Information
-
Patent Grant
-
6584770
-
Patent Number
6,584,770
-
Date Filed
Wednesday, September 12, 200123 years ago
-
Date Issued
Tuesday, July 1, 200321 years ago
-
Inventors
-
Original Assignees
-
Examiners
- Look; Edward K.
- Kershteyn; Igor
Agents
- Mattingly, Stanger, Malur, P.C.
-
CPC
-
US Classifications
Field of Search
US
- 060 452
- 060 422
- 091 446
- 091 447
-
International Classifications
-
Abstract
In a hydraulic drive system in which a target compensated differential pressure for each of pressure compensating valves 21a, 21b is set in accordance with a differential pressure between a pump delivery pressure and a maximum load pressure, and a target LS differential pressure is set as a variable value depending on a revolution speed of an engine 1, a fixed throttle 32 and a signal pressure variable relief valve 33 are disposed in a maximum load pressure line 35. A relief setting pressure PLMAXO of the signal pressure variable relief valve 33 is set so as to satisfy PLMAXO=PR−PGR+a (where a is a value smaller than PGR) with respect to a target LS differential pressure PGR and a setting pressure PR of the main relief valve 30.
Description
TECHNICAL FIELD
The present invention relates to a hydraulic drive system for a construction machine, such as a hydraulic excavator, in which load sensing control is performed to hold a delivery pressure of a hydraulic pump higher than a maximum load pressure of a plurality of actuators by a target differential pressure, and in which differential pressures across a plurality of directional control valves are each controlled by a pressure compensating valve. More particularly, the present invention relates to a hydraulic drive system in which a target compensated differential pressure of each pressure compensating valve is set by a differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators, and the target differential pressure in the load sensing control is variably set depending on an engine revolution speed.
BACKGROUND ART
A hydraulic drive system, in which load sensing control is performed to hold a delivery pressure of a hydraulic pump higher than a maximum load pressure of a plurality of actuators by a target differential pressure, is called a load sensing system (hereinafter referred to also as an “LS system”). Usually, in the LS system, differential pressures across a plurality of directional control valves are each controlled by a pressure compensating valve so that a hydraulic fluid can be supplied to the actuators at a ratio depending on opening areas of the directional control valves regardless of the magnitude of load pressure during the combined operation in which the plurality of actuators are simultaneously driven.
In connection with such an LS system, JP,A 10-196604 discloses a hydraulic drive system in which a differential pressure (hereinafter referred to as an “LS differential pressure”) between a delivery pressure of a hydraulic pump and a maximum load pressure of a plurality of actuators is introduced to pressure compensating valves for setting a target compensated differential pressure of each pressure compensating valve by the LS differential pressure, and in which a target differential pressure (hereinafter referred to as a “target LS differential pressure”) in the load sensing control is variably set depending on an engine revolution speed.
By setting the target compensated differential pressure of each pressure compensating valve by the LS differential pressure, when a saturation state, where a delivery rate of the hydraulic pump is insufficient for satisfying a flow rate demanded by the plurality of directional control valves, occurs during the combined operation in which the plurality of actrators are simultaneously driven, the LS differential pressure is lowered depending on a degree of saturation, and the target compensated differential pressure of each pressure compensating valve is also reduced correspondingly. Therefore, the delivery rate of the hydraulic pump can be redistributed at a ratio of flow rates demanded by the respective actuators. Such a system is based on the concept of the invention disclosed in JP,A 60-11706.
By variably setting the target LS differential pressure depending on the engine revolution speed, when the engine revolution speed is lowered, the target LS differential pressure is also reduced correspondingly. Accordingly, even when a control lever for the directional control valve is operated in the same input amount as in the rated state, the flow rate of the hydraulic fluid supplied to the actuator is reduced and the actuator speed is slowed down. As a result, the actuator speed can be obtained corresponding to the engine revolution speed and fine operability can be improved.
Further, in connection with the LS system, GB2195745A discloses a system in which a signal pressure relief valve is disposed in a maximum load pressure line for detecting a maximum load pressure as a signal pressure, a setting pressure of the signal pressure relief valve is set to be lower than a setting pressure of a main relief valve, and the maximum load pressure having an upper limit restricted by the signal pressure relief valve is introduced to each pressure compensating valve. By providing the signal pressure relief valve in the maximum load pressure line, even when a load pressure of any one actuator reaches the setting pressure of the main relief valve and a delivery pressure of a hydraulic pump becomes equal to the maximum load pressure during the combined operation in which a plurality of actuators are simultaneously driven, it is possible to prevent all of the pressures compensating valves from being fully closed and hence prevent all of the actuators from being stopped, because the signal pressure in the maximum load pressure line is reduced to a level lower than the delivery pressure of the hydraulic pump.
DISCLOSURE OF THE INVENTION
However, the prior-art systems described above have problems as follows.
In the prior art disclosed in JP,A 10-196604, as described above, the LS differential pressure is introduced as the target compensated differential pressure to the pressure compensating valve. During the combined operation in which a plurality of actuators are simultaneously driven, therefore, when the load pressure of any one actuator reaches the setting pressure of the main relief valve and the delivery pressure of the hydraulic pump becomes equal to the maximum load pressure, the LS differential pressure is reduced to 0 and the pressure compensating valves are all fully closed. Consequently, no hydraulic fluid is supplied to the other actuators as well, of which load pressures do not yet reach the relief pressure, and the actuators are all stopped.
By providing the signal pressure relief valve, disclosed in GB2195745A, in the maximum load pressure line of the hydraulic drive system disclosed in JP,A 10-196604, even when the delivery pressure of the hydraulic pump becomes equal to the maximum load pressure as mentioned above, the signal pressure in the detection line is reduced to a level lower than the delivery pressure of the hydraulic pump. It is hence possible to prevent all of the pressure compensating valves from being fully closed and prevent all of the actuators from being stopped. Such an arrangement, however, causes another problem.
In the hydraulic drive system disclosed in JP,A 10-196604, the target LS differential pressure is variably set depending on the engine revolution speed. Therefore, the target LS differential pressure differs between when the engine revolution speed is set to a rated value and when the engine revolution speed is set to a lower value. The target LS differential pressure is smaller in the latter case than in the former case, and the actual LS differential pressure is also reduced correspondingly. Accordingly, if the setting pressure of the signal pressure relief valve is set to be lower than the setting pressure of the main relief valve by a value corresponding to the LS differential pressure during the rated rotation, the following problem occurs. During the rated rotation, the LS differential pressure resulting when the load pressure of the actuator is low and the main relief valve is not operated is equal to the differential pressure between the delivery pressure of the hydraulic pump and the signal pressure in the detection line resulting when the load pressure rises up to the setting pressure of the main relief valve, and hence the target compensated differential pressure of the pressure compensating valve is not changed. However, when the engine revolution speed is set to a lower value, the LS differential pressure is reduced to a level lower than that during the rated rotation as described above, while the differential pressure between the setting pressure of the signal pressure relief valve and the setting pressure of the main relief valve remains the same as the LS differential pressure during the rated rotation. Accordingly, the differential pressure between the delivery pressure of the hydraulic pump and the signal pressure in the detection line resulting when the load pressure rises up to the setting pressure of the main relief valve is larger than the LS differential pressure resulting when the load pressure of the actuator is low and the main relief valve is not operated, whereby the target compensated differential pressure introduced to the pressure compensating valve is increased. As a result, when the load pressure of any one actuator reaches the setting pressure of the main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the hydraulic fluid is supplied to the other actuators at a larger flow rate than so far, and the other actuators are sped up. Operability in the combined operation is hence remarkably impaired.
A first object of the present invention is to provide a hydraulic drive system wherein, even when a load pressure of any one actuator reaches a setting pressure of a main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the other actuators are not stopped and good operability in the combined operation is obtained.
A second object of the present invention is to provide a hydraulic drive system wherein, even when a load pressure of any one actuator reaches a setting pressure of a main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the other actuators are not sped up and good operability in the combined operation is obtained.
(1) To achieve the above first object, 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 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, pump control means for performing load sensing control to hold a delivery pressure of the hydraulic pump higher than a maximum load pressure of the plurality of actuators by a target differential pressure, and a main relief valve for Restricting an upper limit of the delivery pressure of the hydraulic pump, a target compensated differential pressure for each of the plurality of pressure compensating values being set in accordance with a differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators, a target differential pressure in the load sensing control being set as a variable value depending on a revolution speed of the engine, wherein the hydraulic drive system further comprises target compensated differential pressure modifying means for setting, as the target compensated differential pressure for each of the plurality of pressure compensating valves, a modification value different from the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators, when the delivery pressure of the hydraulic pump rises up to a setting pressure of the main relief Valve.
Thus, the target compensated differential pressure modifying means is provided to set, as the target compensated differential pressure, the modification value different from the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure, when the delivery pressure of the hydraulic pump rises up to the setting pressure of the main relief valve. Accordingly, even when the load pressure of any one actuator reaches the setting pressure of the main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the target compensated differential pressure is not reduced down to 0, the pressure compensating valves are not closed, and the hydraulic fluid can be supplied to the other actuators. As a result, the other actuators are not stopped and good operability in the combined operation is ensured.
(2) Also, to achieve the above second object, according to the present invention, the modification value in the above (1) is a variable value depending on the revolution speed of the engine.
With that feature, when the engine revolution speed is lowered and the target differential pressure in the load sensing control, which is set as the variable value depending on the engine revolution speed, is reduced, the modification value set as the target compensated differential pressure is also reduced correspondingly. Therefore, even when the load pressure of any one actuator reaches the setting pressure of the main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the target compensated differential pressure is avoided from increasing beyond the target differential pressure in the load sensing control, thus resulting in that the other actuators are not sped up and good operability in the combined operation is ensured.
(3) Further, to achieve the above second object, according to the present invention, the modification value in the above (1) is equal to or smaller than the target differential pressure in the load sensing control set as a variable value depending on the revolution speed of the engine.
With that feature, when the engine revolution speed is lowered and the target differential pressure in the load sensing control, which is set as the variable value depending on the engine revolution speed, is reduced, the modification value set as the target compensated differential pressure is also reduced correspondingly. Therefore, even when the load pressure of any one actuator reaches the setting pressure of the main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the target compensated differential pressure is avoided from increasing beyond the target differential pressure in the load sensing control, thus resulting in that the other actuators are not sped up and good operability in the combined operation is ensured.
(4) In the above (1), preferably, the target compensated differential pressure modifying means includes a signal pressure relief valve which is provided in a maximum load pressure line for detecting the maximum load pressure, and which reduces an upper limit of the maximum load pressure detected by the maximum load pressure line to be lower than the setting pressure of the main relief valve by the modification value.
With that feature, when the delivery pressure of the hydraulic pump rises up to the setting pressure of the main relief valve, the maximum load pressure detected as a signal pressure by the maximum load pressure line is reduced to be lower than the setting pressure of the main relief valve by the modification value. Accordingly, the modification value set as the target compensated differential pressure becomes different from the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators.
(5) Still further, to achieve the above second object, according to the present invention, the signal pressure relief valve in the above (4) is a variable relief valve, and assuming a relief setting pressure of the variable relief valve to be P
LMAX0
, the target differential pressure in the load sensing control to be P
GR
, and the setting pressure of the main relief valve to be P
R
, the relief setting pressure P
LMAX0
of the variable relief valve is set so as to satisfy:
P
LMAX0
=P
R
−P
GR
+α
(where α is a value smaller than P
GR
)
With that feature, the modification value set as the target compensated differential pressure by the target compensated differential pressure modifying means is provided by P
R
−P
LMAX0
=P
GR
−α, which has a value smaller than P
GR
(i.e., the target differential pressure in the load sensing control set as a variable value depending on the revolution speed of the engine). Accordingly, as mentioned in the above (3), even when the load-pressure of any one actuator reaches the setting pressure of the main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the target compensated differential pressure is avoided from increasing beyond the target differential pressure in the load sensing control, thus resulting in that the other actuators are not sped up and good operability in the combined operation is ensured.
Also, by setting the modification value set as the target compensated differential pressure to not P
GR
, but P
GR
−α that is smaller than P
GR
, it is possible to stably perform the load sensing control by the pump control means using a signal pressure corresponding to the same relief setting pressure P
LMAX0
, and to improve stability of the system.
(6) Still further, to achieve the above second object, according to the present invention, the target compensated differential pressure modifying means in the above (1) includes a selector valve for changing over the target compensated differential pressure from the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators to the target differential pressure in the load sensing control, immediately before the delivery pressure of the hydraulic pump rises up to the setting pressure of the main relief valve.
With that feature, when the delivery pressure of the hydraulic pump rises up to the setting pressure of the main relief valve, the target differential pressure in the load sensing control is set as the target compensated differential pressure (modification value). Accordingly, as mentioned in the above (3), even when the load pressure of any one actuator reaches the setting pressure of the main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the target compensated differential pressure is avoided from increasing beyond the target differential pressure in the load sensing control, thus resulting in that the other actuators are not sped up and good operability in the combined operation is ensured.
Also, by changing over the signal pressure using the selector valve, the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators can be employed in the load sensing control by the pump control means after the relief. It is hence possible to stably perform the load sensing control and to improve stability of the system.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1
is a hydraulic circuit diagram showing a hydraulic drive system according to a first embodiment of the present invention.
FIG. 2
is a graph showing override characteristics of a signal pressure variable relief valve.
FIG. 3
is a graph showing the relationship between an actual maximum load pressure and a pressure (signal pressure) in a signal pressure line controlled by the signal pressure variable relief valve.
FIG. 4
is a hydraulic circuit diagram showing Comparative Example 1.
FIG. 5
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure, a maximum load pressure, and a target compensated differential pressure resulting when the combined operation of boom raising and swirl is performed in Comparative Example 1.
FIG. 6
is a hydraulic circuit diagram showing Comparative Example 2.
FIG. 7
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure, a signal pressure, and a target compensated differential pressure resulting when the combined operation of boom raising and swing is performed in Comparative Example 2, and changes over time of the same status variables resulting when the combined operation of boom raising and swing is performed Comparative Example 3 at a rated engine revolution speed.
FIG. 8
is a hydraulic circuit diagram showing Comparative Example 3.
FIG. 9
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure, a signal pressure, and a target compensated differential pressure resulting when the combined operation of boom raising and swing is performed in Comparative Example 3 at an engine revolution speed set lower than the rated value.
FIG. 10
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure, a signal pressure, and a target compensated differential pressure resulting when the combined operation of boom raising and swing is performed in Comparative Example 1.
FIG. 11
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure, a signal pressure, and a target compensated differential pressure resulting when the combined operation of boom raising and swing is performed in a first embodiment of the present invention at an engine revolution speed set lower than the rated-value.
FIG. 12
is a hydraulic circuit diagram showing a hydraulic drive system according to a second embodiment of the present invention.
FIG. 13
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure, a signal pressure, and a target compensated differential pressure resulting when the combined operation of boom raising and swing in a second embodiment of the present invention at the rated engine revolution speed.
FIG. 14
is a chart showing changes over time of a boom stroke, a swing angular speed; a pump delivery pressure, a signal pressure, and a target compensated differential pressure resulting when the combined operation of boom raising and swing is performed in the second embodiment of the present invention at an engine revolution speed set lower than the rated value.
FIG. 15
is a hydraulic circuit diagram showing a hydraulic drive system according to a third embodiment of the present invention.
FIG. 16
is a hydraulic circuit diagram showing a hydraulic drive system according to a fourth embodiment of the present invention.
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 of this first embodiment comprises an engine
1
, a hydraulic source
2
, a valve apparatus
3
, a plurality of actuators
4
a
,
4
b
, . . . , and a target LS differential pressure generating circuit
5
.
The hydraulic source
2
includes a variable displacement hydraulic pump
10
and a fixed displacement pilot pump
11
, which are both driven by the engine
1
, and also includes an LS/horsepower control regulator
12
for controlling a tilting (displacement) of the hydraulic pump
10
. The LS/horsepower control regulator
12
comprises a horsepower control tilting actuator
12
a
for reducing the tilting of the hydraulic pump
10
when a delivery pressure of the hydraulic pump
10
increases, and an LS control valve
12
b
and an LS control tilting actuator
12
c
for performing load sensing control to hold the delivery pressure of the hydraulic pump
10
to be higher than a maximum load pressure of a plurality of actuators
4
a
,
4
b
, . . . by a target differential pressure.
The LS control valve
12
b
has a pressure receiving section
12
d
positioned on the side acting to reduce a pressure supplied to the actuator
12
c
for increasing the tilting of the hydraulic pump
10
, and a pressure receiving section
12
e
positioned on the side acting to increase a pressure supplied to the actuator
12
c
for reducing the tilting of the hydraulic pump
10
. A target differential pressure in the load sensing control, i.e., a target LS differential pressure, which is given as an output pressure of a pressure control valve
51
(described later) in the target LS differential pressure generating circuit
5
, is introduced to the pressure receiving section
12
d
, and an output pressure of a pressure control valve
34
(usually a differential pressure between the delivery pressure of the hydraulic pump
10
and the maximum load pressure, that is, an LS differential pressure), is introduced as a load-sensing control signal pressure to the pressure receiving section
12
e
. In
FIG. 1
, a mark * affixed to a line connected to a reservoir port of the LS control valve
12
b
means that the line is connected to a line, also denoted by a mark *, branched from an inlet reservoir line of the hydraulic pump
10
.
The valve apparatus
3
includes valve sections
3
a
,
3
b
, . . . corresponding respectively to the actuators
4
a
,
4
b
, . . . , and another valve section
3
p
. A plurality of closed center directional control valves
20
a
,
20
b
, . . . , a plurality of pressure compensating valves
21
a
,
21
b
, . . . , and shuttle valves
22
a
,
22
b
, . . . constituting a part of a maximum load pressure detecting circuit are disposed respectively in the valve sections
3
a
,
3
b
, . . . , whereas a main relief valve
30
, a variable unloading valve
31
, a fixed throttle
32
, a signal pressure variable relief valve
33
, and the aforesaid pressure control valve
34
are disposed in the valve section
3
p.
The directional control valves
20
a
,
20
b
, . . . are connected to a hydraulic fluid supply line
8
which is in turn connected to a delivery line
7
of the hydraulic pump
10
, and control respective flow rates and directions of the hydraulic fluid supplied to the actuators
4
a
,
4
b
, . . . from the hydraulic pump
2
. Also, the directional control valves
20
a
,
20
b
, . . . are provided with load ports
23
a
,
23
b
, . . . for taking out respective load pressures of the actuators
4
a
,
4
b
, . . . when the actuators are driven. The load pressures taken out by the load ports
23
a
,
23
b
, . . . are supplied to one input ports of the shuttle valves
22
a
,
22
b
, . . . , respectively. The shuttle valves
22
a
,
22
b
, . . . are connected in a tournament fashion so that the maximum load pressure is detected as a signal pressure by a maximum load pressure line
35
connected to an output port of the shuttle valve
22
a
of the final stage.
The pressure compensating valves
21
a
,
21
b
, . . . are disposed respectively upstream of the directional control valves
20
a
,
20
b
, . . . , and control differential pressures across meter-in throttles of the directional control valves
20
a
,
20
b
, . . . so as to be kept equal to each other. To that end, the pressure compensating valves
21
a
,
21
b
, . . . have respectively pressure receiving sections
25
a
,
25
b
, . . ;
26
a
,
26
b
, . . . operating in the opening direction, and pressure receiving sections
27
a
,
27
b
, . . . operating in the closing direction. The output pressure of the pressure control valve
34
(usually the LS differential pressure) is introduced to the pressure receiving sections
25
a
,
25
b
, . . . . The load pressures of the actuators
4
a
,
4
b
, . . . (pressures downstream of the meter-in throttles of the directional control valves
20
a
,
20
b
, . . . ) taken out by the load ports
23
a
,
23
b
, . . . of the directional control valves
20
a
,
20
b
, . . . are introduced to the pressure receiving section
26
a
,
26
b
, . . . . Pressures upstream of the meter-in throttles of the directional control valves
20
a
,
20
b
, . . . are introduced to the pressure receiving sections
27
a
,
27
b
, . . . , respectively. Then, in accordance with the output pressure of the pressure control valve
34
(usually the LS differential pressure) introduced to the pressure receiving sections
25
a
,
25
b
, . . . , the pressure compensating valves
21
a
,
21
b
, . . . set the introduced output pressure as a target compensated differential pressure, and control differential pressures across the directional control valves
20
a
,
20
b
, . . . so as to be kept equal to the target compensated differential pressure.
By constructing the pressure compensating valves
21
a
,
21
b
, . . . as described above, during the combined operation in which a plurality of actuators
4
a
,
4
b
, . . . are simultaneously driven, the hydraulic fluid can be supplied to the actuators at a ratio depending on opening areas of the meter-in throttles of the directional control valves
20
a
,
20
b
, . . . , regardless of the magnitudes of load pressures. Also, even when a saturation state, where a delivery rate of the hydraulic pump
10
is insufficient for satisfying a flow rate demanded by the directional control valves
20
a
,
20
b
, . . . , occurs during the combined operation, the LS differential pressure is lowered depending on a degree of saturation, and the target compensated differential pressure for each of the pressure compensating valves
21
a
,
21
b
, . . . is also reduced correspondingly. Therefore, the delivery rate of the hydraulic pump
10
can be redistributed at a ratio of flow rates demanded by the actuators
4
a
,
4
b, . . . .
The main relief valve
30
is connected to the hydraulic fluid supply line
8
, and restricts an upper limit of the delivery pressure of the hydraulic pump
10
. The main relief valve
30
has a spring
30
a
for setting a relief pressure.
The variable unloading valve
31
is also connected to the hydraulic fluid supply line
8
, and operates to limit the differential pressure between the delivery pressure of the hydraulic pump
10
and the maximum load pressure to a value slightly larger than the target LS differential pressure that is the output pressure of the pressure control valve
51
. To that end, the variable unloading valve
31
has pressure receiving sections
31
a
,
31
b
operating in the closing direction, a spring
31
c
operating in the closing direction, and a pressure receiving section
31
d
operating in the opening direction. The pressure (maximum load pressure) in the maximum load pressure line
35
and the target LS differential pressure given as the output pressure of the pressure control valve
51
are introduced respectively to the pressure receiving sections
31
a
,
31
b
, and the delivery pressure of the hydraulic pump
10
is introduced to the pressure receiving section
31
d.
The fixed throttle
32
and the signal pressure variable relief valve
33
function to modify the maximum load pressure detected by the maximum load pressure line
35
when the delivery pressure of the hydraulic pump
10
rises up to the setting pressure of the main relief valve
30
, so that the output pressure of the pressure control valve
34
will not become 0. The fixed throttle
32
is provided midway the maximum load pressure line
35
, and the signal pressure variable relief valve
33
is connected to a portion (hereinafter referred to as a “signal pressure line”)
35
a
of the maximum load pressure line
35
downstream of the fixed throttle
32
. The signal pressure variable relief valve
33
reduces an upper limit of the maximum load pressure detected by the signal pressure line
35
a
to a level lower than the setting pressure of the main relief valve
30
by a value resulting from subtracting an LS control adjustment value a (i.e., a value for ensuring controllability of the LS control valve
12
b
; described later) from the target LS differential pressure given as the output pressure of the pressure control valve
51
. To that end, the signal pressure variable relief valve
33
has a spring
33
a
operating in the closing direction as a relief pressure setting means, and a pressure receiving section
33
b
operating in the opening direction. The target LS differential pressure given as the output pressure of the pressure control valve
51
is introduced to the pressure receiving section
33
b
, and a setting pressure P
LMAX0
(described later) of the variable relief valve
33
is provided by a difference value between a setting value of the spring
33
a
and the target LS differential pressure. Also, the setting value of the spring
33
a
is set to a value greater than a pressure (setting pressure P
R
) corresponding to a setting value of the spring
30
a
of the main relief valve
30
by the aforesaid value α. With such an arrangement, when the maximum load pressure detected by the signal pressure line
35
a
rises up to a value resulting from subtracting the target LS differential pressure from the pressure (=setting pressure of the main relief valve
30
+α) corresponding to the setting value of the spring
33
a
, the signal pressure variable relief valve
33
is operated to prevent the detected maximum load pressure from rising further.
The pressure control valve
34
is a differential pressure generating valve for outputting, as an absolute pressure, a differential pressure between a pressure in the hydraulic fluid supply line
8
(the delivery pressure of the hydraulic pump
10
) and a pressure in the signal pressure line
35
a
(maximum load pressure). The pressure control valve
34
has a pressure receiving section
34
a
operating in the direction to increase the pressure, and pressure receiving sections
34
b
,
34
c
operating in the direction to reduce the pressure. The pressure in the hydraulic fluid supply line
8
is introduced to the pressure receiving section
34
a
, and the signal pressure in the signal pressure line
35
a
and an output pressure of the pressure control valve
34
itself are introduced respectively to the pressure receiving sections
34
b
,
34
c
. Under a balanced condition among those pressures, the pressure control valve
34
outputs, based on a pressure of the pilot pump
11
, a pressure equal to the differential pressure (LS differential pressure) between the pressure in the hydraulic fluid supply line
8
and the signal pressure in the signal pressure line
35
a
to a signal pressure line
36
. The output pressure of the pressure control valve
34
is supplied via signal pressure lines
36
a
,
36
b
to the pressure receiving section
12
e
of the LS control valve
12
b
and to the pressure receiving sections
25
a
,
25
b
, . . . of the pressure compensating valves
21
a
,
21
b
, . . . .
Incidentally, the arrangement for outputting, as an absolute pressure, the LS differential pressure by the pressure control valve
34
is proposed by the invention disclosed in JP,A 10-89304.
The target LS differential pressure generating circuit
5
comprises a flow rate detecting valve
50
and a pressure generating valve
51
. The flow rate detecting valve
50
has a throttle
50
a
which is disposed in a delivery line
9
of the pilot pump
11
. A relief valve
40
for specifying a base pressure of a pilot hydraulic source is connected to a portion
9
a
of the delivery line
9
downstream of the flow rate detecting valve
50
, and the line
9
a
is connected to, e.g., remote control valves (not shown) for generating pilot pressures to shift the directional control valves
20
a
,
20
b
, . . . . The line
9
a
is also connected to an input port of the pressure control valve
34
via a branched line
9
b
and serves as a hydraulic source of the pressure control valve
34
.
The flow rate detecting valve
50
detects a flow rate of the hydraulic fluid flowing through the delivery line
9
as change of a differential pressure across the throttle
50
a
, and the detected differential pressure is employed as the target LS differential pressure. Herein, the flow rate of the hydraulic fluid flowing through the delivery line
9
represents a delivery rate of the pilot pump
11
, and the delivery rate of the pilot pump
11
is changed depending on the revolution speed of the engine
1
. Thus, detecting the flow rate of the hydraulic fluid flowing through the delivery line
9
means detection of the revolution speed of the engine
1
. For example, as the revolution speed of the engine
1
lowers, the flow rate of the hydraulic fluid flowing through the delivery line
9
is reduced and hence the differential pressure across the throttle
50
a
is lowered.
The throttle
50
a
is constructed as a variable throttle having an opening area that varies continuously. The flow rate detecting valve
50
further comprises a pressure receiving section
50
b
operating in the opening direction, and a pressure receiving section
50
c
and a spring
50
d
both operating in the throttling direction. A pressure upstream of the variable throttle
50
a
is introduced to the pressure receiving section
50
b,
and a pressure downstream of the variable throttle
50
a
is introduced to the pressure receiving section
50
c.
An opening area of the variable throttle
50
a
is thereby changed depending on a differential pressure across itself. By thus constructing the flow rate detecting valve
50
and employing the differential pressure across the variable throttle
50
a
as the LS target differential pressure, a saturation phenomenon occurred depending on the engine revolution speed can be improved and good fine operability can be obtained even when the engine revolution speed is set to a low value. The foregoing point is described in detail in JP,A 10-196604.
The pressure generating valve
51
is a differential pressure generating valve for outputting, as an absolute pressure, the differential pressure across the variable throttle
50
a
. The pressure generating valve
51
has a pressure receiving section
51
a
operating in the direction to increase the pressure and pressure receiving sections
51
b
,
51
c
both operating in the direction to reduce the pressure. The pressure upstream of the variable throttle
50
a
is introduced to the pressure receiving section
51
a
, and the signal pressure downstream of the variable throttle
50
a
and an output pressure of the pressure generating valve
51
itself are introduced respectively to the pressure receiving sections
51
b
,
51
c
. Under a balanced condition among those pressures, the pressure generating valve
51
outputs, based on a pressure in the line
9
a
, a pressure equal to the differential pressure across the variable throttle
50
a
to a signal pressure line
53
. The output pressure of the pressure control valve
51
is supplied, as the LS target differential pressure, to the pressure receiving section
12
d
of the LS control valve
12
b
via a signal pressure line
53
a
, and the same output pressure is also supplied, via a signal pressure line
53
b
, to the pressure receiving section
31
b
of the variable unloading valve
31
and to the pressure receiving section
33
b
of the signal pressure variable relief valve.
Herein, the opening area of the variable throttle
50
a
is set, for example, so as to provide a desired LS target differential pressure of about 15 kgf/cm
2
when the engine
1
is rotated in the rated state.
FIG. 2
shows override characteristics of the signal pressure variable relief valve
33
. In
FIG. 2
, P
LMAX0
represents the setting pressure of the signal pressure variable relief valve
33
, P
R
represents the setting pressure of the main relief valve
30
, and P
GR
represents the target LS differential pressure that varies depending on the engine revolution speed.
The setting pressure P
LMAX0
of the signal pressure variable relief valve
33
is controlled so as to satisfy the following formula with respect to the target LS differential pressure P
GR
:
P
LMAX0
=P
R
−P
GR
+α
where α is an LS control adjustment value (described later)
Specifically, as the engine revolution speed lowers, the target LS differential pressure P
GR
is reduced and hence the setting pressure P
LMAX0
of the signal pressure variable relief valve
33
is increased correspondingly.
FIG. 3
shows the relationship between an actual maximum load pressure detected by the load pressure line
35
and the pressure (signal pressure) in the signal pressure line
35
a
resulting when the setting pressure P
LMAX0
of the signal pressure variable relief valve
33
is controlled as described above. In
FIG. 3
, P
LMAX
represents the actual maximum load pressure and P
LMAX
′ represents the signal pressure.
Until the actual maximum load pressure P
LMAX
reaches the same level as the setting pressure P
LMAX0
of the signal pressure variable relief valve
33
, the signal pressure variable relief valve
33
is not operated, thus resulting in P
LMAX
′=P
LMAX
. When the actual maximum load pressure P
LMAX
exceeds the setting pressure P
LMAX0
of the signal pressure variable relief valve
33
, the signal pressure variable relief valve
33
is operated, whereby the pressure P
LMAX
′ in the signal pressure line
35
a
does not rise further and reaches a uppermost limit (remains constant) at P
LMAX0
. Also, since P
LMAX0
increases as the engine revolution speed lowers, the uppermost limit signal pressure P
LMAX
′ is also increased.
Consequently, assuming that the delivery pressure of the hydraulic pump
10
is Ps and the target compensated differential pressure for each of the pressure compensating valves
21
a
,
21
b
, . . . is Pc, the target compensated differential pressure Pc, which is set by the pressure outputted from the pressure control valve
34
to the pressure receiving sections
25
a
,
25
b
, . . . of the pressure compensating valves
21
a
,
21
b
, . . . upon relief through the signal pressure variable relief valve
33
, is expressed by:
Pc=Ps−P
LMAX0
Because of Ps=P
R
′
Pc=P
GR
−α.
The operation of this embodiment having the above-described construction will be described below in comparison with Comparative Examples based on the prior art.
FIG. 4
shows Comparative Example 1 constructed by modifying the hydraulic drive system of this embodiment, show in
FIG. 1
, based on the prior art disclosed in JP,A 10-196604. In the construction of Comparative Example 1, the valve apparatus
3
shown in
FIG. 1
is replaced by a valve apparatus
301
; the fixed throttle
32
and the signal pressure variable relief valve
33
shown in
FIG. 1
are not provided in a valve section
301
p
of the valve apparatus
301
; and the maximum load pressure detected by the maximum load pressure line
35
is directly introduced to the pressure control valve
34
.
With the construction of Comparative Example 1, during the combined operation in which, for example, the actuators
4
a
,
4
b
are simultaneously driven, when the load pressure of one actuator reaches the setting pressure of the main relief valve
30
, no hydraulic fluid is supplied to the other actuator, of which load pressure does not yet reach the setting pressure of the main relief valve
30
. In other words, when the load pressure of any one actuator reaches the setting pressure of the main relief valve
30
during the combined operation, the actuators are all stopped.
FIG. 5
shows an example of the operation of Comparative Example 1.
FIG. 5
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure Ps, a maximum load pressure P
LMAX
, and a target compensated differential pressure Pc resulting when the combined operation of boom raising and swing, i.e., a typical excavation work of a hydraulic excavator, is performed with the actuator
4
a
serving as a swing motor of the hydraulic excavator and the actuator
4
b
serving as a boom cylinder of the hydraulic excavator.
In
FIG. 5
, when the boom cylinder
4
b
reaches the stroke end, both of the maximum load pressure P
LMAX
and the pump delivery pressure Ps rise up to the setting pressure of the main relief valve
30
. This results in Ps=P
LMAX
. Therefore, the output pressure Pc outputted as the target compensated differential pressure to the pressure compensating valves
21
a
,
21
b
from the pressure control valve
34
is provided by Pc(=Ps−P
LMAX
)=0(kgf/cm
2
), and only the differential pressures across the directional control valves
20
a
,
20
b
act upon the pressure receiving sections
26
a
,
27
a
;
26
b
,
27
b
of the pressure compensating valves
21
a
,
21
b.
If some hydraulic fluid flows through the directional control valves
20
a
,
20
b
in that condition, spools of the pressure compensating valves
21
a
,
21
b
are subjected to forces acting in the closing direction. On this occasion, there are flows of the hydraulic fluid as long as the pressure compensating valves
21
a
,
21
b
are opened. Hence, the pressure compensating valves
21
a
,
21
b
are continuously subjected to forces acting in the closing direction until they are fully closed. Therefore, the pressure compensating valves
21
a
,
21
b
are eventually fully closed. With the full closing of the pressure compensating valves
21
a
,
21
b
, the supply of the hydraulic fluid to the swing motor
4
a
is ceased and the swing angular speed is reduced down to 0.
Thus, when the boom cylinder
4
b
reaches the stroke end and the load pressure of the boom cylinder
4
b
rises up to the setting pressure of the main relief valve
30
during the combined operation of boom raising and swing, the swing is stopped and the operability is remarkably impaired.
As means for solving the drawback mentioned above, it is conceivable, as disclosed in GB2195745A, to provide a signal pressure relief valve for setting an upper limit of P
LMAX
as a signal pressure, and to set the setting pressure of the signal pressure relief valve to be lower than the setting pressure of the main relief valve
30
so that Ps=P
LMAX
is not resulted upon relief through the main relief valve
30
.
Such a construction is shown as Comparative Example 2 in FIG.
6
. Comparative Example 2 differs from the hydraulic drive system of this embodiment shown in
FIG. 1
as follows. The target LS differential pressure generating circuit
5
is removed, and instead of the LS control valve
12
b
shown in
FIG. 1
, an LS control valve
112
b
having a spring
112
d
for setting the LS target value as a constant value is provided in an LS/horsepower control regulator
112
of a hydraulic source
102
. Further, the valve apparatus
3
shown in
FIG. 1
is replaced by a valve apparatus
302
, and instead of the variable unloading valve
31
and the signal pressure variable relief valve
33
shown in
FIG. 1
, a variable unloading valve
131
and a signal pressure relief valve
133
having setting pressures fixedly set by springs
131
c
,
133
a
, respectively, are provided in a valve section
302
p
of the valve apparatus
302
.
By providing the signal pressure relief valve
133
in the maximum load pressure line
35
through the fixed throttle
32
and introducing a pressure P
LMAX
′ in the signal pressure line
35
a
, which has been controlled by the signal pressure relief valve
133
, to the pressure control valve
34
, the pressure P
LMAX
′ lower than the setting pressure of the main relief valve
30
is introduced as a signal pressure to the pressure control valve
34
upon relief through the main relief valve
30
.
FIG. 7
is a chart showing changes over time of a boom stroke, a swing angular speed, a pump delivery pressure Ps, a pressure (signal pressure) P
LMAX
′ in the signal pressure line
35
a
, and a target compensated differential pressure Pc resulting when the combined operation of boom raising and swing is performed in Comparative Example 2.
In
FIG. 7
, when the boom cylinder
4
b
reaches the stroke end, both of the maximum load pressure P
LMAX
and the pump delivery pressure Ps rise up to the setting pressure of the main relief valve
30
. At this time, the pressure P
LMAX
′ in the signal pressure line
35
a
controlled by the signal pressure relief valve
133
is limited to a level lower than the setting pressure of the main relief valve
30
. Therefore, the output pressure Pc (=Ps−P
LMAX
′) outputted as the target compensated differential pressure to the pressure compensating valves
21
a
,
21
b
from the pressure control valve
34
is not reduced down to 0, but given by the differential pressure between the setting pressure of the main relief valve
30
and the setting pressure of the signal pressure relief valve
133
.
Herein, by setting the setting pressure P
LMAX0
of the signal pressure relief valve
133
as defined in the following formula, the target compensated differential pressure is not changed between during the boom operation before the main relief valve
30
is operated and when the main relief valve
30
is operated:
P
LMAX
=main relief setting value−target
LS
differential pressure
Consequently, even when the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, the swing is not stopped and the operability in the combined operation is maintained.
However, if the above-mentioned solving means is directly applied to the hydraulic drive system disclosed in JP,A 10-196604, another drawbacks occurs.
Such a construction is shown as Comparative Example 3 in FIG.
8
. Comparative Example 3 is constructed by modifying the hydraulic drive system of this embodiment, shown in
FIG. 1
, based on the concept of the prior art disclosed in GB2195745A. The valve apparatus
3
shown in
FIG. 1
is replaced by a valve apparatus
303
, and instead of the signal pressure variable relief valve
33
shown in
FIG. 1
, a signal pressure relief valve
133
having a setting pressure fixedly set by a springs
133
a
is provided in a valve section
303
p
of the valve apparatus
303
. Note that Comparative Example 3 represents the basic concept of the embodiment shown in FIG.
1
and constitutes a part of the present invention.
The signal pressure relief valve
133
operates in the same manner as in Comparative Example 2. Additionally, in Comparative Example 3, the target LS differential pressure is varied depending on the engine revolution speed. The setting pressure of the spring
133
a
of the signal pressure relief valve
133
is set lower than the setting pressure of the main relief valve
30
by an amount corresponding to the target LS differential pressure resulting when the engine revolution speed is set to the rated value.
The operation of Comparative Example 3 at the engine revolution speed set to the rated value is the same as in Comparative Example 2. Hence, as shown in
FIG. 7
, even when the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief during the combined operation of boom raising and swing, the swing angular speed is not reduced and the operability in the combined operation is maintained.
On the other hand, when the engine revolution speed is set to a level lower than the rated value, the target LS differential pressure is lowered in Comparative Example 3 so that the speeds of the actuators
4
a
,
4
b
are reduced with respect to the same input amounts from control levers of the directional control valves
20
a
,
20
b
, . . . as in the rated state.
FIG. 9
is a chart showing changes over time of the same status variables as shown in
FIG. 7
resulting when the combined operation of boom raising and swing is performed in Comparative Example 3 at an engine revolution speed set lower than the rated value.
Referring to
FIG. 9
, in the boom-raising operation before the main relief valve
30
is operated for relief, the pump delivery pressure Ps is held higher than the maximum load pressure P
LMAX
(=P
LMAX
′) by the target LS differential pressure. Since the target LS differential pressure in this case is lower than that resulting when the engine revolution speed is set to the rated value, the differential pressure Ps−P
LMAX
between the pump delivery pressure and the engine revolution speed, i.e., the target compensated differential pressure Pc of the pressure compensating valves
21
a
,
21
b
set by the output pressure of the pressure control valve
34
, is maintained to a level lower than when the engine revolution speed is set to the rated value.
When the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, the pressure P
LMAX
′ in the signal pressure line
35
a
is limited by the signal pressure relief valve
133
to a level lower than the maximum load pressure P
LMAX
. In this case, the difference between the pump delivery pressure Ps and the signal pressure P
LMAX
′ is given as the target LS differential pressure at the rated engine revolution speed, the target compensated differential pressure Pc of the pressure compensating valves
21
a
,
21
b
set by the output pressure of the pressure control valve
34
is increased from a level during the boom operation before the relief.
Consequently, the angular speed of the swing in the combined operation with a boom is increased at the same time as when the boom cylinder
4
b
reaches the stroke end. As a result, the operability in the combined operation is remarkably impaired.
In this embodiment, as described above, the signal pressure relief valve
33
is constructed as a variable relief valve, and the setting value of the variable relief valve is varied depending on the target LS differential pressure that changes with the engine revolution speed. The above-mentioned drawback can be overcome with such an arrangement.
The operation of the system of this embodiment in the combined operation of boom raising and swing, for example, will be described below as with Comparative Examples.
FIG. 10
is a chart showing changes over time of the same status variables as shown in
FIG. 7
resulting when the combined operation of boom raising and swing is performed in the system of this embodiment at an engine revolution speed set to the rated value.
FIG. 11
is a chart showing changes over time of the same status variables as shown in
FIG. 7
resulting when the combined operation of boom raising and swing is performed in the system of this embodiment at an engine revolution speed set lower than the rated value.
Referring to
FIG. 10
, in the boom-raising operation before the main relief valve
30
is operated for relief, the signal pressure variable relief valve
33
is not operated and the maximum load pressure P
LMAX
is directly detected as the signal pressure P
LMAX
′ by the signal pressure line
35
a
. Also, the pump delivery pressure Ps is held higher than the maximum load pressure P
LMAX
(=P
LMAX
′) by the target LS differential pressure P
GR
. Therefore, the target compensated differential pressure Pc of the pressure compensating valves
21
a
,
21
b
set by the output pressure of the pressure control valve
34
is equal to the differential pressure Ps−P
LMAX
between the pump delivery pressure and the engine revolution speed, i.e., the target LS differential pressure P
GR
, (Pc=P
GR
).
When the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, both of the maximum load pressure P
LMAX
and the pump delivery pressure Ps rise up to the setting pressure P
R
of the main relief valve
30
. At this time, the setting pressure P
LMAX0
of the signal pressure variable relief valve
33
is controlled so as to satisfy P
LMAX0
=P
R
−P
GR
+a with respect to the target LS differential pressure P
GR
, and the pressure P
LMAX
′ in the signal pressure line
35
a
controlled by the signal pressure variable relief valve
33
is limited to P
LMAX
′=P
R
−P
GR
+a that is lower than the setting pressure P
R
of the main relief valve
30
. Therefore, the output pressure Pc (=Ps−P
LMAX
′) outputted as the target compensated differential pressure to the pressure compensating valves
21
a
,
21
b
from the pressure control valve
34
is not reduced down to 0, but given by the differential pressure between the setting pressure of the main relief valve
30
and the setting pressure of the signal pressure variable relief valve
33
, i.e., Pc=P
GR
−a.
As a result, even when the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, the swing is not stopped and the operability in the combined operation is maintained.
The system of this embodiment operates likewise also when the engine revolution speed is set to a level lower than the rated value. More specifically, referring to
FIG. 11
, in the boom-raising operation before the main relief valve
30
is operated for relief, the target compensated differential pressure Pc of the pressure compensating valves
21
a
,
21
b
is equal to the target LS differential pressure P
GR
(Pc=P
GR
). When the boom cylinder
4
b
reaches the stroke end, the target compensated differential pressure Pc (=Ps−P
LMAX
′) of the pressure compensating valves
21
a
,
21
b
is not reduced down to 0, but given by the differential pressure between the setting pressure of the main relief valve
30
and the setting pressure of the signal pressure variable relief valve
33
(Pc =P
GR
−a) . In this case, however, since the target LS differential pressure P
GR
is lower than that when the engine revolution speed is set to the rated value, the target compensated differential pressure Pc of the pressure compensating valves
21
a
,
21
b
is maintained at a level lower than when the engine revolution speed is set to the rated value.
As a result, even when the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, the swing is not stopped and the operability in the combined operation is maintained with no increase of the swing angular speed.
Furthermore, in this embodiment, the setting pressure P
LMAX0
of the signal pressure variable relief valve
33
is set to P
LMAX0
=P
R
−P
GR
+α, instead of P
LMAX0
=P
R
−P
GR
, with respect to the target LS differential pressure P
GR
. The advantage resulting from such setting will be described below.
The output pressure Pc of the pressure control valve
34
is also supplied as the load-sensing control signal pressure to the LS control valve
12
b
of the LS/horsepower control regulator
12
. To the LS control valve
12
b
, there are introduced the target LS differential pressure P
GR
in the direction to increase the delivery rate of the hydraulic pump
10
and the load-sensing control signal pressure Pc in the direction to reduce the delivery rate of the hydraulic pump
10
. By setting of Pc=P
GR
−α, therefore, the pump delivery rate is controlled so as to maximize within the range of horsepower control flow rate provided by the horsepower control tilting actuator
12
a
upon relief through the main relief valve
30
.
Assuming a=0, for example, the LS control valve
12
b
is subjected to the same signal pressure at the pressure receiving sections
12
d,
12
e
at both ends thereof, and therefore loses controllability. This results in unstable operation of the LS control valve
12
b
under effects caused by variations in the setting pressure of the main relief valve
30
and the setting pressure of the signal pressure variable relief valve
33
.
For the reason mentioned above, setting the LS control adjustment value α ensures the stable operation of the system.
By the setting of α, however, the target compensated differential pressure Pc outputted from the pressure control valve
34
upon relief through the main relief valve
30
becomes lower than that during the operation before the relief by α (Pc=P
GR
→Pc=P
GR
−α), and the speed of the other actuator operated in the combined operation is lowered (see FIGS.
10
and
11
). Taking into account the above problem, in practice, a is set to be in a range in which the operator does not feel noticeably speed change during the operation. By way of example, a can be set as given below:
α=Pc
0
×0.14
where Pc
0
is the target LS differential pressure at the rated engine revolution speed.
With this embodiment, as described above, even when the load pressure of any one actuator reaches the setting pressure of the main relief valve
30
during the combined operation in which a plurality of actuators
4
a
,
4
b
, . . . are simultaneously driven, the other actuators are neither stopped nor sped up, and good operability in the combined operation is maintained.
A second embodiment of the present invention will be described with reference to
FIGS. 12
to
14
. In these drawings, identical members to those shown in
FIG. 1
are denoted by the same reference numerals.
Referring to
FIG. 12
, a hydraulic drive system of this embodiment includes a valve apparatus
3
A. In a valve section
3
Ap of the valve apparatus
3
A, the fixed throttle
32
and the signal pressure variable relief valve
33
shown in
FIG. 1
are not provided, and the maximum load pressure detected by the maximum load pressure line
35
is directly introduced to the pressure control valve
34
. Further, the system of this embodiment includes a selector valve
60
capable of selecting one of the output pressure of the pressure control valve
34
and the output pressure of the pressure control valve
51
, i.e., the target LS differential pressure. An output pressure of the selector valve
60
is introduced to the pressure receiving sections
25
a
,
25
b
, . . . of the pressure compensating valves
21
a
,
21
b
, . . . for setting the target compensated differential pressure.
The selector valve
60
has two input ports
60
a
,
60
b
and one output port
60
c
. The output pressure of the pressure control valve
34
is introduced to the input port
60
a
via the signal pressure line
36
and a signal pressure line
36
c
branched from it. The output pressure of the pressure control valve
51
, i.e., the target LS differential pressure, is introduced to the input port
60
b
via the signal pressure line
53
b
and a signal pressure line
53
c
branched from it. The output port
60
c
is connected to the pressure receiving sections
25
a
,
25
b
, . . . of the pressure compensating valves
21
a
,
21
b
, . . . via a signal pressure line
61
so that the output pressure of the selector valve
60
is introduced to the pressure receiving sections
25
a
,
25
b, . . . .
Also, the selector valve
60
has a spring
60
d
operating in the direction to select the first input port
60
a
, and pressure receiving sections
60
e
,
60
f
operating in the direction to select the second input port
60
b
. The maximum load pressure is introduced to the pressure receiving section
60
e
via the maximum load pressure line
35
and a signal pressure line
35
b
branched from it. The output pressure of the pressure control valve
51
, i.e., the target LS differential pressure, is introduced to the pressure receiving section
60
f
via a signal pressure line
53
d
branched from the signal pressure line
53
c
. The spring
60
d
is set to have the strength that provides the same value in terms of pressure as the setting pressure of the main relief valve
30
, i.e., the same strength as the spring
30
a
of the main relief valve
30
.
Further, the selector valve
60
has variable throttles
60
g
,
60
h
for varying pressure in a smooth and continuous manner when the selector valve
60
is shifted from a position where the pressure at the first input port
60
a
is selected as shown, to a position where the pressure at the second input port
60
b
is selected.
FIG. 13
is a chart showing changes over time of the same status variables as shown in
FIG. 10
resulting when the combined operation of boom raising and swing is performed in the system of this embodiment at an engine revolution speed set to the rated value.
FIG. 14
is a chart showing changes over time of the same status variables as shown in
FIG. 11
resulting when the combined operation of boom raising and swing is performed in the system of this embodiment at an engine revolution speed set lower than the rated value.
Referring to
FIG. 13
, in the boom-raising operation before the main relief valve
30
is operated for relief, the selector valve
60
is in the position as shown, and the output pressure Pc of the pressure control valve
34
is selected as an output pressure Pc′ of the selector valve
60
and then set as the target compensated differential pressure of the pressure compensating valves
21
a
,
21
b
, . . . . Also, the pump delivery pressure Ps is held higher than the maximum load pressure P
LMAX
by the target LS differential pressure P
GR
. Therefore, a target compensated differential pressure Pc′ of the pressure compensating valves
21
a
,
21
b
, . . . set by the output pressure of the pressure control valve
34
is equal to the target LS differential pressure P
GR
(Pc′=P
GR
).
When the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, the selector valve
60
is shifted, whereupon the target LS differential pressure P
GR
given by the output pressure of the pressure control valve
53
is selected as an output pressure Pc′ of the selector valve
60
and then set as the target compensated differential pressure of the pressure compensating valves
21
a
,
21
b
, . . . (Pc′=P
GR
). The output pressure Pc of the pressure control valve
34
at this time is Pc=0.
As a result, even when the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, the swing is not stopped and the operability in the combined operation is maintained.
The system of this embodiment operates likewise also when the engine revolution speed is set to a level lower than the rated value. More specifically, referring to
FIG. 14
, in the boom-raising operation before the main relief valve
30
is operated for relief, the output pressure Pc(=Pc′) of the pressure control valve
34
is set as the target compensated differential pressure of the pressure compensating valves
21
a
,
21
b
, . . . , and this target compensated differential pressure Pc′ is equal to the target LS differential pressure P
GR
(Pc′=P
GR
). When the boom cylinder
4
b
reaches the stroke end, the target LS differential pressure P
GR
given by the output pressure of the pressure control valve
53
is set as the target compensated differential pressure of the pressure compensating valves
21
a
,
21
b
, . . . (Pc′=P
GR
). The output pressure Pc of the pressure control valve
34
at this time is Pc=0. In this case, however, since the target LS differential pressure P
GR
is lower than that when the engine revolution speed is set to the rated value, the target compensated differential pressure Pc′ of the pressure compensating valves
21
a
,
21
b
, . . . is maintained at a level lower than when the engine revolution speed is set to the rated value.
As a result, even when the boom cylinder
4
b
reaches the stroke end and the main relief valve
30
is operated for relief, the swing is not stopped and the operability in the combined operation is maintained with no increase of the swing angular speed.
Furthermore, the output pressure Pc(=0) of the pressure control valve
34
is supplied to the LS control valve
12
b
of the LS/horsepower control regulator
12
, and the pump delivery rate is controlled so as to maximize within the range of horsepower control flow rate provided by the horsepower control tilting actuator
12
a.
Accordingly, this embodiment can also provide similar advantages as those in the first embodiment. In addition, with this embodiment, the speeds of the other actuators are avoided from lowering upon relief through the main relief valve
30
, and the LS control valve
12
b
of the horsepower control regulator
12
can be operated with stability.
A third embodiment of the present invention will be described with reference to FIG.
15
. In
FIG. 15
, identical members to those shown in
FIG. 1
are denoted by the same reference numerals. While, in the first and second embodiments, the differential pressure between the pump delivery pressure and the maximum load pressure is generated as an absolute pressure by the pressure control valve
34
and introduced to the pressure compensating valves and the LS control valve, the pump delivery pressure and the maximum load pressure are separately introduced as they are in this embodiment.
Referring to
FIG. 15
, a hydraulic drive system of this embodiment includes a hydraulic source
2
B and a valve apparatus
3
B. The hydraulic source
2
B and the valve apparatus
3
B have constructions different from those in the first embodiment.
More specifically, the hydraulic source
2
B includes an LS/horsepower control regulator
12
B for controlling the tilting (displacement) of the hydraulic pump
10
. The LS/horsepower control regulator
12
B comprises a horsepower control valve
12
Ba, an LS control valve
12
Bb, and a servo piston
12
Bc. The horsepower control valve
12
Ba and the servo piston
12
Bc cooperatively perform horsepower control for decreasing the tilting of the hydraulic pump
10
, while the LS control valve
12
Bb and the servo piston
12
Bc cooperatively perform load sensing control for holding the delivery pressure of the hydraulic pump
10
to be higher than the maximum load pressure of a plurality of actuators
4
a
,
4
b
,
4
c
by the target differential pressure.
The LS control valve
12
Bb includes a first operation drive unit
12
Bd and a second operation drive unit
12
Be which are each of the piston type and are disposed at an end of the LS control valve
12
Bb on the side acting to raise a pressure in a bottom-side chamber of the servo piston
12
Bc and to increase the tilting of the hydraulic pump
10
. The first operation drive unit
12
Bd has a pressure bearing section
70
a
on the side acting to increase the tilting and a pressure bearing section
70
b
on the side acting to decrease the tilting. The target differential pressure for the load sensing control (target LS differential pressure), given as the output pressure of the pressure control valve
51
of the target LS differential pressure generating circuit
5
, is introduced to the pressure bearing section
70
a
on the side acting to increase the tilting, and the pressure bearing section
70
b
on the side acting to decrease the tilting is communicated with a reservoir. The second operation drive unit
12
Be has a pressure bearing section
70
c
on the side acting to decrease the tilting and a pressure bearing section
70
d
on the side acting to increase the tilting. The delivery pressure of the hydraulic pump
10
is introduced to the pressure bearing section
70
c
on the side acting to decrease the tilting, and the pressure in the signal pressure line
35
a
(usually the maximum load pressure) is introduced to the pressure bearing section
70
d
on the side acting to increase the tilting.
The valve apparatus
3
B includes valve sections
3
Ba,
3
Bb,
3
Bc corresponding respectively to the actuators
4
a
,
4
b
,
4
c
, and another valve section
3
Bp. A plurality of closed center directional control valves
20
Ba,
20
Bb,
20
Bc and a plurality of pressure compensating valves
21
Ba,
21
Bb,
12
Bc are disposed respectively in the valve sections
3
Ba,
3
Bb,
3
Bc, whereas shuttle valves
22
a
,
22
b
constituting a part of a maximum load pressure detecting circuit, a main relief valve
30
, a fixed throttle
32
, and a signal pressure variable relief valve
33
are disposed in the valve section
3
Bp. The aforesaid pressure control valve
34
used in the first and second embodiments are not disposed in the valve section
3
Bp. Additionally, a variable unloading valve is omitted from the drawing.
The pressure compensating valve
21
Ba has pressure receiving sections
73
a
,
26
a
operating in the opening direction, and pressure receiving sections
27
a
,
74
a
operating in the closing direction. As with the first embodiment, the load pressure of the actuator
4
a
(pressure downstream of a meter-in throttle of the directional control valve
20
Ba) and a pressure upstream of the meter-in throttles of the directional control valve
20
Ba are introduced to the pressure receiving section
26
a
,
27
a
, respectively. On the other hand, the delivery pressure of the hydraulic pump
10
is introduced to the pressure receiving section
73
a
, and the pressure in the signal pressure line
35
a
(usually the maximum load pressure) is introduced to the pressure receiving section
74
a
. The pressure compensating valves
21
Bb,
21
Bc are also similarly constructed.
In the maximum load pressure line
35
, as with the first embodiment, the fixed throttle
32
and the signal pressure relief valve
33
are disposed. A setting pressure of the signal pressure relief valve
33
is set to be lower than a setting pressure of the main relief valve
30
, and the signal pressure relief valve
33
is constructed as a variable relief valve, of which setting pressure varies depending on the target LS differential pressure that changes with the engine revolution speed.
This embodiment having the above-described construction is essentially the same as the first embodiment except that the pump delivery pressure and the maximum load pressure are separately introduced, as they are, to the second operation drive unit
12
Be of the LS control valve
12
Bb and the pressure compensating valves
21
Ba,
21
Bb,
21
Bc instead of generating the differential pressure (absolute pressure) between the pump delivery pressure and the pressure in the signal pressure line
35
a
(usually the maximum load pressure) by the pressure control valve
34
and then introducing the generated differential pressure to those components. Hence, with the operation of the fixed throttle
32
and the signal pressure variable relief valve
33
, this embodiment can also provide similar advantages as those in the first embodiment.
A fourth embodiment of the present invention will be described with reference to FIG.
16
. In
FIG. 16
, identical members to those shown in
FIGS. 1 and 15
are denoted by the same reference numerals. While, in the first to third embodiments, the pressure compensating valve is of the before orifice type wherein it is disposed upstream of the meter-in throttle of the directional control valve, this embodiment employs a pressure compensating valve of the after orifice type wherein it is disposed downstream of the meter-in throttle of the directional control valve.
Referring to
FIG. 16
, a hydraulic drive system of this embodiment includes a valve apparatus
3
C. The valve apparatus
3
C has a construction different from that in the first embodiment.
The valve apparatus
3
C includes valve sections
3
Ca,
3
Cb,
3
Cc corresponding respectively to the actuators
4
a
,
4
b
,
4
c
, and another valve section
3
Bp. A plurality of closed center directional control valves
20
Ca,
20
Cb,
20
Cc and a plurality of pressure compensating valves
21
Ca,
21
Cb,
21
Cc are disposed respectively in the valve sections
3
Ca,
3
Cb,
3
Cc, whereas shuttle valves
22
a
,
22
b
constituting a part of a maximum load pressure detecting circuit, a main relief valve
30
, a fixed throttle
32
, and a signal pressure variable relief valve
33
are disposed in the valve section
3
Bp.
The pressure compensating valve
21
Ca is positioned downstream of meter-in throttles
81
,
82
of a directional control valve
20
Ca, and has a pressure receiving section
83
a operating in the opening direction and a pressure receiving section
84
a
operating in the closing direction. A pressure downstream of the meter-in throttle of the directional control valve
20
Ca is introduced to the pressure receiving section
83
a,
and the pressure in the signal pressure line
35
a
(usually the maximum load pressure) is introduced to the pressure receiving section
84
a
. The pressure compensating valves
21
Cb,
21
Cc are also similarly constructed.
In the case of employing the pressure compensating valves
21
Ca,
21
Cb,
12
Cc of the after orifice type like this embodiment, the pressures downstream of the meter-in throttles of the directional control valves
20
Ca,
20
Cb,
20
Cc are all controlled to a level substantially equal to the pressure in the signal pressure line
35
a
during the combined operation in which the actuators
4
a
,
4
b
,
4
c
are simultaneously driven. As a result, differential pressures across the meter-in throttles of the directional control valves
20
Ca,
20
Cb,
20
Cc are also controlled substantially in a similar manner. Thus, as with the case of employing the pressure compensating valves
21
Ca,
21
Cb,
12
Cc of the before orifice type, the hydraulic fluid can be supplied at a ratio depending on opening areas of the meter-in throttles of the directional control valves
20
Ca,
20
Cb,
20
Cc regardless of the magnitudes of load pressures and in the event of a saturation state where the delivery rate of the hydraulic pump
10
is insufficient for satisfying a demanded flow rate.
Also in this embodiment, the fixed throttle
32
and the signal pressure relief valve
33
are disposed in the maximum load pressure line
35
. A setting pressure of the signal pressure relief valve
33
is set to be lower than a setting pressure of the main relief valve
30
, and the signal pressure relief valve
33
is constructed as a variable relief valve, of which setting pressure varies depending on the target LS differential pressure that changes with the engine revolution speed. Therefore, even when the load pressure of any one actuator reaches the setting pressure of the main relief valve
30
during the combined operation in which a plurality of actuators
4
a
,
4
b
,
4
c
are simultaneously driven, the other actuators are neither stopped nor sped up, and good operability in the combined operation is maintained.
Industrial Applicability
According to the present invention, even when a load pressure of any one actuator reaches a setting pressure of a main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the other actuators are not stopped and good operability in the combined operation can be ensured.
Also, according to the present invention, even when a load pressure of any one actuator reaches a setting pressure of a main relief valve during the combined operation in which a plurality of actuators are simultaneously driven, the other actuators are not sped up and good operability in the combined operation can be ensured.
Simultaneously, a pump LS control system can be held in a stable condition.
Claims
- 1. A hydraulic drive system comprising an engine (1), a variable displacement hydraulic pump (10) driven by said engine, a plurality of actuators (4a,4b) driven by a hydraulic fluid delivered from said hydraulic pump, a plurality of directional control valves (20a,20b; 20Ba,20Bb; 20Ca,20Cb) 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 (21a,21b; 21Ba,21Bb; 21Ca,21Cb) for controlling respective differential pressures across said plurality of directional control valves, pump control means (12; 12B) for performing load sensing control to hold a delivery pressure of said hydraulic pump higher than a maximum load pressure of said plurality of actuators by a target differential pressure, and a main relief valve (30) for restricting an upper limit of the delivery pressure of said hydraulic pump, a target compensated differential pressure (Pc) for each of said plurality of pressure compensating valves being set in accordance with a differential pressure (Ps−PLMAX) between the delivery pressure of said hydraulic pump and the maximum load pressure of said plurality of actuators, a target differential pressure (PGR) in said load sensing control being set as a variable value depending on a revolution speed of said engine, wherein:said hydraulic drive system further comprises target compensated differential pressure modifying means (32,33; 60) for setting, as the target compensated differential pressure (Pc) for each of said plurality of pressure compensating valves (21a,21b; 21Ba,21Bb; 21Ca,21Cb), a modification value (PGR−α; PGR) different from the differential pressure between the delivery pressure of said hydraulic pump and the maximum load pressure of said plurality of actuators (4a,4b), when the delivery pressure of said hydraulic pump (10) rises up to a setting pressure of said main relief valve (30).
- 2. A hydraulic drive system according to claim 1, wherein said modification value (PGR−α; PGR) is a variable value depending on the revolution speed of said engine (1).
- 3. A hydraulic drive system according to claim 1, wherein said modification value (PGR−α; PGR) is equal to or smaller than the target differential pressure (PGR) In said load sensing control set as a variable value depending on the revolution speed of said engine (1).
- 4. A hydraulic drive system according to claim 1, wherein said target compensated differential pressure modifying means (32,33) includes a signal pressure relief valve (33) which is provided in a maximum load pressure line (35,35a) for detecting the maximum load pressure, and which reduces an upper limit of the maximum load pressure detected by said maximum load pressure line to be lower than the setting pressure of said main relief valve (30) by said modification value (PGR−α).
- 5. A hydraulic drive system according to claim 4, wherein said signal pressure relief valve (33) is a variable relief valve, and assuming a relief setting pressure of said variable relief valve to be PLMAX0, the target differential pressure in said load sensing control to be PGR, and the setting pressure of said main relief valve to be PR, the relief setting pressure PLMAX0 of the variable relief valve is set so as to satisfy:PLMAX0=PR−PGR+α(where α is a value smaller than PGR).
- 6. A hydraulic drive system according to claim 1, wherein said target compensated differential pressure modifying means (60) includes a selector valve (60) for changing over the target compensated differential pressure (Pc) from the differential pressure (Ps—PLMAX) between the delivery pressure of said hydraulic pump and the maximum load pressure of said plurality of actuators (4a, 4b) to the target differential pressure (PGR) in said load sensing control, immediately before the delivery pressure of said hydraulic pump (10) rises up to the setting pressure (PR) of said main relief valve (30).
Priority Claims (1)
Number |
Date |
Country |
Kind |
2000-004074 |
Jan 2000 |
JP |
|
PCT Information
Filing Document |
Filing Date |
Country |
Kind |
PCT/JP01/00057 |
|
WO |
00 |
Publishing Document |
Publishing Date |
Country |
Kind |
WO01/51820 |
7/19/2001 |
WO |
A |
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