Note: Descriptions are shown in the official language in which they were submitted.
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BACKGROUND OF THE INVENTION
This invention relates generally to load responsive
fluid control valves and to fluid power systems incorporating
such valves, which systems are supplied by a single fixed or
variable displacement pump. Such control valves are equipped
with an automatic load responsive control and can be used in
a multiple load system, in which a plurality of loads is
individually controlled under positive and negative load
conditions by separate control valves.
In more particular aspects this invention relates to
direction and flow control valves capable of controlling simul-
taneously a number of loads under both positive and negative
load conditions.
In still more particular aspects this invention
relates to direction and flow control valves capable of con-
trolling ~ multaneously multiple positive and negative loads,
which while controlling a negative load interrupt pump flow
to the motor providing the motor inlet with fluid from the
pressurized system exhaust.
Closed centre load responsive fluid control valves
are very desirable for a number of reasons. They permit load
control with reduced power losses and therefore, increased
system efficiency and when controlling one load at a time
provide a feature of flow control irrespective of the varia-
tion in the magnitude of the load. Normally such valves
include a load responsive control, which automatically main-
tains pump discharge pressure at a level higher, by a constant
pressure differential, than the pressure required to sustain
the load. A variable orifice, introduced between pump and
load, varies the flow supplied to the load, each orifice area
corre~ponding to a different flow level, which is maintained
constant irrespective of variation in magnitude of the load.
The application of such a system is, however, limited by one
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basic system disadvantage.
Normally in such a system the load responsive valve
control can maintain a constant pressure differential and
therefore constant flow characteristics when operating only
one load at a time. With two or more loads, simultaneously
controlled, only the highest of the loads will retain the flow
control characteristics, the speed of actuation of lower loads
varying with the change in magnitude of the highest load. A
fluid control valve for such a system is shown in U. S. Patent
#3,488,953 issued to Haussler.
This drawback can be overcome in part by the
provision of a proportional valve as disclosed in my U. S.
Patent #3,470,694 dated October 7, 1969 and also in U. S.
Patent #3,455,210 issued to Allen on July 15, 1969. However,
while these valves are effective in controlling positive loads
they do not retain flow control characteristics when control-
ling negative loads, which instead of taking supply the energy
to the fluid system and hence the speed of actuation of such
a load in a negative load system will vary with the magnitude
of the negative load. Especially with socalled overcentre
loads, where a positive load may become a negative load, such
a valve will lose its speed control characteristics in the
negative mode.
This drawback can be overcome by the provision of
a load responsive fluid control valve as disclosed in my U. S.
Patent #3,744,517 issued July 10, 1973 and my U. S. Patent
#3,882,896 issued May 13, 1975. However, while these valves
are effective in controlling both positive and negative loads,
with pump pressure responding to the highest pressure of a
sy~tem load being controlled, they utilize a controlling ori-
fice located in the motor exhaust during negative load mode
of operation and therefore control the fluid flow out of the
fluid motor ~hese valves al~o during control of negative
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loads supply the motor inlet with throttled down fluid from
the pump circuit, therefore using flow from the pump, while
controlling a negative load. In certain fluid power control
systems it is preferable, while controlling a negative load,
to supply fluid to the motor inlet from the motor exhaust
circuit instead of using pump capacity.
These drawbacks can be overcome in part by provision
of fluid control valves as disclosed in U. S. Patent #3,807,447
issued to Masuda on April 30, 1974. However, while these
valves utilize actuator exhaust fluid for actuator inlet flow
requirement when controlling negative loads and also utilize
a controlling orifice located between the pump and the actuator
while controlling positive and negative loads they regulate
actuator inlet pressure by bypassing fluid to a down stream
load circuit. Masuda's valves and their proportional control
system are based on series type circuit in which excess fluid
flow is successively diverted from one valve to the other and
in which loads arranged in series determine the system pressure.
In such a system flow to the last valve operating a load must
be delivered through all of the bypass sections of all of the
other system valves, resulting in fluid throttling loss-.
These valves are not adaptable to simultaneous control of mul-
tiple loads in parallel circuit and they do not provide system
load control pressure signal to the pump flow control mechanism.
SUMMARY OF THE INVENTION
It is therefore a principal object of this invention
to provide improved load responsive fluid direction and flow
control valves which block system pump from motor inlet and
supply it with system exhaust flow when controlling negative
loads, while transmitting control signals to system pump to
maintain the pressure of the system pump higher, by a constant
pressure differential, than the highest pressure of the system
posi~ive load being controlled.
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Another object of this invention is to provide load
responsive fluid direction and flow control valves, which
load responsive fluid direction and flow control valves are
provided with a pressurized exhanst manifold, flow from which
supplies the inlet flow requirements of motors controlling
negative loads.
It is a further object of this invention to provide
load responsive fluid direction and flow control valves which
retain their control characteristics during control of positive
loads, while responding to a pressure differential developed
across a variable orifice located between the pump and the
actuator and which retain their control characteristics during
control of negative loads while responding to a pressure dif-
ferential developed across a variable orifice located between
actuator and exhaust manifold.
Briefly the foregoing and other additional objects
and advantages of this invention are accomplished by providing
a novel load responsive fluid control system for use during
proportional simultaneous control of multiple positive and
negative loads. A system pump is controlled in respect to
pressure signal transmitted from system valves, corresponding
to the highest system load pressure. Exhaust circuit of the
system is pressurized, the exhaust flow being used to provide
inlet flow requirements of motors controlling negative loads.
Valve controls during control of positive and negative loads
respond to pressure differentials developed across variable
orifices in the actuator inlet and outlet.
Additional objects of this invention will become
apparent when referring to the preferred embodiments of the
invention as shown in the accompanying drawing and described
in the following detailed description.
DESCRIPTION OP THE DRA'~ING
Fig. 1 i~ a longitudinal sectional view of an
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embodiment of a flow control valve having a positive load
control responsive to actuator upstream pressure differential
and negative load controls responsive to actuator down ~tream
pressure differential for use in load responsive fluid control
system, with lines, system flow control, system pump, second
load responsive valve, exhaust relief valve and system reser-
voir shown diagramatically.
DESCRIPTION OF THE PREFERRE~ EMBODIMENT
Referring now to Fig. 1, an embodiment of a flow
control valve, generally designated as 10, is shown interposed
between diagramatically shown fluid motor 11 driving load L
and a pump 12 of a fixed displacement or variable displacement
type driven through a shaft 13 by a prime mover not shown.
Similarly, a flow control valve 14, indentical to
flow control valve 10, is interposed between a diagramatically
shown fluid motor 15 driving a load W and the pump 12. Fluid
flow from the pump 12 to flow control valves 10 and 14 is
regulated by a pump flow control 16. If pump 12 is of a fixed
displacement type pump flow control 16 is a differential
pressure relief valve, which in a well known manner, by bypas-
sing fluid from the pump 12 to a reservoir 17, maintains dis-
charge pressure of pump 12 at a level, higher by a constant
pressure differential, than load pressure developed in fluid
motor 11 or 15. If pump 12 is of a variable displacement type
pump flow control 16 is a differential pressure compensator,
well known in the art, which by changing displacement of pump
12 maintains discharge pressure of pump 12 at a level, higher
by a constant pressure differential, than load pressure devel-
oped in fluid motor 11 or 15.
The flow control valve 10 is of a fourway type and
has a housing 18 provided with a bore 19 axially guiding a
valve spool 20. The valve spool 20 is equipped with lands
21, 22, 23 and outlet metering land 24 which in neutral
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position of the valve spool 20, as shown in Fig. 1 isolate a
fluid supply chamber 25, load chambers 26 and 27, outlet
chambers 28 and 29 and fluid unloading chamber 30. The meter-
ing land 24 is equipped with metering slots 31 and 32, which
upon displacement of the metering land 24, from neutral posi-
tion in either direction, connects for fluid flow the outlet
chamber 28 with the unloading chamber 30. The unloading cham-
ber 30 is connected through slots 34, of a negative load
control spool 35, to the exhaust chamber 36. The negative
load control spool 35 having slots 34, provided with throttling
edges 37, projects into control space 38 and is biased towards
position, as shown, by spring 39. The negative load control
spool 35 is provided with stop 40 limiting its displacement
against surface 41. The exhaust chamber 36 in turn is connec-
ted through exhaust line 42, an exhaust relief valve, generally
designated as 43, and line 44 to the reservoir 17.
The pump 12 through its discharge line 45 and load
check valve 46 is connected to a fluid inlet chamber 47.
Similarly discharge line 45 is connected through load check
valve 48 with the inlet chamber of the fluid control valve 14.
The control bore 49 connects the fluid inlet chamber 47 with
the fluid supply chamber 25. The control spool 50, axially
slidable in the control bore 49, projects on one end into
space 51 connected to the fluid supply chamber 25 by passages
52 and 53 and restriction orifice 54 and abuts against piston
55 defining space 56 connected by passage 57 to outlet chamber
29. The control spool 50 on the other end projects into con-
trol space 58 which is connected by passage 59 with the posi-
tive load sensing ports 60 and 61 and through leakage orifice
62 and line 63 to a second exhaust chamber 64 and to upstream
pressure of the exhaust relief valve 43. Similarly control
space and leakage orifice of the control valve 14 is connected
by line 65 to the upstream pressure of exhaust relief valve
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34. The control spool 50 is provided with slots 66 termina-
ting in throttling edges 67 and a sealing land 69 equipped
with control surface 68, isolating the control space 58, the
supply chamber 25 and the second exhaust chamber 64. The
control spool 50 is biased by a control spring 70 towards
position, in which slots 66 connect the fluid supply chamber
25 with the fluid inlet chamber 47. The control spool 50 with
its sealing land 69 having a control surface 68 isolates, in
the position as shown in Fig. 1, the fluid supply chamber 24
from the second fluid exhaust chamber 64. Displacement of
the sealing land 69 from right to left cross-connects through
control surface 68 the fluid supply chamber 25 and the second
fluid exhaust chamber 64, the maximum displacement of the
control spool 50 being limited by surface 71.
If the pump 12 is of a fixed displacement type ex-
cess pump flow from the differential pressure relief valve or
pump flow control 16 is delivered through line 72 to the ex-
haust line 42, which communicates with the exhaust chamber 36,
the second exhaust chamber 64, a bypass check valve 73, the
exhaust relief valve 43 and through line 65 with all of the
exhaust passages of the flow control valve 14. The bypass
check valve 73 is interposed between exhaust line 42 and the
fluid supply chamber 25.
Positive load sensing ports 60 and 61, located be-
tween load chambers 26 and 27 and the supply chamber 25 and
blocked in neutral position of valve spool 20 by land 22, are
connected through signal passage 74, a check valve 75 and
signal line 76 to the pump flow control 16. In a similar
manner positive load sensing ports of flow control valve 14
are connected through line 77, a check valve 78 and signal
line 76 to the pump flow control 16.
The exhau~t relief valve, generally designated as
43, interposed between combined exhau~t circuits of flow
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control valves 10 and 14 including bypass circuit of pump 12
and reservoir 17, is provided with a throttling member 79
biased by a spring 80 towards engagement with seat 81.
The land 22 of the valve spool 20 is equipped with
signal slots 82 and 83 located in the plane of positive load
sensing ports 60 and 61 and metering slots 84 and 85 which in
a well known manner can be circumferentially spaced in respect
to each other and in respect to the signal slots 82 and 83.
Signal slots 82 and 83, in a well known manner, can be substi-
tuted by end surfaces of land 22. A suitable device is pro-
vided to prevent relative rotation of the spool 20 in respect
to bore 19.
The preferable sequencing of the control spool 50 is
such that when moved from right to left, when throttling edges
67 close communication between the inlet chamber 47 and the
supply chamber 25, control surface 68 is positioned at the
point of opening communication between the supply chamber 25
and the second exhaust chamber 64. Further movement of the
control spool 50 from right to left will gradually establish
full flow communication between the second exhaust chamber 64
and the supply chamber 25.
The sequencing of the lands and slots of valve spool
20 preferably is such that when displacedin either direction
from its neutral position, as shown in Fig. 1, one of the load
chambers 26 or 27 is first connected by the signal slot 82 or
83 to the positive load sensing port 60 or 61 while load
chambers 26 and 27 are still isolated from the supply chamber
25 and the outlet chambers 28 and 29. Further displacement
of the valve spool 20 from its neutral position connects load
chamber 26 or 27 through timing surface 86 or 87 with outlet
char~er 28 or 29, while land 22 still isolates the supply
chamber 25 from load chambers 26 and 27 and metering land 24
~till isolate~ outlet chambers 28 and 29 from the unloading
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chamber 30. Still further displacement of valve spool 20 will
connect load chamber 26 or 27 through metering slots 84 or 85
with the fluid supply chamber 25 while metering land 24 will
connect through metering slots 31 or 32 outlet chambers 28 and
29 with the unloading chamber 30.
As previously described the pump flow control 16, in
a well known manner, will regulate fluid flow delivered from
pump 12 to discharge line 45, to maintain the pressure in dis-
charge line 45 higher, by a constant pressure differential,
than the highest load pressure signal transmitted through the
check valve system to the signal line 76. Therefore with valve
spools of flow control valves 10 and 14 in their neutral posi-
tion blocking positive load sensing ports 60 and 61, signal
pressure input to the pump flow control 16 from the signal
line 76 will be at minimum pressure level.
With pump 12 of a fixed displacement type started
up the pump flow control 16 will bypass through line 72,
exhaust line 42, the exhaust relief valve 43 and line 44 all
of pump flow to the system reservoir 17 at minimum pressure
level equivalent to preload in the spring 80, while automati-
cally maintaining pressure in discharge line 45 at a constant
pressure, higher by a constant pressure differential, than
pressure in signal line 76 or pressure in exhaust line 42.
Therefore all of pump flow is diverted by the pump flow control
16 to the low pressure exhaust circuit, as previously described,
without being used by flow control valves 10 and 14. Since
signal line 76 is connected by passage 59 with control space
58, which in turn is connected through leakage orifice 62 and
line 63 to upstream of exhaust relief valve 43, the bypass
pressure in the discharge line 45 will be higher, by a constant
pressure differential, than the pressure in exhaust line 42,
which equals the pre~sure setting of the exhaust relief valve
43. ~hi~ pump bypass pressure transmitted through passages
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52 and 53 and restriction orifice 54 to space 51 reacts on
the cross-sectional area of control spool 50 and against the
bias of control spring 70 moves the control spool 50 from
right to left, closing with throttling edges 67 the passage
between the inlet chamber 47 and the supply chamber 25. Supply
chamber 25 is connected through bypass check valve 73 with
pressure existing in exhaust line 42. The pressure setting of
exhaust relief valve 43 is selected to provide the necessary
pressure drop through metering slots 84 and 85 to maintain
load chamber 26 or 27 at above atmospheric pressure.
With pump 12 of a variable displacement type, under
working conditions, minimum flow to the system exhaust mani-
fold, composed of line 72, 65, exhaust line 42 and exhaust
pressure relief valve 43, may have to be diverted from the
pump 12, to maintain the system exhaust manifold pressurized.
A pressure reducing type regulator can be used, which upon
system exhaust manifold pressure dropping below the setting of
the exhaust pressure relief valve 43, will throttle some of
the pump discharge flow and supply it to the exhaust manifold,
to maintain it at a certain preselected minimum pressure level.
Assume that the load chamber 26 is subjected to a
positive load. The initial displacement of the valve spool 20
to the right will connect the load chamber 26 through signal
slot 82 with positive load port 60, while lands 21, 22 and 23
still isolate the supply chamber 25, load chambers 26 and 27
and outlet chambers 28 and 29. As previously described posi-
tive load signal, transmitted from positive load sensing port
60, through signal passage 74, check valve system and signal
line 76 to the pump flow control 16 will increase the pressure
in di~charge line 45 to a level, which is higher by a constant
pressure differential than the load pressure signal. The load
pre~ure, transmitted through passage 59 to control space 58,
will move the positive load control spool 50 to the right,
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opening through slots 66 communication between the inlet cham-
ber 47 and the supply chamber 25. Communication will be main-
tained between the supply chamber 25 and the inlet chamber 47,
as long as the pump flow control 16 maintains a constant pres-
sure differential between the pump discharge pressure and the
positive load pressure.
Further displacement of the valve spool 20 to the
right will connect through timing surface 87 the load chamber
27 with outlet chambers 28 and 29, while land 22 still isolates
the load chamber 26 from the supply chamber 25 and the metering
land 24 still isolates the outlet chamber 28 from the unloading
chamber 30. Since the load chamber 27 is subjected to low
pressure no change in position of the negative load control
spool 35 will take place.
Still further displacement of the valve spool 20 to
the right will connect the load chamber 26, through metering
slot 84, with the supply chamber 25 and will also connect
through metering slot 32 the outlet chamber 28 with the unload-
ing chamber 30. In a manner as previously described, the pump
flow control 16 will maintain a constant pressure differential
across orifice, created by displacement of metering slot 84,
the flow into the load chamber 26 being proportional to the
area of the orifice and therefore displacement of the valve
spool 20 from its neutral position and independent of the mag-
nitude of the load L.
Assume that while controlling positive load L through
the flow control valve 10, a higher positive load W is actuated
through the flow control valve 15. Higher load pressure signal
from the flow control valve 15 will be transmitted through the
check valve system to the pump flow control 16, which will now
maintain ~y~tem pre~sure, higher by a constant pressure differ-
ential, than pre~sure generated by po~itive load W. In a
manner a~ previou~ly de~cribed, the presRure drop through
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metering slot 84 will increase, therefore increasing the
pressure differential between space 51 and control space 58
The positive load control spool 50 will move into its modula-
ting position, throttling with throttling edges 67 the fluid
flowing from the inlet chamber 47 to the supply chamber 25,
to maintain a constant pressure differential between the supply
chamber 25 and the load chamber 26, thus controlling fluid flow
through metering slot 84.
Assume that the load chamber 27 is subjected to a
negative load L and that the valve spool 20 is displaced from
its neutral position to the right while, as previously descri-
bed, the positive load control spool 50 is maintained by the
pump standby pressure in a position blocking communication
between the inlet chamber 47 and the supply chamber 25. Ini-
tial displacement of the valve spool 20 will connect through
signal slot 82 the load chamber 26 with the positive load sen-
sing port 60. Since the load chamber 26 is subjected to low
pressure neither the pump flow control 16 nor the positive
load control spool 50 will react to it.
Further displacement of valve spool 20 will connect
negative load pressure from load chamber 27 with outlet cham-
bers 29 and 28, while the metering land 24 still isolates the
outlet chamber 28 from the unloading chamber 30. Negative
load pressure, from the outlet chamber 28 will be transmitted
through passage 88 to control space 38, where reacting on the
cross-sectional area of the negative load control spool 35
will move it against the bias of the spring 39, all the way to
the left, blocking communication between the unloading chamber
30 and the exhaust chamber 36. The negative load pressure
from the outlet chamber 29 will be transmitted through passage
57 to space 56 where, reacting on the cross-sectional area of
the pi~ton 55, will generate a force moving the control ~pool
50, again~t biasing force of control spring 70, all the way to
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the left, with sealing land 69 coming in contact with surface
71. The control spool 50 will be maintained in this position
by the piston 55 during control of negative load, isolating
with throttling edges 67 the inlet chamber 47 from the supply
chamber 25 and establishing full flow communication, by dis-
placement of control surface 68 of the sealing land 69, between
the second exhaust chamber 64 and the supply chamber 25.
Further displacement of valve spool 20 to the right
will connect through metering slot 32 the outlet chamber 28
with the unloading chamber 30, while also connecting through
metering slots 84 the load chamber 26 with the supply chamber
25. Since the unloading chamber 30 is isolated by position of
the negative load control spool 35, the pressure in the unload-
ing chamber 30 will begin to rise, until it will reach a level,
at which force generated on the cross-sectional area of the
negative load control spool 35, by the pressure in control
space 38, will equal the sum of the force generated on the
same cross-sectional area by the pressure in the unloading
chamber 30 and the biasing force of the spring 39. At this
point the negative load control spool 35 will move from left
to right into a modulating position, in which fluid flow from
the unloading chamber 30 to the exhaust chamber 36 will be
throttled by the throttling edges 37, to automatically main-
tain a constant pressure differential, equivalent to biasing
force of the spring 39, between the outlet chamber 28 and the
unloading chamber 30. Since during control of negative load
a constant pressure differential is maintained across the
orifice, created by the displacement of metering slot 32, by
the throttling action of negative load control spool 35, fluid
flow through metering slot 32 will be proportional to the
di~placement of the valve spool 20 and constant for each spe-
cific position o metering slot 32, irrespective of the change
in the magnitude of the negative load L.
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As previously described during control of negative
load the control spool 50 will be maintained by the piston 55
in a position, where it isolates the inlet cham~er 47 from the
supply chamber 25, while establishing full flow communication
between the second exhaust chamber 64 with the supply chamber
25. In this way, during control of negative load, the supply
chamber 25 is connected through the second exhaust chamber 64,
line 63 and exhaust line 42 to the total exhaust manifold of
flow control valve 10 and 14, pressurized by the exhaust re-
lief valve 43. The pressure setting of the exhaust relief val-
ve 43 is high enough to provide the necessary pressure drop
through metering slot 84, at the highest rates of flow from
the supply chamber 25 to the load chamber 26, without pressure
in the load chamber 26 dropping below atmospheric level, thus
preventing any possibility of cavitation. In this way, during
control of negative load, inlet flow requirement of the actu-
ator is not supplied from the pump circuit but from the pres-
surized exhaust circuit of flow control valves 10 and 14, con-
serving the pump flow and increasing system efficiency. If
negative load pressure is not sufficiently high to provide
constant pressure drop through metering slot 32, the negative
load control spool 35 will move to the right from its modula-
ting and throttling position, the negative load pressure in
the outlet chamber 29 and space 56 will drop to a level at
which the pressure in space 51, due to the setting of the ex-
hau~t relief valve 43, with the biasing force of control spring
70 will move the piston 55 to the right together with the con-
trol spool 50 and the control system will revert to its posi-
tive load mode of operation, providing the energy to load L
from the pump circuit to maintain a constant pressure differen-
tial across metering slot 84, which will also maintain a con-
stant pres~ure diferential across metering slot 32. During
control of negative load the inlet flow requirement of the
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actuator is supplied from the outlet flow from the actuator,
bypass flow from pump flow control and the exhaust circuits
of all of the other system flow control valves. During control
of negative load system exhaust manifold is not only connected
with the supply chamber through the second exhaust chamber but
also through bypass check valve 73. The bypass check valve 73
is only needed in certain applications, in which a change from
positive to negative load can occur at high rates of flow, like
for example in hydrostatic transmission circuit, well known in
the art. For most of the applications controlling hydraulic
cylinders the bypass check valve 73 can be dispensed with.
As shown in Fig. 1 the piston 55 is of a larger di-
ameter than the control spool 50. If desired, in a well known
manner, the piston 55 can be slidably mounted in a bore provi-
ded in plug 89 and be of the same or even smaller diameter than
control spool 50. In Fig. 1 space 56 is connected by passage
57 to outlet chamber 29 and receives negative load pressure
signal from the outlet chamber 29. Space 56 can be connected
to negative load sensing ports located in the bore 19 between
load and outlet chambers and be selectiveIy communicable by
the valve spool 20 with the negative load pressure existing in
the load chambers.
Although the preferred embodiments of this invention
have been shown and described in detail it is recognized that
the invention is not limited to the precise form and structure
shown and various modifications and rearrangements as will oc-
cur to those skilled in the art upon full comprehension of this
invention may be resorted to without departing from the scope
of the invention as defined in the claims.
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