TECHNICAL FIELD
The present invention relates to a displacement
control system for controlling a displacement of a variable
displacement type hydraulic pump to be employed in a
hydraulic circuit supplying a pressurized fluid to an
actuator of a construction machine or so forth.
BACKGROUND ART
As a hydraulic circuit supplying a pressurized
fluid to an actuator of a construction machine and so
forth, one supplying a discharged pressurized fluid of the
hydraulic pump to the actuator through the operation valve,
has been known. When a closed center type operation valve
which shuts off a pump port at a neutral position, is
employed as the operation valve in such hydraulic circuit,
a discharge passage of the hydraulic pump becomes dead
ended while the operation valve is in the neutral position,
to make the discharged pressurized fluid at high pressure.
Thus, a driving horse power consumption of an engine
driving the hydraulic pump becomes large.
As a hydraulic circuit resolving this problem,
there has been known a circuit, in which a variable
displacement type hydraulic pump (hereinafter referred to
as a variable hydraulic pump) is employed as the hydraulic
pump, a displacement (a discharge amount per one revolution
cycle) of the variable hydraulic pump is controlled to make
the displacement smaller when a differential pressure
between an inlet side pressure (pump discharge pressure)
and an outlet side pressure (load pressure) is large and to
make the displacement larger when the differential pressure
is small and whereby to make the differential pressure
constant, and a discharge flow rate (displacement x number
of revolution per unit period) of the variable hydraulic
pump can be a value corresponding to an opening degree (a
communication area between a pump port and an actuator
port) of the operation valve, as disclosed in Japanese
Unexamined Utility Model Publication (Kokai) No. Heisei 5-86003.
However, in such a hydraulic circuit, the
discharge flow rate of the variable hydraulic pump is
displacement x number of revolution per unit period, and
thus becomes small when an engine speed is low and large
when the engine speed is high even when the displacement is
constant, to differentiate the differential pressure of the
inlet side pressure and the outlet side pressure even when
the opening degree of the operation valve is the same.
Therefore, even when the opening degree of the operation
valve is the same, the displacement of the variable
hydraulic pump is controlled to be larger when the engine
speed is low and to be smaller when the engine speed is
high so that the discharge flow rate becomes a value
corresponding to the opening degree.
Therefore, even if the engine speed is lowered
when fine operation for fine actuation of the actuator by
reducing a supply flow rate to the actuator is desired, the
displacement of the variable hydraulic pump becomes larger
to increase discharge flow rate of the variable hydraulic
pump to make it impossible to perform fine operation.
Therefore, as a solution for such drawback, in
the foregoing Japanese Unexamined Utility Model Publication
No. Heisei 5-86003, a fixed displacement type hydraulic
pump (hereinafter referred to a fixed displacement
hydraulic pump) is driven by the engine which drives the
variable hydraulic pump, and a drain circuit including a
restriction and a relief valve is connected to a discharge
passage of the fixed displacement hydraulic pump, a
pressure on the side of the fixed displacement hydraulic
pump in relation to a junction in the discharge passage is
detected to control the displacement of the variable pump
depending upon the detected pressure.
Thus, since the detected pressure becomes a value
corresponding to the engine speed, the displacement of the
variable hydraulic pump can be controlled with taking the
engine speed into account. As a result, the discharge flow
rate of the variable hydraulic pump can be controlled with
taking the engine speed into account, and whereby the
differential pressure of the pump discharge pressure and
the load pressure becomes a value corresponding to the
engine speed.
However, in the construction set forth above,
since a part of the discharged pressurized fluid of the
fixed displacement hydraulic pump is flowed into a tank
through the restriction and the relieve valve, the
discharged pressurized fluid cannot be used effectively.
Also, when the discharged pressurized fluid is supplied to
other hydraulic device, a flow rate to be flowed into the
tank is reduced to cause a variation of the differential
pressure between the upstream side and the downstream side
of the restriction to cause a variation of the detected
pressure to vary displacement control characteristics of
the variable hydraulic pump.
On the other hand, when the engine speed is
extremely low, the discharge flow rate of the fixed
displacement hydraulic pump can be too small to elevate the
hydraulic pressure to a set pressure of the relief valve.
In such case, the displacement of the variable hydraulic
pump cannot be controlled with taking the engine speed into
account.
Therefore, in view of the problems set out above,
it is an object of the present invention to provide a
displacement control system for a variable displacement
type hydraulic pump, in which a discharge flow rate of the
variable displacement type hydraulic pump becomes extremely
small at low engine speed to improve operability in fine
operation, a discharged pressurized fluid of a fixed
displacement type hydraulic pump can be effectively used
without flowing out to a tank, a displacement control
characteristics of the variable displacement type hydraulic
pump can be maintained constant even when the discharged
pressure of the fixed displacement type hydraulic pump is
supplied to other hydraulic device, and the displacement of
the variable displacement type hydraulic pump can be
controlled with taking an engine speed into account even
when the engine speed is extremely low.
DISCLOSURE OF THE INVENTION
In order to accomplish the above-mentioned
object, according to one aspect of the present invention,
a displacement control system for a variable
displacement type hydraulic pump comprising a cylinder for
varying a displacement of the variable displacement type
hydraulic pump, a control valve for controlling supply and
drain of a discharge pressure of the variable displacement
type hydraulic pump to and from the cylinder, a fixed
displacement type hydraulic pump driven simultaneously with
the variable displacement type hydraulic pump by a common
engine and a restriction provided in a discharge passage of
the fixed displacement type hydraulic pump, the control valve being operated for switching by
comparison of a differential pressure between the discharge
pressure and a load pressure and a differential pressure
between upstream and downstream of the restriction as a set
differential pressure, for controlling a displacement of
the variable displacement type hydraulic pump via the
cylinder so that the differential pressure becomes a value
corresponding to the set differential pressure.
With the construction set forth above, when the
engine speed is low, the discharge flow rate of the fixed
displacement type hydraulic pump becomes smaller to make
the differential pressure between upstream and downstream
of the restriction smaller to make the set differential
pressure smaller. On the other hand, when the engine speed
is high, the discharge flow rate of the fixed displacement
type hydraulic pump becomes larger to make the differential
pressure between upstream and downstream of the restriction
larger to make the set differential pressure larger.
Accordingly, when the engine is in low speed
revolution, the discharge flow rate of the variable
displacement type hydraulic pump becomes extremely small to
improve operability in fine operation.
On the other hand, since the set differential
pressure of the control valve is varied by the differential
pressure between upstream and downstream of the restriction
provided in the discharge passage of the fixed displacement
type hydraulic pump, the discharged pressurized fluid of
the fixed displacement type hydraulic pump can be used
effectively without flowing out to the tank. Also, when
the engine speed is constant, the differential pressure
between upstream and downstream of the restriction is not
varied even when the discharged pressurized fluid of the
fixed displacement type hydraulic pump is supplied to other
hydraulic device, the displacement control characteristics
of the variable displacement type hydraulic pump can be
constant. Furthermore, the differential pressure between
upstream and downstream of the restriction is generated
even at extremely low engine speed to enable control with
taking the engine speed into account.
According to the second aspect of the present
invention, a displacement control system comprises a
cylinder for varying a displacement of the variable
displacement type hydraulic pump, a control valve for
controlling supply and drain of a discharge pressure of the
variable displacement type hydraulic pump to and from the
cylinder, a fixed displacement type hydraulic pump driven
simultaneously with the variable displacement type
hydraulic pump by a common engine and a restriction
provided in a discharge passage of the fixed displacement
type hydraulic pump, a first additional restriction
provided in a pilot circuit between the downstream side of
the restriction and the control valve and a second
additional restriction provided in a drain passage
connected to downstream side of the first additional
restriction,
the control valve being operated for switching by
comparison of a differential pressure of the discharge
pressure and a load pressure and a differential pressure
between the upstream side of the restriction and the
downstream side of the first additional restriction as a
set differential pressure, for controlling a displacement
of the variable displacement type hydraulic pump via the
cylinder so that the differential pressure becomes a value
corresponding to the set differential pressure.
According to the third aspect of the present
invention, a displacement control system for a variable
displacement type hydraulic pump comprising a cylinder for
varying a displacement of the variable displacement type
hydraulic pump, a control valve for controlling supply and
drain of a discharge pressure of the variable displacement
type hydraulic pump to and from the cylinder, a fixed
displacement type hydraulic pump driven simultaneously with
the variable displacement type hydraulic pump by a common
engine, a restriction provided in a discharge passage of
the fixed displacement type hydraulic pump and a switching
valve for switching supply and drain of the discharge
pressure to and from the control valve and being associated
with the cylinder via the spring,
the control valve being operated for switching by
comparison of a differential pressure between the discharge
pressure and a load pressure and a differential pressure
between upstream and downstream of the restriction as a set
differential pressure, for controlling a displacement of
the variable displacement type hydraulic pump via the
cylinder so that the differential pressure becomes a value
corresponding to the set differential pressure, and the switching valve being operated for switching
by comparison of the discharge pressure and a mounting load
of the spring, for controlling a displacement of the
variable displacement type hydraulic pump via the control
valve and the cylinder to maintain an input torque of the
variable displacement type hydraulic pump constant.
BRIEF DESCRIPTION OF THE DRAWINGS
The present invention will be understood more
fully from the detailed description given herebelow and
from the accompanying drawings of the preferred embodiment
of the invention, which, however, should not be taken to be
limitative to the present invention, but are for
explanation and understanding only.
Fig. 1 is a diagrammatic explanatory illustration
of a construction of the first embodiment of a displacement
control system for a variable displacement type hydraulic
pump according to the present invention; Fig. 2 is a chart showing a relationship between
an engine speed and a discharge flow rate of the variable
hydraulic pump in the first embodiment; Fig. 3 is a section showing a particular
construction of a control valve of the first embodiment; Fig. 4 is a diagrammatic explanatory illustration
of a construction of the second embodiment of a
displacement control system for a variable displacement
type hydraulic pump according to the present invention; and Fig. 5 is a diagrammatic explanatory illustration
of a construction of the third embodiment of a displacement
control system for a variable displacement type hydraulic
pump according to the present invention.
BEST MODE FOR IMPLEMENTING THE INVENTION
The preferred embodiment of a displacement
control system for a variable displacement type hydraulic
pump will be discussed hereinafter with reference to the
accompanying drawings.
Fig. 1 shows the first embodiment of a
displacement control system for a variable displacement
type hydraulic pump according to the present invention. As
shown in Fig. 1, by an engine 1, the variable displacement
type hydraulic pump 2 (hereinafter referred to as a
variable hydraulic pump 2) and a fixed displacement type
hydraulic pump 3 (hereinafter referred to as a fixed
displacement hydraulic pump 3) are driven. Then, a
discharge passage 4 of the variable hydraulic pump 2 is
connected to a pump port 6 of an operation valve 5. First
and second actuator ports 7 and 8 of the operation valve 5
are respectively connected to a first chamber 12 and a
second chamber 13 of an actuator 11 via respective of first
and second circuits 9 and 10. A tank port 14 is connected
to a tank 15.
A discharge passage 16 of the fixed displacement
hydraulic pump 3 is connected to an inlet side of a
hydraulic pilot valve 17. By operating the hydraulic pilot
valve 17, a discharged pressurized fluid of the fixed
displacement hydraulic pump 3 is supplied to the other
hydraulic device. The discharge passage 16 is provided
with a restriction 18.
The reference numeral 21 denotes a cylinder which
has a piston 22, a large diameter pressure receiving
chamber 23 and a small diameter pressure receiving chamber
24 defined at both sides of the piston 22, and the piston
22 is connected to a swash plate 20 of the variable
hydraulic pump 2. Then, the swash plate 20 for varying a
displacement of the variable hydraulic pump 2 is designed
to vary an angle thereof by being pivotally tilted by a
piston 22 of a swash plate control cylinder 21. The piston
22 of the cylinder 21 is moved in a displacement reducing
direction by a pressurized fluid of the large diameter
pressure receiving chamber 23 and moved in a displacement
increasing direction by a pressurized fluid of the small
diameter pressure receiving chamber 24 and a spring 25.
The large diameter pressure receiving chamber 23
is selectively connected to the tank 15 or the discharge
passage 4, by a control valve 26 and the small diameter
pressure receiving chamber 24 is connected to the discharge
passage 4.
The control valve 26 has a first port 27
connected to the discharge passage 4 via the small diameter
pressure receiving chamber 24 and a second port 28
connected to the large diameter pressure receiving chamber
23 and a tank port 29. The control valve 26 is changed
over to a drain position A to shut off the first port 27
and to communicate the second port 28 with the tank port 29
by a pressurized fluid of a first pressure receiving
portion 30 and a pressurized fluid of a first auxiliary
pressure receiving portion 31, and is changed over to a
supply position B to communicate the first portion 27 with
the second port 28 and to shut off the tank port 29 by the
pressurized fluid of a second auxiliary pressure receiving
portions 32 a pressurized fluid of a second auxiliary
pressure receiving portion 33.
The first pressure receiving portion 30 is
connected to an outlet side of a shuttle valve 19 detecting
a higher pressure of the first circuit 9 and the second
circuit 10 to be supplied with the outlet side pressure
(load pressure) of the operation valve 5. On the other
hand, the second pressure receiving portion 32 is
communicated with the discharge passage 4 via the small
diameter pressure receiving chamber 24 to be supplied with
the inlet side pressure (pump discharge pressure) of the
operation valve 5. Furthermore, the first auxiliary
pressure receiving portion 31 is connected to a upstream
side of the restriction 18, and the second auxiliary
pressure receiving portion 33 is connected to a downstream
side of the restriction 18.
A revolution speed of the engine 1 is controlled
by feeding an engine speed command signal generated from an
operation member 34, such as an accelerator pedal or the
like, to a control governor 35 of the engine 1. For
example, when the operation member 34 is placed at a low
speed position a, the engine speed becomes low speed, and
at a medium speed position b, the engine speed becomes
medium speed, and at a high speed position c, the engine
speed becomes high speed.
Next, operation of the first embodiment set forth
above will be discussed.
The control valve 26 is switched into the supply
position B to pivoted the swash plate 4 in a direction to
reduce the tilt angle thereof when the differential
pressure between the inlet side pressure (pump discharge
pressure) and the outlet side pressure (load pressure) of
the operation valve 5 is greater than the set pressure, and
is switched into the drain position A to tilt the swash
plate 4 in a direction to increase the tilt angle thereof
when the differential pressure is smaller than the set
pressure to make the differential pressure between the
inlet side pressure and the outlet side pressure of the
operation valve 5 constant. By this, the discharge flow
rate of the variable hydraulic pump 2 becomes a value
corresponding to an opening degree (a communication area of
the pump port 6 and the first or second actuator ports 7 or
8) of the operation valve 5.
On the other hand, a set differential pressure of
the control valve 26 is varied depending upon a
differential pressure between upstream and downstream of
the restriction 18. The differential pressure between
upstream and downstream of the restriction 18 is
proportional to a square of a discharge flow rate (engine
speed) of the fixed displacement hydraulic pump 3.
Thus, the set differential pressure of the
control valve 26 becomes small when the engine speed is
low, and becomes large when the engine speed is high. On
fine operation at low engine speed, the discharge flow rate
of the variable hydraulic pump 2 is reduced than that at
the high speed to improve a fine operation ability.
Namely, the differential pressure between
upstream and downstream of the restriction 18 is
proportional to a square of the engine speed (discharge
flow rate of the fixed displacement hydraulic pump 3) as
shown in Fig. 2(a). The set differential pressure between
control valve 26 is linearly proportional to the
differential pressure of the upstream and downstream of the
restriction 18 as shown in Fig. 2(b). The displacement of
the variable hydraulic pump 2 is proportional to a square
of the set differential pressure as shown in Fig. 2(c).
Therefore, the discharge flow rate of the variable
hydraulic pump 2 is linearly proportional to the engine
speed as shown in Fig. 2(d).
Next, a particular construction of the control
valve 26 will be explained.
As shown in Fig. 3, a sleeve 42 is threadingly
inserted into a sleeve bore 41 formed in a valve body 40,
such as a housing or the like of the variable displacement
pump 2. The sleeve bore 41 is formed with a first inflow
port 65, a flow out port 66, a control port 67, a pump
pressure supply port 68 and a second inflow port 69. The
sleeve 42 is formed with a first port 43, a second port 44,
a third port 45, a fourth port 46 and a fifth port 47.
Within a spool insertion bore 42a located at an axial
center portion of the sleeve 42, a spool 48 is slidably
disposed.
The spool 48 has a first small diameter portion
50, a second small diameter portion 51 and a third small
diameter portion 52. In an axial bore 48a formed at one
end portion of the spool 48, a small diameter portion 54 of
a stationary piston 53 inserted in the spool insertion bore
42a is inserted to define a first pressure receiving
chamber 55 (the second pressure receiving portion 32 of
Fig. 1) and a second pressure receiving chamber 56 (the
second auxiliary pressure receiving chamber 33 of Fig. 1).
Then, the first pressure receiving chamber 55 is
communicated with the second small diameter portion 51
through a fluid conduit 57 formed in the axial center
portion of the spool 48. Also, by projecting the third
small diameter portion 52 into the spring insertion hole
42b through an axial center bore 42c of the sleeve 42, a
third pressure receiving chamber 58 is defined.
To one end of the spring insertion hole 42b of
the sleeve 42, a cylindrical tip end portion 60 of the
threaded rod 59 is engaged to define a fourth pressure
receiving chamber 61. Furthermore, the threaded rod 59 is
threadingly engaged with the sleeve 42 and fixed by
tightening a lock nut 62. Then, in the axial center
portion of the threaded rod 59, a bore 64 communicated with
the fourth pressure receiving chamber 61 via the
cylindrical tip end portion 60, is formed and extended from
a piping joint portion 63 at the other end. Also, in the
spring insertion hole 42b, the spring 49 is disposed
between the other end of the spool 48 and the cylindrical
tip end 60 of the threaded rod 59.
It should be noted that a pressure receiving
diameter d1 of the first pressure receiving chamber 55 and
a pressure receiving diameter d2 of the fourth pressure
receiving chamber 61 are equal to each other.
The first port 43 is communicated with the second
pressure receiving chamber 56 and the first inflow port 65.
The first inflow port 65 is connected to the downstream
side of the restriction 18 shown in Fig. 1. The second
port 44 is connected to the flow out port 66 (the tank port
29 in Fig. 1). The flow out port 66 is communicated with
the tank 15 shown in Fig. 1. The third port 45 is
connected to a control port 67 (the second port 28 in Fig.
1). The control port 67 is connected to the large diameter
pressure receiving chamber 23. The fourth port 46 is
connected to a pump pressure supply port 68 (the first port
27 in Fig. 1). The pump pressure supply port 68 is
connected to the discharge passage 4 via the small diameter
chamber 24. The fifth port 47 is connected to the third
pressure receiving chamber 58 (the first auxiliary pressure
receiving portion 31 in Fig. 1) and a second inflow port
69. The second inflow port 69 is connected to the upstream
side of the restriction 18 shown in Fig. 1. To the piping
joint portion 63, a not shown hose is connected. Through
these, the outlet side pressure (load pressure) of the
operation valve 5 is supplied to the fourth pressure
receiving portion 61 (first pressure receiving portion 30
in Fig. 1).
Next, operation of the control valve 26 is
discussed.
A pump discharge pressure P1 of the variable
hydraulic pump 2 is supplied to the fourth port 46 through
the pump pressure supply port 68, and supplied to the first
pressure receiving chamber 55 via the second small diameter
portion 51 and the fluid conduit 57 to thrust the spool 48
toward right. The outlet side pressure (load pressure) PLS
of the operation valve is supplied to the fourth pressure
receiving chamber 61 to thrust the spool 48 toward left.
Here, when the differential pressure of the pump
discharge pressure P1 and the load pressure PLS is zero (P1
= PLS ), namely when the opening degree of the operation
valve is maximum, since the pressure receiving diameter d1
of the first pressure receiving chamber 55 and the pressure
receiving diameter d2 of the fourth pressure receiving
chamber 61 are equal to each other, the spool 48 is thrust
toward left by the spring 49, as shown in Fig. 3.
By this, through the first small diameter portion
50 of the spool 48, the second port 44 communicates with
the third port 45 to establish communication of the control
port 67 with the tank 15 via the flow out port 66 to place
the control valve 26 at the drain position A in Fig. 1.
Then, the pressurized fluid of the large diameter pressure
receiving chamber 23 of the swash plate control valve 21
flows out, the displacement of the variable hydraulic pump
2 is increased, accordingly.
Once the displacement of the variable hydraulic
pump 2 becomes large to increase the discharge flow rate,
the flow rate of the pressurized fluid flowing from the
pump port 6 of the operation valve 5 to the first or second
actuator port 7 or 8 is increased. Thus, when the opening
degree of the operation valve 5 is maintained as that in
the condition set forth above, a pressure loss of the
operation valve 5 is increased to make the differential
pressure of the pump discharge pressure P1 and the load
pressure PLS large.
By this, a force acting on the spool 48 becomes
π/4d1 2 x P1 - π/2d2 2 x PLS = π/4d2 2 x ΔPLS > 0
When the force becomes greater than a mounting load of the
spring 49, the spool 48 is moved toward right to be placed
in a condition shown in Fig. 3 to block the communication
between the second port 44 and the third port 45. Thus,
the communication between the control port 67 and the flow
out port 66 is blocked.
On the other hand, since d1 = d2 as set forth
above, if π/4d2 2 x PLS and an initial mounting load of the
spring 49 are equal to each other, the spool 48 is
constantly stopped at the position shown in Fig. 3. Thus,
the differential pressure ΔPLS between the discharge
pressure P1 of the pump and the load pressure PLS is always
maintained constant. The differential pressure (the set
differential pressure) is determined by the initial
mounting load of the spring 49.
On the other hand, the spool 48 is thrust toward
right by the pressure of the downstream side of the
restriction 18 acting in the second pressure receiving
chamber 56 and thrust toward left by the pressure of the
upstream side of the restriction 18 acting in the third
pressure receiving chamber 58. Thus, when the spool 48 is
thrust toward left by the differential pressure between
upstream and downstream of the restriction 18, the set
differential pressure is greater than that in the case set
forth above.
Namely, a condition upon stopping of the spool 48
at the position shown in Fig. 3 is π/4d2 2 x ΔPLS = mounting
load of the spring 49 + π/4(D2 - d1 2) x (P2 - P3). It should
be noted that D is a diameter of the spool 48, P2 is the
pressure at upstream side of the restriction 18, and P3 is
the pressure at the downstream side of the restriction 18.
When the desired differential pressure ΔPLS is
set at the engine speed set in the foregoing condition,
when the engine speed is low, the discharge flow rate of
the fixed displacement hydraulic pump 3 is decreased to
reduce the differential pressure (P2 - P3) at upstream and
downstream of the restriction 18 to lower the set
differential pressure. Thus, the spool 48 is moved toward
right to establish communication between the third port 45
and the fourth port 46. Thus, a condition where the
control valve 26 of Fig. 1 is placed at the supply position
B, in which the discharge pressure fluid of the variable
hydraulic pump 2 flows to the control port 67 and then
supplied to the large diameter pressure receiving chamber
23, is established. By this, the swash plate 20 is tilted
in the displacement reducing direction to reduce the
displacement of the variable hydraulic pump 2 to
significantly smaller the discharge flow rate of the
variable hydraulic pump 2 than that at high speed.
Fig. 4 shows the second embodiment of the
displacement control system for the variable displacement
type hydraulic pump according to the present invention.
This is constructed by providing a first restriction 71 in
a pilot circuit 70 connecting the downstream side of the
restriction 18 in the discharge passage 16 of the fixed
displacement hydraulic pump 3 and the second auxiliary
pressure receiving portion 33 of the control valve 26, and
connecting the downstream side of the first restriction 71
with the tank 15 through a drain circuit 73, and providing
a second restriction 73 in the drain circuit 72.
The discharge passage 16 of the fixed
displacement hydraulic pump 3 is connected to a swiveling
hydraulic motor 76 via an auxiliary operation valve 75. By
switching the auxiliary operation valve 75 to the supply
position, a swiveling hydraulic motor 76 is driven.
Next, operation of the second embodiment will be
discussed.
By providing the first restriction 71 in the
pilot circuit 70 and, in conjunction therewith, connecting
the downstream side of the restriction 71 to the tank 15
via the second restriction 73, a pressure P4 acting in the
second auxiliary pressure receiving portion 33 of the
control valve 26 is lowered upon passing through the first
restriction 71 to be lower than the downstream side
pressure P3 of the restriction 18 to establish P3 > P4.
A ratio of lowering of pressure becomes a given
ratio determined by the flow area of the first restriction
71 and the flow area of the second restriction 73.
Upon driving the swiveling hydraulic motor 76 by
supplying the discharged pressurized fluid of the fixed
displacement hydraulic pump 3 by placing the auxiliary
operation valve 75 at the supply position, since a start up
torque of the swiveling hydraulic motor 76 is large, the
discharge pressure of the fixed displacement hydraulic pump
3 upon starting up becomes significantly high, and during
steady swiveling action, the discharge pressure of the
fixed hydraulic pump 3 becomes low.
On the other hand, the fixed displacement
hydraulic pump 3 is constructed with a gear pump, for
example. When the discharge pressure becomes high,
internal leakage amount is increased to lower efficiency.
Therefore, even at the same revolution speed, the discharge
flow rate of the fixed displacement hydraulic pump 3 in
high pressure is reduced than that in the low pressure.
Once the discharge flow rate of the fixed
displacement hydraulic pump 3 is reduced, the differential
pressure (P2 - P3) between upstream and downstream of the
restriction 18 becomes smaller. Therefore, upon starting
up of the swiveling hydraulic motor 76, the differential
pressure (P2 - P) between upstream and downstream of the
restriction 18 becomes smaller. Also, during steady
revolution, the differential pressure (P2 - P3) between
upstream and downstream of the restriction 18 becomes
large.
Therefore, if the first and second restrictions
71 and 73 are not provided in the pilot circuit 70 and when
the pressure P3 at the downstream side of the restriction
18 directly acts on the second auxiliary pressure receiving
portion 33 of the control valve 26, the differential
pressure between the pressure acting in the first auxiliary
pressure receiving portion 31 of the control valve 26 and
the pressure acting in the second auxiliary pressure
receiving portion 33 is varied between that upon starting
up of the swiveling hydraulic motor 76 and that during
steady revolution to differentiate the discharge flow rate
of the variable hydraulic motor 2.
However, as shown in Fig. 4, by providing the
first and second restrictions 71 and 73 in the pilot
circuit 70, the pressure P4 acting in the second auxiliary
pressure receiving portion 33 of the control valve 26
becomes lower pressure at a given ratio than the pressure
P3 of the downstream side of the restriction 18. Therefore,
the discharge flow rate of the fixed displacement hydraulic
pump 3 as set forth above is reduced by lowering of
efficiency due to the discharge pressure, the differential
pressure between the pressures in the first auxiliary
pressure receiving portion 31 and the second auxiliary
pressure receiving portion 32 of the control valve 26
becomes substantially constant so that the discharge flow
rate of the variable hydraulic pump 2 will not be varied
between the start-up and by steady revolution of the
swiveling hydraulic motor 76.
For example, it is assumed that the discharge
pressure of the fixed displacement hydraulic pump 3 is 50
kg/cm2 and the discharge flow rate is 20 l/min, the pressure
P3 at the downstream side of the restriction 18 is 40
kg/cm2, and the pressure P4 in the second auxiliary pressure
receiving portion 33 is 39.5 kg/cm2 during steady revolution
of the swiveling hydraulic motor 76, the differential
pressure (P2 - P4) between the pressure P2 acting in the
first auxiliary pressure receiving portion 31 and the
pressure P4 acting in the second auxiliary pressure
receiving portion 33 of the control valve 26 becomes 10
kg/cm2.
In the condition set forth above, upon starting
up of the swiveling hydraulic motor, when the discharge
pressure of the fixed displacement hydraulic pump 3 is 200
kg/cm2 and the discharge flow rate is 18 l/min, the
downstream side pressure P3 of the restriction 18 becomes
192 kg/cm2 and the pressure P4 acting in the second
auxiliary pressure receiving portion 33 becomes
substantially 189.5 kg/cm2. Thus, the foregoing
differential pressure (P2 - P4) becomes substantially 10
kg/cm2.
Fig. 5 shows the third embodiment of the
displacement control system of the variable displacement
type hydraulic pump according to the present invention.
The drain port 29 of the control valve 26 is selectively
connected to one of the tank 15 and the discharge passage
4 by the switching valve 80 for an input torque control.
The switching valve 80 is changed over to a drain
position C by the spring 81 and is changed over to a supply
position D by the discharge pressure of the variable
displacement pump 2 acting in the first pressure receiving
portion 82, and an external pressure acting in the second
pressure receiving portion 83. Furthermore, the spring 81
is associated with the piston 22 by a link 84.
Since the shown embodiment is constructed as set
forth above, when the pump discharge pressure of the
variable displacement pump 2 becomes higher than a pressure
corresponding to the mounting load of the spring 81, the
switching valve 80 is changed over to the supply position
D. Thus, the discharge pressure flows into the large
diameter pressure receiving chamber 23 via the control
valve 26. Therefore, the piston 22 is moved toward right
to pivot the swash plate 20 to tilt in the direction of
smaller displacement. By movement of the piston 22, the
mounting load of the spring 81 is increased via the link 84
to change over the switching valve 80 back to the drain
position C.
Thus, by repeating such operation, the
displacement of the variable hydraulic pump 2 is controlled
so that an input torque (pump discharge pressure x
displacement) becomes constant.
As set forth above, according to the present
invention, when the revolution speed of the engine 1 is
low, the discharge flow rate of the fixed displacement type
hydraulic pump 3 is reduced. By this, the differential
pressure between upstream and downstream of the restriction
18 becomes smaller to make the set differential pressure of
the control valve 26 smaller. On the other hand, when the
revolution speed of the engine 1 is high, the discharge
flow rate of the fixed displacement type hydraulic pump 3
is increased. Associating therewith, the differential
pressure between upstream and downstream of the restriction
18 becomes larger to make the set differential pressure of
the control valve 26 larger.
Accordingly, when the engine 1 is in low
revolution speed, the discharge flow rate of the variable
displacement type hydraulic pump 2 is significantly smaller
to improve the operability in fine motion.
On the other hand, since the set differential
pressure of the control valve 26 is varied by the
differential pressure between upstream and downstream of
the restriction 18 provided in the discharge passage 16 of
the fixed displacement type hydraulic pump 3, the
discharged pressurized fluid of the fixed displacement type
hydraulic pump 3 does not flow out to the tank 15 and thus
can be used effectively. Also, as long as the engine speed
is constant, the differential pressure between upstream and
downstream of the restriction 18 will not be varied even
when the discharged pressurized fluid of the fixed
displacement type hydraulic pump 3 is supplied to other
hydraulic device. Thus, the displacement control
characteristics of the variable displacement type hydraulic
pump 2 can be made constant. Furthermore, even if the
engine 1 is in extremely low speed, the differential
pressure between upstream and downstream of the restriction
18 is caused to make it possible to control the
displacement of the variable displacement type hydraulic
pump 2 with taking the engine speed into account.
On the other hand, according to the present
invention, the pressure P4 acting in the second auxiliary
pressure receiving portion 33 of the control valve 26
becomes a value corresponding to the pressure P3 at
downstream of the restriction 18 provided in the discharge
passage 16 of the fixed displacement type hydraulic pump 3
lowered by a given ratio. When the revolution speed of the
engine 1 is constant and the discharge flow rate is varied
associating with the variation of efficiency due to the
discharge pressure of the fixed displacement type hydraulic
pump 3, the pressure difference between the first auxiliary
pressure receiving portion 31 and the second auxiliary
pressure receiving portion 33 of the control valve 26
becomes substantially constant.
Accordingly, even when the revolution speed of
the fixed displacement type hydraulic pump 3 is the same
and the discharge pressure thereof is varied between high
pressure and low pressure, the set differential pressure of
the control valve 26 can be constant to make the discharge
flow rate of the variable displacement type hydraulic pump
2 can be substantially constant.
Although the invention has been illustrated and
described with respect to exemplary embodiment thereof, it
should be understood by those skilled in the art that the
foregoing and various other changes, omissions and
additions may be made therein and thereto, without
departing from the spirit and scope of the present
invention. Therefore, the present invention should not be
understood as limited to the specific embodiment set out
above but to include all possible embodiments which can be
embodied within a scope encompassed and equivalents thereof
with respect to the feature set out in the appended claims.