U.S. patent number 3,777,773 [Application Number 05/228,020] was granted by the patent office on 1973-12-11 for pressure compensating valve mechanism.
This patent grant is currently assigned to Koehring Company. Invention is credited to William N. Tolbert.
United States Patent |
3,777,773 |
Tolbert |
December 11, 1973 |
PRESSURE COMPENSATING VALVE MECHANISM
Abstract
A mechanism which, in one form, is useful as an unloading valve,
and in another form as a pressure compensating valve. It comprises
a valve plunger having a bypass position allowing all source fluid
entering an inlet port to flow to an outlet port, and having a feed
position compelling flow of source fluid to a feeder port in an
amount depending upon the extent the plunger is displaced from its
bypass position. The plunger is held in a feed position under
substantially strong force exerted thereon by a primary spring and
by pressure fluid from the feedback port. When the feedback port is
vented, the plunger moves to its bypass position under force
exerted thereon by pressure fluid from the inlet port, against the
force of the primary spring diminished by the force of a secondary
plunger spring.
Inventors: |
Tolbert; William N. (Pewaukee,
WI) |
Assignee: |
Koehring Company (Milwaukee,
WI)
|
Family
ID: |
22855421 |
Appl.
No.: |
05/228,020 |
Filed: |
February 22, 1972 |
Current U.S.
Class: |
137/115.15;
137/115.21; 91/451; 91/446; 137/596.13 |
Current CPC
Class: |
F15B
13/0417 (20130101); Y10T 137/263 (20150401); Y10T
137/87185 (20150401); Y10T 137/2615 (20150401) |
Current International
Class: |
F15B
13/00 (20060101); F15B 13/04 (20060101); F15b
011/02 () |
Field of
Search: |
;137/115,116.3,117,522,523,596.13 ;91/446,451 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Nilson; Robert G.
Claims
The invention is defined by the following claims:
1. A fluid flow controlling valve mechanism having a body with an
inlet and an outlet and feeder and feedback ports all opening to a
bore containing a fluid pressure responsive valve plunger, means to
impose the pressure of fluid at the inlet upon a first surface of
the plunger to urge it in one axial direction toward a normal
position at which pressure fluid entering the bore from the inlet
flows to the outlet in bypass relation to the feeder port, and
means by which the pressure of fluid at the feedback port can be
imposed upon a second surface of the plunger to move it in the
opposite axial direction toward a feed position compelling inlet
fluid to flow to the feeder port, characterized by:
A. a strong primary spring connected with the plunger to resist
movement thereof in said one axial direction, out of its feed
position;
B. a weaker secondary spring for the plunger;
C. and means connected with the secondary spring and rendered
operative by the plunger in the normal position of the latter for
effecting imposition of secondary spring force upon the plunger in
opposition to the force of the primary spring to thus enable the
plunger to be held in its normal position by inlet fluid at
substantially low pressure.
2. The fluid flow controlling valve mechanism of claim 1 further
characterized by:
A. said means which is connected with the secondary spring
comprising a piston which is movable in one direction from one
position to another position at which the secondary spring is
ineffective to oppose the primary spring;
B. and means for effecting motion of the piston to said other
position thereof in consequence of rise in fluid pressure at one of
said ports.
3. The fluid flow controlling valve mechanism of claim 1, wherein
the piston is confined between the primary and secondary springs
and can be held in said one position thereof by the primary
spring.
4. The fluid flow controlling valve mechanism of claim 3, wherein
pressure of fluid at the feedback port acts upon the piston to move
the same to said other position thereof at which it effects
relaxation of the secondary spring.
5. The fluid flow controlling valve mechanism of claim 1, further
characterized by:
A. said means which is connected with the secondary spring
comprising a piston confined between said springs and movable
axially toward said second surface on the plunger to relax the
secondary spring;
B. and means for effecting movement of the piston axially toward
said second surface on the plunger in consequence of rise in
pressure at the feedback port.
6. The fluid flow controlling valve mechanism of claim 5, further
characterized by:
A. the secondary spring being confined in the bottom portion of a
cylinder in which said piston operates;
B. and means providing a passageway communicating the bottom
portion of said cylinder with the feedback port.
7. The fluid flow controlling valve mechanism of claim 6, further
characterized by:
A. the plunger having a well in its end adjacent to said piston, in
which the primary spring is received;
B. the feedback port opening to said well;
C. and the piston having an axial passage therethrough
communicating said well with the bottom of the cylinder.
8. The fluid flow controlling valve mechanism of claim 7, further
characterized by:
A. the opposite ends of the piston being subjected to the pressure
of feedback fluid in said well and in the bottom portion of the
cylinder;
B. and the cylinder and piston having correspondingly larger
diameter portions at the end of the piston remote from the
plunger.
9. The fluid flow controlling valve mechanism of claim 1, further
characterized by:
A. said springs being connected in series with the plunger whereby
the force which they exert thereon in its normal position is equal
to the product of their ratings divided by the sum thereof;
B. and said means which is connected with the secondary spring
being operable in the feed position of the plunger, in response to
pressure of fluid at said feed back port, for substantially
nullifying the effect of said secondary spring upon the
plunger.
10. A pressure compensating valve mechanism having inlet, outlet,
feeder and feedback ports all opening to a bore containing a fluid
pressure responsive valve plunger, said mechanism being of the type
having provision for delivery of inlet fluid to one end of the bore
where it can exert force on the plunger and normally hold it in a
neutral position at which inlet fluid can flow to the outlet in
bypass relation to the feeder port, and wherein the plunger is
movable to an operating position limiting said bypass flow and
diverting inlet fluid to the feeder port, characterized by:
A. means opening to the other end of the bore providing a cylinder
coaxial with the bore;
B. a piston movable axially in the cylinder;
C. a primary compression spring reacting between the piston and the
plunger and operable in the neutral position of the latter to hold
the piston in engagement with an abutment in the outer end of the
cylinder;
D. a secondary compression spring confined in said outer end of the
cylinder and normally held in a loaded condition by the piston
under the force which the primary spring exerts thereon in said
neutral position of the plunger, whereby the force which the
primary spring exerts upon the plunger tending to move it toward
its operating position is substantially diminished except at times
when the secondary spring is relaxed;
E. and means communicating said outer end of the cylinder with the
feedback port so that the force which feedback fluid exerts upon
the piston will move the same in the direction to relax the
secondary spring.
11. A pressure compensating valve mechanism having a plunger
movable axially in a bore in the body of the mechanism and having
means to impose the pressure of fluid at an inlet upon a first
surface of the plunger to urge it axially in one direction in the
bore toward a normal position at which fluid entering the bore from
the inlet flows to an outlet in bypass relation to a feeder port,
and having a feedback port by which the pressure of feedback fluid
can be imposed upon a second surface of the plunger to move it in
the opposite axial direction toward a feed position compelling
inlet fluid to flow to the feeder port, characterized by:
A. a primary plunger spring connected with the plunger to
substantially strongly resist movement thereof in said one
direction out of its feed position;
B. a secondary plunger spring which is weaker than the primary
plunger spring;
C. means defining a cylinder;
D. a piston in the cylinder connected with the secondary spring and
operable in said normal position of the plunger to render the
secondary spring effective to impose force upon the plunger in
opposition to the force of the primary spring providing the piston
is in one axial position at which it engages an abutment on the
body, said piston being movable away from said abutment to a second
position at which it renders the secondary spring ineffective;
E. and means for effecting movement of the piston to its said
second position in consequence of rise in fluid pressure at one of
said ports.
12. The pressure compensating valve mechanism of claim 11 further
characterized by:
A. said primary and secondary springs being connected with the
plunger at opposite ends thereof;
B. and said piston being movable to its said second position in
consequence of rise in fluid pressure at the feeder port.
13. A pressure compensating valve mechanism having a plunger
operable in a feed position to regulate the rate at which pump
fluid flows from an inlet to a feeder port in accordance with
variations in the pressure differential between the feeder port and
a feedback port, and wherein the plunger is held in a normal
position displaced from its feed position under force which inlet
fluid exerts upon one surface thereof to allow inlet fluid to flow
to an outlet in bypass relation to the feeder port, characterized
by:
A. a strong primary plunger spring;
B. a weaker secondary plunger spring;
C. means connecting the primary spring with the plunger so that the
primary spring will substantially strongly resist movement of the
plunger out of its feed position;
D. means connecting the secondary spring with the plunger;
E. said last named means being rendered operative by the plunger,
in said normal position thereof, to impose the force of the
secondary spring upon the plunger in offsetting relation to the
force of the primary spring;
F. and said last named means being rendered operative in
consequence of fluid pressure at one of said ports, in the feed
position of the plunger, to nullify said force offsetting effect of
the secondary spring upon the plunger.
Description
This invention relates to controls for fluid motors, and has more
particular reference to bypass type pressure compensating valve
mechanisms such as are used conjunctionally with a closed center
control valve to govern the speed of a fluid motor.
In general, such pressure compensating valve mechanisms comprise a
fluid pressure actuatable valve plunger or spool which cooperates
with the valve element or main spool of the control valve to
regulate flow of pump fluid to the inlet of the control valve in
accordance with variations in the pressure differential between
pump output fluid and that motor port of the control valve through
which pump fluid is being supplied to the motor.
Ordinarily, when the spool of a closed center control valve is in a
neutral or hold position closing off the motor ports from both
supply and return passages, the pressure of pump output fluid
cannot be dissipated through the control valve. At that time, pump
output fluid acts upon one end of the plunger of the associated
pressure compensating valve, against an opposing spring force
acting upon the other end of the plunger, to hold the plunger in a
bypass position at which all of the pump output fluid can be
bypassed to an outlet port. However, when the main spool is in an
operating position directing source fluid to one of the motor
ports, fluid pressure corresponding to that which exists at said
motor port is fed back to the pressure compensating valve mechanism
and is imposed upon the other end of its plunger to act with the
spring force thereon in holding the plunger in a feed position at
which it maintains the drop or differential between pump output and
motor port pressures at a constant value. That value, of course, is
determined by the size of the variable flow restriction or orifice
provided by the valve spool and across which source fluid flows to
the selected motor port. Thus, bypass flow is reduced by the volume
of pressure fluid allowed to flow across the orifice of the valve
spool to the motor port in any given operating position of the
spool.
It is customary to provide such pressure compensating valve
plungers with a spring to strongly resist plunger movement out of
its feed position during operation. As is well known in the art,
precise control over the speed of the governed fluid motor is
impossible unless this spring force is great enough to assure the
desired quick response of the compensating plunger to the smallest
variations in pump output and/or feedback pressure under all
operating conditions.
While the importance of strongly biasing the pressure compensating
plunger toward its feed position was widely recognized, this
prerequisite nevertheless gave rise to a serious problem which
until now has defied solution. As stated, the compensating plunger
must be held in a bypass position at times when the control valve
element is in its neutral or hold position. The purpose of this, of
course, is to unload the pump just as an open center control valve
does when its valve element is in a neutral position.
However, it was heretofore possible to only partly unload the pump
in the neutral position of the control valve element. This was due
to the fact that at that time the pump had to generate an output
pressure of a magnitude which, when imposed upon the compensating
plunger, was capable of holding it in its bypass position against
the force of the strong plunger spring. As a result, not only was
considerable power wasted, but undesirable heating was an
unavoidable consequence of such unproductive effort on the part of
the pump.
With the foregoing in mind, it is a purpose of the present
invention to provide a compensating valve mechanism having a fluid
pressure actuatable plunger movable between feed and bypass
positions, and in the latter of which positions the plunger can be
held by an exceptionally low pump output pressure, while in the
feed position of the plunger, a desirably strong spring force is
exerted thereon to strongly resist return movement thereof toward
its bypass position.
More specifically, it is an object of the invention to provide a
pressure compensating valve mechanism wherein a substantially
strong primary spring bearing upon the plunger urges it toward its
feed position, and a weaker secondary spring connected with the
plunger normally acts to diminish the force of the primary spring
and thus assure a low pressure drop across the inlet and outlet
ports of the mechanism in the bypass position of the plunger.
In this respect, it is a further object of the invention to provide
a valve mechanism such as described above where the primary and
secondary springs normally act in series upon the plunger to only
lightly resist motion thereof toward its bypass position under
force of inlet fluid thereon, and wherein the force of feedback
fluid is utilized to render the secondary spring ineffective so
that the full force of the primary spring plus that of feedback
fluid acting on the plunger is then made available to resist
movement of the plunger out of its feed position.
Another object of the invention is to adapt the series-connected
spring concept mentioned in the preceding paragraph to bypass valve
mechanisms per se, as distinguished from pressure compensating
valve mechanisms.
In a preferred embodiment of the invention, a piston is confined
between the primary and secondary springs and is actuatable in a
direction to relax the secondary plunger spring in consequence of
subjection of the piston to feedback pressure from the governed
motor.
With these observations and objectives in mind, the manner in which
the invention achieves its purpose will be appreciated from the
following description and the accompanying drawings, which
exemplify the invention, it being understood that changes may be
made in the specific apparatus disclosed herein without departing
from the essentials of the invention set forth in the appended
claims.
The accompanying drawings illustrate several complete examples of
embodiments of the invention constructed according to the best
modes so far devised for the practical applications of the
principles thereof, and in which:
FIG. 1 is a diagrammatic view illustrating a fluid pressure
operated system embodying a pressure compensating valve mechanism
of this invention, showing the plunger thereof in its bypass
position;
FIG. 2 is a view of the pressure compensating valve mechanism of
FIG. 1, but showing the plunger thereof in a feed position;
FIG. 3 is a diagrammatic view of a pump unloading valve embodying
this invention; and
FIGS. 4 and 5 are diagrammatic views of a pressure compensating
valve mechanism of modified construction.
Referring now to the accompanying drawings, and particularly to
FIGS. 1 and 2 thereof, the numeral 5 designates a reversible fluid
motor, here shown as a double acting hydraulic cylinder having a
piston rod 6 by which the cylinder can be operatively connected to
a load (not shown). The cylinder is supplied with pressure fluid
from a pump P, at the dictate of a control valve 8. The control
valve has been shown by way of example as a single spool valve of a
conventional type. Its valve spool 9 is shiftable axially to
working positions at opposite sides of a neutral position (shown),
to direct output fluid from the pump through a supply passage 10 to
one or the other of a pair of motor ports 11, 12 connecting with
the opposite sides of the cylinder.
The control valve is of the closed center type having feeder
passage means comprising an upstream branch 14 that can be
considered as the inlet of the valve, and a downstream branch 15
which is communicable with the upstream branch 14 through the bore
16 in which the valve spool operates. The downstream branch 15 of
the feeder passage is communicated with the supply passage 10
through a load holding check valve 17.
The supply passage is of inverted U-shape to provide branches 18
and 19 which intersect the bore at locations inwardly adjacent to
the junctions of the bore with the service passages 11 and 12,
respectively. The valve spool is grooved to provide lands which
control communication of the motor ports 11, 12 with either the
supply passage branches 18, 19 or with exhaust passages 20, 21
which intersect the bore 16 at locations outwardly adjacent to the
junctions of the bore with the motor ports 11 and 12,
respectively.
A central land 22 on the spool is situated to close off
communication between the feeder passage branches 15, 16 in the
neutral or hold position of the valve spool shown. Throttle notches
23 and 24 in the left and right hand ends respectively, of the land
22 provide for adjusting the rate at which pressure fluid flows
from the upstream feeder branch 14 to the downstream feeder branch
15, upon shifting of the valve spool in either direction out of
neutral to a flow metering position short of a full operating
position of the spool.
In a right hand flow metering position of the valve spool 9,
pressure fluid entering the inlet 14 flows through throttle groove
23 and to the rod end of cylinder 5 via the feeder passage branch
15, check valve 17 and branch 18 of the supply passage 10 then in
communication with service passage 11. In a left hand flow metering
position of the spool, pressure fluid entering the inlet 14 flows
through throttle notch 24 and to the head end of the cylinder via
feeder branch 15, check valve 17 and supply branch 19 then in
communication with the service passage 12. In each case, the
non-selected service passage is communicated with its adjacent
exhaust passage branch 20 or 21 to conduct to the reservoir of the
system pressure fluid which is expelled from the cylinder 5.
The pressure compensating valve mechanism 25 of this invention is
connected in the system between the pump P and the control valve 8,
and its purpose is to maintain the flow of pressure fluid to the
cylinder at a constant rate determined by the metering position or
setting of the control spool 9. As is customary, the pressure
compensating valve is provided with a fluid pressure actuatable
plunger 26 which is sensitive to the pressure drop across the
orifice provided by either throttle notch 23, 24, and which
regulates the flow of pump fluid to the inlet 14 in accordance with
variations in said pressure drop from that value thereof which
exists when fluid flows to the motor at the desired rate.
The plunger 26 of the pressure compensating valve mechanism has
hollow opposite end portions, and it is slidably received in a bore
27 which is closed at each end. Five different passages open to the
bore at axially spaced zones. Reading from left to right, these
passages comprise a low pressure or return port 28, a feedback port
29, an outlet port 30, an inlet port 31, and a feeder port 32.
The return port 28 can comprise part of the exhaust passage 20; the
feeder port 32 is communicated with the upstream feeder branch 14
by means of a duct 33; and the feedback port 29 can be communicated
with either motor port 11, 12 through a shuttle valve in a
conventional way. In the single spool control valve shown, however,
the feedback port 29 is communicated with the branch 18 of the
supply passage 10 by means of a duct 34; or it can be connected to
feeder branch 15 with the same results. The feedback port 29 is
vented to the reservoir of the system when the control valve spool
9 is in its neutral position by means of an axial passage 35 in the
valve spool which communicates at one end with the supply passage
branch 18 through a radial passage 36 in the spool, and with the
exhaust branch 20 at its other end through a radial passage 37 in
the spool.
In the neutral position of the valve spool 9, seen in FIG. 1, the
land 22 on the spool closes off communication between the feeder
branches 14 and 15. Pump output fluid then entering the inlet port
31 of the compensating valve acts upon the right hand end of the
plunger 26 to hold the same in a bypass position at which all of
the incoming pump fluid flows to the outlet port via a
circumferential groove 26' in the plunger communicating ports 30
and 31. For this purpose, the pressure chamber 38 provided by the
right hand end of the bore 27 is communicated with the inlet port
31, as by a radial hole 39 in the hollow end portion of the
plunger. The hole 39 is at all times in direct communication with
the feeder port 32, and consequently also with the inlet port
31.
When the spool 9 of the control valve is moved to a full operating
position at either side of neutral, it provides unrestricted
communication between the feeder branches 14 and 15, and the
pressure of fluid in the supply passage 10 will be at substantially
the same valve as that of pressure fluid in whichever motor port 11
or 12 is then in communication with it. As is customary, this
pressure is fed back to the compensating valve mechanism and
imposed upon the left hand end of the compensating plunger 26 in
opposition to the force which pump output fluid in pressure chamber
38 exerts upon the right hand end of the plunger. A spring also
customarily urges the plunger toward its feed position.
According to this invention, the spring bias for the plunger is
provided by a pair of compression springs, namely, a substantially
strong primary spring 40 and a weaker secondary or auxiliary spring
41. Both of these springs are located at the left hand end of the
plunger, at opposite axial ends of a stepped piston 42. The primary
spring is confined in the well 43 provided by the hollow left hand
end portion of the plunger, between the bottom of the well and the
adjacent inner end portion 44 of the piston, which is smaller in
diameter than the well. A washer 45 is preferably interposed
between the piston and the primary spring. The piston is axially
slidable in a stepped cylinder 46 which is coaxial with the bore
27, and the small diameter inner end portion of which cylinder
opens to the bore.
One end portion of the secondary spring extends into a well 47 in
the larger diameter outer end portion 48 of the piston, and the
other end of the spring extends into a well 49 in a cap 50 which
closes the large diameter outer end portion of the cylinder. The
spring 41, of course, is confined between the bottoms of the wells
47 and 49.
The large diameter end portion of the cylinder 46 has greater axial
length than that of the piston portion therein, so as to provide a
stop 51 which cooperates with the inner end 52 of the cap 50 to
define the limits of axial sliding motion of the piston.
It should here be observed that the piston is circumferentially
reduced between its large and small diameter ends, as at 53. The
space thus provided between the large diameter end of the piston
and the inner stop 51 is vented through a hole 54 in the wall of
the cartridge 55 in which the piston is housed. The inner end of
this cartridge also provides a stop to define the neutral or full
bypass position of the compensating plunger 26.
It is important to note that the piston is formed with a passage 57
which extends axially therethrough, so that the fluid pressures in
the spring chambers at the opposite ends of the piston 42 can be
equalized.
In the neutral or hold position of the control valve spool 9 seen
in FIG. 1, the compensating plunger occupies a bypass position at
which all of the pump output fluid entering its inlet port 31 flows
to the outlet port 30, in bypass relation to the feeder port 32.
This condition is brought about by reason of the fact that the
primary and secondary spring chambers 43 and 47 are vented to the
exhaust passage 20 in the control valve through a hole 58 in the
side wall of the chamber 43, which hole opens to the feedback port
29 and is always in communication therewith. In this connection, it
should be recalled that the feedback port 29 is communicated with
the exhaust passage 20 via the duct 34, branch 18 of the supply
passage, and the passages 35, 36 and 37 in the control spool 9 in
the neutral position of the latter.
When the spring chambers are vented in this fashion, the pressure
of pump output fluid at the inlet port 31, acting upon the right
hand end of the compensating plunger 26 in chamber 38, is easily
able to hold the plunger in its bypass position against the action
of the primary and secondary plunger springs 40 and 41,
respectively. It should be appreciated at this point that the
primary and secondary springs can be said to act in opposition to
one another. The total force which they exert upon the plunger
tending to move it out of its bypass position is actually far less
than the force which the primary spring alone would be capable of
exerting upon the plunger. In fact, because the primary and
secondary springs 40 and 41 are connected in series with the
plunger, the total force which they exert thereon is computed in
the same way as the resistance of electrical resistors in parallel.
Thus, if the ratings of the primary and secondary springs are
represented by R1 and R2, respectively, the total force RT can be
found using the formula
RT = (R1 .times. R2/R1 + R2)
Merely by way of example, if R1 equals 100 psi and R2 equals 50
psi, the total force RT which can be exerted on the plunger is 5000
psi divided by 150 psi, or only slightly over 33 psi. Thus it will
be seen that the pressure drop across the inlet and outlet ports 31
and 30 will be held to a very desirable low value whenever the
control valve spool is in its neutral position and the compensating
plunger 26 is in a corresponding neutral or bypass position such as
seen in FIG. 1.
The spring bias acting upon the compensating plunger becomes much
greater, as is essential, whenever the control spool 9 is actuated
to either working position at which fluid at high pressure flows
from the supply passage 10 to the selected motor port 11 or 12.
Assuming that the control spool is moved to the left, pump fluid
will then flow from the feeder passage 14 to motor port 12 via
throttle groove 24, feeder branch 15, check valve 17, and branch 19
of the supply passage. A metered amount of pump fluid then flows
into the head end of cylinder 5 to effect extension of its piston
rod 6.
The pressure of fluid then in the cylinder head will be transmitted
to spring chambers 43 and 47 via supply passage 10, duct 34,
feedback port 29, hole 58 in the compensating plunger, and axial
passage 57 in piston 42.
Feedback fluid then flows into chamber 47 and moves the piston to
the right to its limit of motion seen in FIG. 2, so as to then
relax the secondary spring 41. When that occurs, the secondary
spring 41 becomes ineffective, and the full force of the primary
spring is then exerted on the compensating plunger, without
opposition from the secondary spring. The position of the
compensating plunger, in all operating positions of the control
valve spool, will then depend upon pump output pressure as imposed
upon the right hand end of the plunger, and upon the combined
forces of feedback fluid and of the primary spring 40 acting upon
the left hand end of the plunger.
At any given setting of the main valve spool, the position of the
compensating plunger will be automatically adjusted to effect
regulation of fluid pressure at the feeder port 32 in accordance
with variations in pump output pressure at port 31 and/or in
response to variations in feedback pressure; and the plunger will
be moved in response to such variations in pressure in the
direction to compensate for the changed condition and thereby
maintain the pressure at port 32 at a constant value. That is to
say, for example, that the plunger will be caused to move to the
right to increase fluid pressure at the feeder port 32 in
consequence of a rise in pressure at the feedback port 29; and it
will be caused to move to the left to decrease the feeder port
pressure as a consequence of a decrease in pressure at the feedback
port 29.
It will be seen, therefore, that in the single spool control valve
illustrated, the compensating plunger functions to maintain
constant pressure at the feeder port 32 by varying the degree of
communication between the pump port 31 and the outlet port 30. In a
plural spool valve, however, the pressure compensating plunger will
function to maintain feeder port pressure constant by varying the
degree of communication between the feeder port and the pump port
31 at times when port 30 is in use as a high pressure carry-over
port for a downstream control valve and the spool of the latter is
operating a fluid motor at a greater pressure than that being
operated by the upstream spool.
Actuation of the control spool 9 to a full flow position, of
course, will effect complete close-off of the outlet port 30 from
the inlet port 31, so that all pump fluid will then be compelled to
flow to the work cylinder.
As soon as the control spool 9 is returned to its neutral position,
of course, the feeder passage 14 is blocked and the feedback port
29 is vented. As a result, pump output fluid flows into chamber 38
at the right hand end of the plunger and promptly moves the same to
its bypass position at which the secondary spring 41 is again
effective to oppose the action of the primary spring 40.
From the above, it will be seen that the spring force tending to
move the compensating plunger out of its bypass position is
desirably low, while a much greater spring force resists movement
of the plunger out of its feed position at times when pressure
fluid is being directed to one end or the other of the cylinder 5.
This last assures the desired precise control over the speed at
which the cylinder operates.
FIG. 3 illustrates how the series spring concept described above
can be used to advantage in a pump unloading valve. The unloading
valve is similar in most respects to the pressure compensating
valve described earlier, although it can be made somewhat simpler.
Thus, its plunger 65 comprises a cup shaped member comparable to
the large diameter end portion of the compensating plunger 26 at
the left hand end thereof, and having its closed end 66 movable
into and out of engagement with a seat 67 provided by a short
portion of the bore 68 in which the valve plunger is axially
slidably received. Reading from left to right, the unloading valve
is also provided with a reservoir port 69, a feedback port 70, an
outlet port 71, an inlet port 72, and a feeder port 73. While the
inlet and outlet ports open to the bore 68 at axially opposite
sides of the bore portion 67, the feeder port 73 can be directly
communicated with the inlet port 72, to form a part thereof, as
shown.
The reservoir port 69 opens to a counterbore 74 in which is secured
the cartridge 75 containing the stepped piston 76. The piston, of
course, is slidable axially in a stepped cylinder provided in the
interior of the cartridge, as before; and the primary and secondary
plunger springs 77 and 78, respectively, are again located at
axially opposite ends of the piston, with the primary spring
extending into the hollow interior 79 of the valve plunger 65.
In the present case, the plunger 65 is shown as extending outwardly
of the bore 68 into the counterbore 74, to engage the inner end of
the cartridge 75, which thus defines the fully open or bypass
position of the plunger. A snap ring 80 confined in a groove in the
exterior of the plunger is engageable with the bottom 81 of the
counterbore to define the closed position of the plunger at which
its right hand end is received within the seat defining bore
portion 67 to block communication between the inlet port 72 and the
outlet port 71.
As before, a hole 83 in the side wall of the plunger 65
communicates with the feedback port 70 and provides for entry of
feedback fluid into the interior of the plunger to effect
pressurization of the chambers at the opposite ends of the piston
76.
The pump unloading valve operates in substantially the same way as
the pressure compensating valve. When the spool of the associated
control valve is in neutral position, the feedback port 70 is
vented and the feeder passage is closed off. Accordingly, pump
output fluid entering the inlet port 72 acts upon the closed end 66
of the plunger and moves the same to its fully open position shown,
against the substantially light resistance of the serially
connected primary and secondary springs 77 and 78. All of the pump
fluid entering port 72 then flows through outlet port 71 and back
to the reservoir, with but slight pressure drop between ports 71
and 72.
As soon as the spool of the associated control valve is shifted to
an operating position allowing pump output fluid to flow through
the feeder passage to the selected motor port, the pressure of
fluid at the latter port is transmitted to the port 70 of the
unloading valve to effect pressurization of both spring chambers,
at the opposite ends of the piston 76. Pressure fluid then flowing
into the secondary spring chamber moves piston 76 toward the
plunger and away from the secondary spring 78 so as to relax the
secondary spring 78, while pressure fluid in the primary spring
chamber acts directly upon the plunger in the direction to move it
toward its closed position. The plunger may then move into the bore
portion 67 under this combination of forces acting thereon, to
close off the outlet port 71 from the inlet port 72 and thereby
constrain pump output fluid entering the inlet port to flow through
the feeder port for passage to the controlled cylinder.
It is also possible for the valve plunger 65 to occupy a partially
closed position allowing some of the pump output fluid to flow to
the outlet port and constraining the remainder to flow to the
feeder port. Such a situation can arise from placement of the
control valve spool in a metering position on the order of that
described earlier, so that the plunger 65 would then respond to
variations in the pressure drop across the orifice provided by the
throttle groove through which the work cylinder was supplied with
pump output fluid.
It is also believed to be obvious to those skilled in the art that
the feedback port 70 of the unloading valve could be communicated
with the motor ports of the associated control valve through a
shuttle valve arrangement in a more or less conventional way.
FIG. 4 illustrates another version of the pressure compensating
valve mechanism of this invention. Its plunger 89, shown in bypass
position, is again provided with a circumferential groove 90, and
it is reciprocable endwise in a bore 91 whose opposite ends are
closed.
In the pump unloading position shown, the plunger groove
communicates a pump inlet port 93 with an outlet port 94 to cause
pump output fluid to bypass a feeder port 95. The plunger occupies
this unloading or bypass position when the spool of the associated
control valve is in its neutral position closing off flow
therethrough of pump output fluid from the feeder port and venting
the left hand end of the bore 91 through a feedback port 96
therein.
In the feed position of the plunger 89, its left hand land can
limit or even close off flow of pump output fluid to the outlet
port from the inlet port to thereby divert a regulatable volume of
pump output fluid to the feeder port 95 through the groove 90. The
plunger moves to its feed position in consequence of actuation of
the spool of the associated control valve to an operating position
allowing for flow of pump output fluid therethrough from the feeder
port and pressurizing the feedback port 96 in the manner described
earlier.
Coaxial wells 97 and 98 provide spring chambers in the opposite
ends of the plunger. A secondary spring 100 situated in the well 98
acts upon a piston 101 in said well to urge the piston toward one
limit of motion at which its outer end abuts the adjacent end of
the bore 91. In that limit of motion, the piston may hold the
secondary spring lightly loaded. The piston 101 can be moved
inwardly into the well 98 by the pressure of fluid in the chamber
102 provided by the right hand end portion of the bore to further
compress the secondary spring. It is for this reason that the well
98 is at all times vented to a reservoir port 103, through a
passageway 104 in the compensating plunger.
The secondary spring 100 tends to move the compensating plunger 89
to the left, toward its bypass or pump unloading position. It acts
upon the plunger in opposition to a primary spring 105 which is
situated in the well 97 in the outer end of the compensating
plunger 89 and tends to move the same toward its feed position.
While the secondary spring thus opposes the primary spring, the
latter is slightly stronger and tends to normally move the plunger
toward the right hand end of the bore 91, to its feed position.
The pressure chamber 102 in the right hand end of the bore 91 is at
all times communicated with the feeder port 95, and hence with the
inlet port 93, via a passageway 106 in the compensating
plunger.
The piston 101 is provided with a stem 107 which projects coaxially
into the secondary spring and toward the bottom of the well 98,
with which it can engage to define the inner limit of motion of the
piston relative to the compensating plunger.
In operation, the compensating plunger will be normally held in its
bypass or pump unloading position as long as the spool of the
associated control valve remains in its neutral position venting
the feedback port 96 and blocking flow of pump output fluid through
the feeder port 95. At that time, pump output fluid flows directly
to the outlet port 94 in bypass relation to the feeder port, and
its pressure is manifested in the chamber 102 at the right hand end
of the compensating plunger. The substantially small force which
pump output fluid is able to exert upon the right hand end of the
plunger, when added to the greater force of the secondary spring
100, produces a total force in excess of that exerted upon the
plunger by the primary spring 105. Hence, the plunger is held in
its bypass position mainly by reason of the force of the secondary
spring, which greatly counteracts the force of the primary
spring.
Thus it will be seen that again the force of the primary spring is
greatly diminished in the bypass position of the plunger, and that
the pressure drop between the inlet port 93 and outlet port 94 will
be held at a desirably low value.
As soon as the spool of the associated control valve is actuated to
an operating position, feedback fluid from the supply passage flows
into the port 96 and the compensating plunger is moved to the right
thereby to limit or even close off fluid flow to the outlet port
94, and to open up the feeder port 95 for the flow of pump fluid
therethrough to the selected motor port of the associated control
valve.
As the pressure rises in feeder passage 95, as it must in order to
overcome the load on the governed work cylinder, the pressure in
chamber 102 rises accordingly and forces piston 101 to its inner
extreme of motion at which its stem engages the bottom of the well
98. The piston then can be said to become part of the plunger, and
the effect of the primary spring is nullified. In the feed position
of the compensating plunger, seen in FIG. 5, therefore, the primary
spring acts upon the plunger without opposition from the secondary
spring and provides the desirably strong bias essential to precise
control over the rate at which pump output fluid flows to the
feeder port 95.
As before, of course, the compensating plunger will respond to
variations in the pressure differential between fluid at the inlet
port 93 and feedback fluid at the feedback port 96, and effect
regulation of fluid flow to the feeder port in accordance with such
variations.
From the foregoing description, together with the accompanying
drawings, it will be appreciated that this invention makes possible
the provision of a pressure compensating valve having a plunger
which can be maintained in a bypass position under relatively light
force, and in which a substantially strong spring force is made
available to resist movement of the plunger out of its feed
position.
Those skilled in the art will appreciate that the invention can be
embodied in forms other than as herein disclosed for purposes of
illustration.
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