery pressure rises beyond a total of the maximum load pressure and a setting value (target LS differential pressure) of a spring 41a, the hydraulic pressure chamber 40a of the actuator 40 is connected to the delivery fluid line 1b of the hydraulic pump 1 and a higher pressure is introduced to the hydraulic pressure chamber 40a, whereupon the piston 40b is moved to the left in FIG. 1 against the force of a spring 40c. Accordingly, the tilting of the swash plate 1a is decreased to reduce the delivery rate of the hydraulic pump 1. Conversely, when the pump delivery pressure lowers down below the total of the maximum load pressure and the setting value (target LS differential pressure) of the spring 41a, the hydraulic pressure chamber 40a of the actuator 40 is connected to the reservoir 19 and the hydraulic pressure chamber 40a is depressurized, whereupon the piston 40b is moved to the right in FIG. 1 by the force of the spring 40c. Accordingly, the tilting of the swash plate 1a is enlarged to increase the delivery rate of the hydraulic pump 1. With the above-described operation of the LS control valve, the delivery rate of the hydraulic pump 1 is controlled such that the pump delivery pressure is held higher than the maximum load pressure by an amount corresponding to the setting value (target LS differential pressure) of the spring 41a.
In the pressure compensating valves 12-16, pressures upstream of the directional control valves 7-11 act in the valve-closing direction, pressures (load pressures) in the detection lines 20-24 given by pressures downstream of the directional control valves 7-11 act in the valve-opening direction, the maximum load pressure introduced to the signal line 37 acts in the valve-closing direction, and the delivery pressure of the hydraulic pump 1 acts in the valve-opening direction. As a result, the differential pressures across the plurality of directional control valves 7-11 are controlled by employing, as the target compensation differential pressure, a differential pressure (hereinafter referred to also as an LS control differential pressure) between the delivery pressure of the hydraulic pump 1, which has been LS-controlled as described above, and the maximum load pressure.
Of the pressures acting on the pressure compensating valves 12-16, the pressures upstream of the directional control valves 7-11 are taken out respectively through signal lines 50a-50e, the pressures (load pressures) in the detection lines 20-24 given by the pressures downstream of the directional control valves 7-11 are taken out respectively through signal lines 51a-51e, the maximum load pressure in the signal line 37 is taken out through signal lines 52 and 52a-52e, and the delivery pressure of the hydraulic pump 1 is taken out through signal lines 53 and 53a-53e. In the pressure compensating valves 13-16, the maximum load pressure taken out through the signal lines 52b-52e is applied to fluid chamber 13a-16a, and the delivery pressure of the hydraulic pump 1 taken out through the signal lines 53b-53e is applied to fluid chamber 13b-16b, thereby setting the target compensation differential pressure. Fluid chambers of the pressure compensating valve 12, which are formed therein to set the target compensation differential pressure, will be described later.
Further, the pressure compensating valve 12 is constructed to have such a load dependent characteristic that when the load pressure of the swing motor 2 rises under a condition where the pressure upstream of the directional control valve 7 acts in the valve-closing direction and the pressure (the load pressure of the swing motor 2 ) in the detection line 20 given by the pressure downstream of the directional control valve 7 acts in the valve-opening direction, the target compensation differential pressure is reduced to restrict the flow rate of the hydraulic fluid passing the directional control valve 7. In addition, the pressure compensating valve 12 includes a lower-limit setting spring 55 provided on the side acting in the valve-opening direction, i.e., on the side setting the target compensation differential pressure. The lower-limit setting spring 55 acts on a spool of the pressure compensating valve 12 only when the target compensation differential pressure of the pressure compensating valves 13-16 for the other sections is reduced down below the setting value of the spring 55, thereby setting a lower limit to prevent the target compensation differential pressure from becoming smaller than the setting value.
The structure of the pressure compensating valve 12 is shown in FIG. 2.
Referring to FIG. 2, the pressure compensating valve 12 has two bodies 101, i.e., a first body 301a and a second body 301b. These bodies are assembled into an integral structure by appropriate means (not shown) such as bolting. In the first body 301a, there are formed a small-diameter bore 321 and a medium-diameter bore 322 in continuation to the small-diameter bore 321. A first spool 311 having a diameter d1 is slidably fitted in the small-diameter bore 321, and a second spool 312 having a diameter d3 (>d1) is slidably fitted in a medium-diameter bore 322. In the second body 301b, there are formed a large-diameter bore 323 in continuation to the medium-diameter bore 322 and a small-diameter bore 325 which is in continuation to the large-diameter bore 323 and has the same diameter as the small-diameter bore 321. A third spool 310 is slidably fitted in the large-diameter bore 323 and the small-diameter bore 325. The third spool 310 has first and second large-diameter portions 313, 314 which are slidably fitted in the large-diameter bore 323 and have a diameter d2 (>d3), and a small-diameter portion 315 which is slidably fitted in the small-diameter bore 325 and has the diameter d1.
A projection 321a is provided at an end surface of the small-diameter bore 321, and a fluid chamber 331 is formed around the projection 321a. A recess 311a for receiving the projection 321a is formed in an end surface of the first spool 311, and a weak initial-position holding spring 350 for pushing the spools in the valve-opening direction is disposed between an end surface of the projection 321a and a bottom portion of the recess 311a. Also, a chamber in which the spring 350 is disposed is communicated with the oil chamber 331, positioned on the outer side, through a passage 321b formed in the projection 321a.
The lower-limit setting spring 55 is disposed over the projection 321 in the oil chamber 331 and is positioned to face the end surface of the first spool 311. In the initial position as shown, the lower-limit setting spring 55 is positioned to face the end surface of the first spool 311, but away from the same, thereby generating no force to push the spools in the valve-closing direction.
Further, a pump port 341 and a load pressure port 342 are formed in the body 301a, while a reservoir port 343, an output port 344, an input port 345 and a maximum load pressure port 346 are formed in the body 301b. The pump port 341 is communicated with the signal line 53a for the delivery pressure of the hydraulic pump 1 and is opened to the fluid chamber 331. The load pressure port 342 is communicated with the load-pressure signal line 51a and is opened to a fluid chamber 332 which is formed in a connecting portion between the small-diameter bore 321 and the medium-diameter bore 322. Further, the reservoir port 343 is communicated with the reservoir 19 and is opened to a fluid chamber 333 formed in the large-diameter bore 323 which surrounds abutting ends of the second spool 312 and the third spool 310. The output port 344 is connected to the load check valve 17a and is opened to a fluid chamber 328 formed in the large-diameter bore 323 between the first and second large-diameter portions 313, 314 of the third spool. The input port 345 is communicated with the pump delivery fluid line 1b and is opened to the input side of a throttle portion 316 which is capable of opening/closing and formed in the second large-diameter portion 314 of the third spool 310. The maximum load pressure port 346 is communicated with the signal line 52a for the maximum load pressure and is opened to a fluid chamber 336 formed in the large-diameter bore 323 in which a continuously stepped portion between the second large-diameter portion 314 and the small-diameter portion 315 of the third spool 310.
Additionally, between the small-diameter portion 315 and an end surface 330 of the small-diameter bore, there is formed a fluid chamber 334 communicating with the fluid chamber 328, to which the output port 344 is opened, through a pilot fluid passage 50a formed within the third spool 310.
The body 301 is constructed by assembling the first body 301a and the second body 301b into an integral structure by appropriate means (not shown) such as bolting. At that time, even if the medium-diameter bore 322 on the side of the first body 301a and the large-diameter bore 323 on the side of the second body 301b are offset from each other, there is no problem in operation because the second spool 312 and the third spool 310 are formed as separate parts and held just in an abutting relation.
With the above construction, in the closing direction of the pressure compensating valve 12, the output pressure (Pz) at the output port 34 acts on a pressure bearing area B1 of the end surface 340 of the small-diameter portion 315 in the fluid chamber 334 through the pilot fluid passage 50a, and the maximum load pressure (PLmax) at the maximum load pressure port 346 acts on a pressure bearing area B2 of the stepped portion in the fluid chamber 336, which is resulted from subtracting the cross-sectional area of the small-diameter portion 315 from the cross-sectional area of the second large-diameter portion 314. Also, in the opening direction of the pressure compensating valve 12, the pump delivery pressure (Ps) acts on a pressure bearing area B1 of the end surface 340 of the first spool 311 in the fluid chamber 331 through the pump port 341, and the load pressure (PL) at the load pressure port 342 acts on a pressure bearing area B3 of the stepped portion in the fluid chamber 332, which is resulted from subtracting the cross-sectional area B1 of the first spool 311 from the cross-sectional area of the second spool 312. Moreover, no force acting to open and close the spools is imposed on a pressure bearing area of the stepped portion in the fluid chamber 333, which is resulted from subtracting the cross-sectional area of the second spool 312 from the cross-sectional area of the first large-diameter portion 313, because the fluid chamber 33 is communicated with the reservoir 19 through the reservoir port 343.
Then, the pressure bearing area B2 and the pressure bearing area B1 of the first spool 311 are set substantially equal to each other (B1=B2), and in addition the pressure bearing area B3 is set to be smaller than the pressure bearing area B1 (=B2) of the first spool (B1>B3), whereby the pressure compensating valve 12 is given a load dependent characteristic under which as the load pressure (PL) of the swing motor 2 increases, the flow rate passing the directional control valve 7 communicating with the swing motor 2 is reduced.
More specifically, considering balance among the hydraulic pressures imposed on the first spool 311, the second spool 312 and the third spool 313, the following formula holds because the pressure compensating valve 12 functions under a condition where B1P-B2PLmax is balanced by B1Pz-B3PL:
B1Ps-B2PLmax=B1Pz-B3PL
From B1=B2:
B1(Ps-PLmax)=B2Pz-B3PL
Ps-PLmax represents the differential pressure (LS control differential pressure) between the delivery pressure Ps of the hydraulic pump 1, which has been LS-controlled, and the maximum load pressure PLmax. Assuming the LS control differential pressure to be .DELTA.Pc, the following formula (1) is resulted:
B1.DELTA.Pc=B2Pz-B3PL (1)
Assuming the differential pressure across the directional control valve 7 to be .DELTA.P,
.DELTA.P=Pz-PL
is obtained. Also, the formula (1) can be modified into:
B1.DELTA.Pc+(B3-B2)PL=B2(Pz-PL)
Accordingly: ##EQU1##
Here, by putting B1/B2=.DELTA. and B3/B2=.beta.:
.DELTA.P=Pz-PL=.alpha..DELTA.Pc-(1-.beta.)PL (3)
Stated otherwise, if B2=B3 holds (there is no area difference between B2 and B3),
.DELTA.P=.alpha..DELTA.Pc
would be resulted and P would be determined only depending on .DELTA.Pc (LS control differential pressure). Because of B2.noteq.B3 (area difference between B2 and B3), .DELTA.P is affected by the load pressure PL depending on the area difference, thereby providing such a load dependent characteristic that as the load pressure PL increases, .DELTA.P is decreased to reduce the flow rate passing the directional control valve 7.
FIG. 3 shows the load dependent characteristic of the pressure compensating valve 12. The horizontal axis of FIG. 3 represents the load pressure denoted by PL, and the vertical axis represents the target compensation differential pressure denoted by .DELTA.Pv. A dotted line indicates, for reference, the target compensation differential pressure of the pressure compensating valves 13-16 for sections other than that for the swing (hereinafter referred to as a swing section). The pressure compensating valves 13-16 not for the swing section each have the target compensation differential pressure .DELTA.Pv that is held at the LS control differential pressure .DELTA.Pc in spite of an increase in the load pressures PL of the associated actuators 3-6. On the other hand, in the pressure compensating valve 12 for the swing section, when the load pressures PL increases, the target compensation differential pressure .DELTA.Pv is reduced depending on an increase in the load pressure PL.
FIG. 4 shows a function of setting a lower limit of the target compensation differential pressure effected by the lower limit setting spring 55 when it is assumed that the pressure compensating valve 12 is not given the load dependent characteristic. The horizontal axis of FIG. 4 represents, by Qr, a total of the flow rates demanded by the directional control valve 7 and the other directional control valves 8-11 (i.e., the valve demanded flow rates). This value corresponds to a total of input amounts by which levers of control lever units (not shown) for shifting the directional control valves 7-11 are operated, i.e., a total demanded flow rate of the swing motor 2 and the actuators. The vertical axis represents the target compensation differential pressure .DELTA.Pv set for the pressure compensating valve 12 and the other pressure compensating valves 13-16. Also, a differential pressure set by the lower limit setting spring 55 (i.e., a lower limit of the target compensation differential pressure) is denoted by Pb.
During the swing-combined operation in which the swing motor 2 and the other actuators are driven simultaneously, when the total Qr of the valve demanded flow rates of the directional control valve 7 and the other directional control valves 8-11 is smaller than a maximum delivery rate Qpmax of the hydraulic pump 1 and hence the delivery rate of the hydraulic pump 1 is not in the saturation state, the target compensation differential pressure .DELTA.Pv of all the pressure compensating valves, including the pressure compensating valve 12, is constant at the LS control differential pressure .DELTA.Pc.
When the total Qr of the valve demanded flow rates exceeds the maximum delivery rate Qpmax of the hydraulic pump 1 and hence the delivery rate of the hydraulic pump 1 is brought into the saturation state, the target compensation differential pressure .DELTA.Pv of all the pressure compensating valves is reduced with a lowering of the LS control differential pressure .DELTA.Pc until the LS control differential pressure .DELTA.Pc lowers down to the differential pressure Pb set by the lower limit setting spring 55 in the pressure compensating valve 12 for the swing section. When the LS control differential pressure .DELTA.Pc lowers down to the differential pressure Pb set by the lower limit setting spring 55, the target compensation differential pressure .DELTA.Pv of the pressure compensating valve 12 is held thereafter at the differential pressure Pb set by the lower limit setting spring 55 and is no more reduced beyond the lower limit, whereas the target compensation differential pressure .DELTA.Pv of the pressure compensating valves not for the swing section continues reducing with a lowering of the LS control differential pressure .DELTA.Pc.
In FIG. 4, a thick broken line indicates change in the target compensation differential pressure .DELTA.Pv of the pressure compensating valves 13-16 not for the swing section during the combined operation including the swing section, and a thin broken line indicates change in the target compensation differential pressure .DELTA.Pv of the pressure compensating valves 13-16 during the combined operation not including the swing section. Since the target compensation differential pressure .DELTA.Pv of the pressure compensating valve 12 for the swing section is not reduced down below the differential pressure Pb set by the lower limit setting spring 55, the target compensation differential pressure .DELTA.Pv of the pressure compensating valves 13-16 not for the swing section during the combined operation including the swing section is reduced at a greater rate than the target compensation differential pressure .DELTA.Pv of the pressure compensating valves 13-16 during the combined operation not including the swing section.
The hydraulic drive system described above is installed, for example, in a hydraulic excavator. FIG. 5 shows an appearance of the hydraulic excavator. Referring to FIG. 5, the hydraulic excavator comprises a lower track structure 200, an upper swing structure 201, and a front operating mechanism 202. The upper swing structure 201 is able to swing on the lower track structure 200 about an axis O, and the front operating mechanism 202 is able to move vertically in front of the upper swing structure 201. The front operating mechanism 202 has a multi-articulated structure comprising a boom 203, an arm 204 and a bucket 205. The boom 203, the arm 204 and the bucket 205 are driven respectively by a boom cylinder 206, an arm cylinder 207 and a bucket cylinder 208 for rotation in a plane that contains the axis O. The swing motor 2 shown in FIG. 1 is an actuator for driving the upper swing structure 202 to swing on the lower track structure 200. Three of the other actuators 3-6 are employed as the boom cylinder 206, the arm cylinder 207 and the bucket cylinder 208.
In the above construction, the fluid chambers 13a-16a, 13b-16b communicating with the signal lines 52b-52e, 53b-53e of the pressure compensating valves 13-16 constitute first means provided respectively in those 13-16 of the plurality of pressure compensating valves 12-16, which are not for the swing section associated with the swing motor 2, and setting, as the target compensation differential pressure, the differential pressure between the delivery pressure of the hydraulic pump 1 and the maximum load pressure among the plurality of actuators 2-6. The fluid chamber 336 (having the pressure bearing area B2=B1) and the fluid chamber 331 (having the pressure bearing area B1) communicating respectively with the signal lines 52a, 53a of the pressure compensating valve 12 constitute second means provided in the pressure compensating valve 12 for the swing section and setting the target compensation differential pressure of the pressure compensating valve 12. The fluid chamber 334 (having the pressure bearing area B1>B3) and the fluid chamber 332 (having the pressure bearing area B3) communicating respectively with the signal lines 50a, 51a of the pressure compensating valve 12 constitute third means provided in at least one 12 of the plurality of pressure compensating valves 12-16, which is for the swing section, and reducing the target compensation differential pressure set by the second means when the load pressure of the swing motor 2 rises, thereby giving a load dependent characteristic to the pressure compensating valve 12 for the swing section. The lower limit setting spring 55 in the pressure compensating valve 12 constitutes fourth means provided in the pressure compensating valve 12 for the swing section and setting a lower limit of the target compensation differential pressure that is set by the second means and modified by the third means.
Further, in this embodiment, the second means (the fluid chambers 331, 336) is means for setting, as the target compensation differential pressure, the differential pressure between the delivery pressure of the hydraulic pump 1 and the maximum load pressure among the plurality of actuators 2-6 as with the first means (the fluid chambers 13a-16a, 13b-16b). The fourth means (the lower limit setting spring 55) functions as lower limit setting means for limiting both reduction in the target compensation differential pressure itself set by the second means (the fluid chambers 331, 336) and reduction in the target compensation differential pressure due to the load dependent characteristic given by the third means (the fluid chambers 332, 334).
Additionally, the fourth means (the lower limit setting spring 55) is biasing means for applying a biasing force to the spool 311 of the pressure compensating valve 12 for the swing section in the valve-opening direction when the target compensation differential pressure set by the second means (the fluid chambers 331, 336) and modified by the third means (the fluid chambers 332, 334) reaches a predetermined value.
The operation of this embodiment thus constructed will be described.
1. Operation of Swing Alone
FIG. 6 is a time chart showing the behavior of the swing-associated pressure compensating valve 12 during the operation of swing alone in which the swing-associated directional control valve 7 is operated and the swing motor 2 is driven solely.
At the start-up of the swing-alone operation, there occurs a rise of the load pressure of the upper swing structure 201 specific to an inertial load. Such a rise of the load pressure is restricted by a safety valve that is constructed by the overload relief valve 60a or 60b disposed in association with the swing motor 2. In this condition, the hydraulic fluid supplied to the swing motor 2 is drained to the reservoir through the safety valve 60a or 60b.
In a conventional general pressure compensating valve, an acceleration feel of the upper swing structure 201, which is an inertial load, has been adjusted with the drain of the hydraulic fluid through the safety valve. In this case, however, since a flow rate of the hydraulic fluid drawn by the swing motor at the start-up is small, most of the hydraulic fluid is drained to the reservoir, thus resulting in an energy loss. Also, it is difficult to keep balance between the LS control of the hydraulic pump and the flow rate compensating function of the pressure compensating valve, causing the operator to feel jerky in the swing operation.
By contrast, this embodiment is free from such a problem because the pressure compensating valve 12 for the swing section has the load dependent characteristic described above.
First, in a condition prior to the start-up where the control lever of the swing-associated control lever unit is not operated, the target compensation differential pressure .DELTA.Pv of the pressure compensating valve 12 is controlled to the LS control differential pressure .DELTA.Pc (t0-t1).
Then, when the swing motor 2 is started up by operating the control lever, the load pressure PL rises due to the inertial load at the same time as the start-up (t1).
With the load dependent characteristic of the pressure compensating valve 12, the target compensation differential pressure .DELTA.Pv is reduced down from the LS control differential pressure .DELTA.Pc until reaching the differential pressure Pb set by the lower limit setting spring 55 (t1). A supply flow rate Qa to the swing motor 2 is controlled to a value corresponding to the differential pressure Pb set by the spring 55. In the case of not including the lower limit setting spring 55, the target compensation differential pressure .DELTA.Pv is further reduced down below Pb (but will not become zero).
When the upper swing structure 201 starts rotation and the swing speed rises, the flow rate of the hydraulic fluid drawn by the swing motor 2 is balanced by the supply flow rate Qa to the swing motor 2 and the load pressure lowers gradually. As a result, the target compensation differential pressure .DELTA.Pv of the pressure compensating valve 12 increases gradually (t2).
When the flow rate of the hydraulic fluid drawn by the swing motor 2 is not balanced by the supply flow rate Qa to the swing motor 2, this condition is fed back, as a rise or fall of the load pressure PL, to the pressure compensating valve 12 for the swing section. With the load dependent characteristic of the pressure compensating valve 12, when the supply flow rate Qa is too large, the load pressure PL increases and therefore the supply flow rate Qa is restricted by the pressure compensating valve 12. Conversely, when the supply flow rate Qa is insufficient, the load pressure PL decreases and therefore the supply flow rate Qa is increased by the pressure compensating valve 12. Such fine adjustment of the pressure compensating valve 12 enables the swing motor 2 to be moderately accelerated without causing hunting that has been generated in the conventional LS control.
At the time when the supply flow rate reaches an intrinsic value, the swing motor comes into a steady state (t3) and the load pressure PL is given by a pressure due to the rotation resistance.
2. Start-up of Another Actuator during Steady Rotation in Swing
FIG. 7 is a time chart showing the behavior of the pressure compensating valves for the respective sections during the combined operation in which, during steady rotation in the swing-alone operation, another actuator, e.g., the boom cylinder, is started up. It is assumed that the actuator 3 serves as the boom cylinder.
During steady rotation in the swing-alone operation, the l |