Note: Descriptions are shown in the official language in which they were submitted.
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CONTR~)L ACTUATION SYSTEM INC~UDING
STAGED DIR~CT DRI~E VALVE WITH FAULT CONTROL
This invention relates generally to a fluid servo system, and
more particularly to an aircraft flight control servo system including a
control actuation system incorporating an electro-mechanically controlled,
hydraulically powered actuator for use in driving a main control valve of the
servo system.
Fluid servo systems are used for many purposes, one being to
position the flight control surfaces of an aircraft. In such an application,
system redundancy is desired to achieve increased reliability in various
modes of operation, such as in a control augmentation or electrical mode.
In conventional electro-hydraulic systems, plural redundant
electro-hydrautic valves have been used in conjunction with plural redundant
servo valve actuators to assure proper position control of the system's main
control servo valve in the event of failure of one of the valves and/or servo
actuators, or one of the corresponding hydraulic systems. Typically, the
servo actuators operate on opposite ends of a linearly movable valve
element of the main control valve and are controlled by the electr~
hydraulic valves located elsewhere in the system housing. ~lthough the
servo valve actuators, alone or together, advantageously are capable of
driving the linear movable valve element against high reaction forces, such
added redundancy results in a complex system with many additionaI electri-
cal and hydraulic elements necessary to perform the various sensing)
eq~ fltion, timing and other control functions. This gives rise to reduced
overall reliability, increased package size and cost, and imposes added
requirements on the associated electronics.
An alternative approach to the electr~hydraulic control system
is an electro-mechanical control system wherein a force motor is cGupled
directly and mechanically to the main control servo valve. In this system,
redundancy has been accomplished by mechanical summation of forces
directly within the multiple coil force motor as opposed to the conventional
electro-hydraulic system where redundancy is achieved by hydraulic force
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summing using multiple electro-hydraulic valves and actuators. If one coil
or its associated electronics should fail, its counterpart channel will
maintain control while the failed channel is uncoupled and made passive.
Such alternative approach, however, has a practical limitation in that direct
drive force motors utiIizing state of the art rare earth magnet materials are
not capable of producing desired high output forces at the main control
servo valve within acceptable size and weight limitations.
In aircraft flight control systems, it also is advantageous and
desirable to provide for controlled recentering of the main control servo
valve in the event of a total failure or shut-down of the electrical
operational mode. This is particularly desirable in those control systems
wherein a manual input to the main servo valve is provided in the event that
a mechanical reversion is necessary after multiple failures have rendered
the electrical mode inoperative. In known servo systems of this type, the
manual input may operate upon the spool of the main servo valve whereas
the electrical input operates upon the movable sleeve of the main servo
valve.
Upon rendering the electrical mode inactive, it is necessary to
move the valve sleeve to a neutral or centered position and lock it against
movement relative to the valve spool controlled by the manual input.
Heretofore, this has been done by using a centering spring device whic~h
moves the valve sleeve to its centered or neutral position and a spring
biased plunger that engages a slot in the valve sleeve to lock the latter
against movement. The plunger normally is maintained out of engagement
with the slot during operation in the electrical mode by hydraulic system
pressure, and may have a tapered nose that engages a similarly tapered slot
in the valve sleeve to assist in centering the valve sleeve.
Some redundant control actuation systems are particularly suited
for use in applications where the required stroke of the main control valve
element is relatively small and about equal the desired stroke for the pilot
valve. In other applications, however, the required stroke of the main
control valve element is relatively long and may be several times longer
than can be the stroke of the pilot valve within acceptable size and weight
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limitations. This would be the case, for example, for flight controls
requiring main control valve flow rates of 15 to 25 gallons per minute and a
stroke of about plus or minus .050 inch or more, whereas the pilot valve
desirably would have a flow rate of less than one gallon per minute and a
stroke of say plus or minus .015 inch.
Also in long stroke applications, the force motor then would be
required to have high output energy capability. The energy required of a
force motor to drive the pilot valve is approximately proportional to force
reguired times stroke over which the force must act, the force level usually
being established by specified valve chip shearing requirements in aircraft
applications. In some systems, the relatively long stroke requirement placed
upon the pilot valve therein imposes an energy penalty on the force motor.
Accordingly, there would be required a higher energy force motor which is
disadvantageous because it is larger and heavier and requires higher
electrical power, associated larger electrical circuit elements and heat
rejection devices.
The control actuation system of the present invention has
particular advantages in applications requiring a relatively long stroke main
control valve. Briefly, the actuation system includes an electro-mechani-
cally controlled, hydraulically powered actuator for driving the main control
valve of a servo actuator control system. The actuator includes a tandern
piston connected to the main control valve which is controllably positioned
by a staged valve having a relatively short stroke whereby a force motor of
minimum size and energy requirements may be used to directly drive the
valve. Despite the relatively small size and output energy capability of the
force motor, the system is capable of driving the main control valve through
a relatively long stroke and` against high reaction forces as the valve is
hydraulically powered by one or both of the hydraulic systems.
In particular, the staged direct drive valve includes a linearly
movable, tubular valve plunger connected at one end to a flexible quill
which extends through and out of the tubular valve plunger for connection to
either a rotary or linear force motor. With a rotary force motor, the
flexible quill has a ball bearing in which is engaged an eccentric pin on the
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force motor drive shaft, and flexing of the quill accommodates the rise and
fall of the bearing during short arcuate movement of the eccentric pin
without applying significant side loads to the valve plunger. ~Alternatively, a
linear force motor may have its linear drive member connected to the
flexible quill whereby flexing of the quill accommodates any misalignment
of the drive member and valve plunger without applying significant side
loads to the valve plunger.
Also, the staged valve includes a fault control valve sleeve
concentric with the valve plunger which, upon shut-down or failure of both
hydraulic systems, moves linearly to render the valve plunger inoperative
and release fluid pressure from opposed, corresponding pressure surfaces of
the tandem piston to respective returns therefor through respective center-
ing rate control orifices in the fault control valve sleeve as the piston is
moved to a neutral position by a centering spring device acting on the main
control valve. For normal operation, the fault control valve sleeve is
movable by fluid pressure from either hydraulic system to a position
permitting controlled differential application of fluid pressure to the
tandem piston sections by the valve plunger. In addition, system pressure is
applied to the actuator through shut-down valves which, upon shut-down of
the system, disconnect the actuator from system pressure sources and
release fluid pressure from other opposed, corresponding pressure surfaces
of the tandem piston sections to return through flow restricting orifices,
whereby the piston is hydrnulically locked ngainst high loads of short
duration.
Preferably, such a control actuation system is used for driving
the main control valve of a dual hydraulic servo actuator control system
which obtains the advantages of both electro-hydraulic and electro-
mechanical control systems while eliminating drawbacks associated there-
with.
Such a control actuation system is capable of being electro-
mechanically controlled by a linear or rotary force motor drive within
acceptable size and weight limitations, and is purticularly suited for use in
applications requiring driving of the main control valve through a relatively
long stroke in relation to the stroke of the force motor drive.
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Such a control actuation system has high reliability, reduced
complexity, and reduced package size and cost in relation to known
comparable systems.
Such a control actuation system is capable of driving the main
control valve against relatively high reaction forces, and preferably effects
re-centering of the main control servo valve at a controlled rate under
system shut~own or failure conditions.
Preferably such a control actuation system may be provided with
a fault control having centering rate control provisions that is responsive to
one or both hydraulic systems and effective regardless of control actuator
stroke position.
Preferably such a control actuation system has high stiffness and
is capable of supporting higll loads.
In accordance with one aspect of the present invention, there is
provided a control actuation system useful in a dual hydraulic servo actuator
control system for operating a relatively long stroke, control valve element
therein, comprising an actuator, a tandem piston axially movable in said
actuator and dl;vingly connectable to the control valve element, staged
valve means operatively connected to said actuator for effecting axial
movement of said piston in opposite directions, and control input means
including force motor means for linearly driving said staged valve means
through a relatively short stroke in relntion to the stroke of said piston to
effect position control of said piston, said piston including two serially
connected piston sections each having axially opposed pressure surfaces, and
said staged valve means including a valve plunger having two serially
connected valving sections respectively for controlling the differential
application of fluid pressure from respective sources thereof on said opposed
pressure surfaces of respective s'aid piston sections to cause axial movement
of said piston in opposite directions in response to relatively short, directly
driven linear rnovement of said valve plunger in opposite directions, a fault
control valve member axially movable in said staged valve means, and
means responsive to fluid pressure from either source thereof being supplied
to said staged valve means for moving said fault control valve member from
12031'~5
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a disabling position blocking the differential application of fluid pressure to
said piston by said valve plunger to an enabling position permitting such
application and release of fluid pressure.
In accordance with a further aspect of the present invention,
there is provided a control actuation system useful in a dual hydraulic servo
actuator control system for operating a relatively long stroke, control valve
element therein, comprising an actuator, a tandem piston axially movable in
said actuator and d~ivil~gly colmectable to the control valve element, staged
valve means operatively connected to said actuator for effecting axial
movement of said piston in opposite directions, and control input means
including force motor means for linearly driving said staged valve means
through a relatively short stroke in relation to the stroke of said piston to
effect position control of said piston, said piston including two serially
connected piston sections each having axially opposed pressure surfaces, and
said staged valve means including a valve plunger having two serially
connected valving sections respectively for controlling the differential
application of fluid pressure from respective sources thereof on said opposed
pressure surfaces of respective said piston sections to cause axial movement
of said piston in opposite directions in response to relatively short, directly
driven linear movement of said valve plunger in opposite directions, said
opposed pressure surfaces of each piston section being opposed to corres-
ponding pressure surfaces of the other piston section, respective means for
simultaneously supplying fluid pressure from such respective sources thereof
to said actuator and staged vaIve means and for disconnecting such supply to
effect system shut-down, centering means for urging said piston to a neutral
position upon system shut-down, flnd menns responsive to system shut-down
for releasing fluid pressure acting on opposed corresponding pressure sur-
faces of said piston sections through respective metering orifices to control
the rate at which said piston is moved to its neutral position by said
centering means, said means responsive to system shut-down including a
fault control valve member axially movable in said staged valve means.
In accordance with a further aspect of the present invention,
there is provided a control actuation system useful in a dual hydraulic servo
::.
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actuator control system for operating a relatively long stroke, control valve
element therein, comprising an actuator, a tandem piston axially movable in
said actuator and drivingly connectable to the control valve element, staged
valve means operatively connected to said actuator for effecting axial
movement of said piston in opposite directions, and control input means
including force motor means for linearly driving said staged valve means
through a relatively short stroke in relation to the stroke of said piston to
effect position control of said piston, said piston including two serially
connected piston sections each having axially opposed pressure surfaces, and
said staged valve means including a valve plunger having two serially
connected valving sections respectively for controlling the differential
application of fluid pressure from respective sources thereof on said opposed
pressure surfaces of respective said piston sections to cause a~ial movement
of said piston in opposite directions in response to relatively short, directly
driven linear movement of said valve plunger in opposite directions, said
opposed pLe3~ e surfaces of each piston section having unequal effective
pressure areas, and means for applying fluid pressure from such respective
sources thereof normally only on the smaller area pre~ ~ul e surface of
respective said piston sections, said valving sections of said valve plunger
being operable upon movement of said plunger either to apply fluid pressure
from such respective sources thereof on the larger area pressure surfaces of
respective said piston sections or to release fluid pressure acting on said
larger area pressure surfaces of respective said piston sections to respective
returns therefor for fluid actuation of said piston in opposite directions, a
fault control valve member a2~iaUy movable in said staged valve means, and
means responsive to the application of fluid pressure from either source
thereof upon said smaller area pressure surfaces of said piston sections for
moving said fault control valve member from a disabling position blocking
such application and release of fluid pressure acting on said larger area
pressure surfaces to an enabling position permitting such application and
release of fluid pressure.
In accordance with a further aspect of the present invention,
there is provided a control actuation system useful in a hydraulic servo
actuator control system for operating a control valve element therein,
~2031'~S
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comprising an actuator, a piston axially movable in said actuator and
drivingly connectable to the controI valve element, staged valve means
operably connectable to control input means for effecting position control
of said piston, said staged valve means in~lu(1ing a linearly movable valve
plunger for directing fluid pressure against said piston to cause axial
movement of said piston, centering means for urging said piston to a neutral
position upon such control input means being rendered inoperative, and
means responsive to such control input means being rendered inoperative for
releasing fluid pressure acting on opposite sides of said piston through
metering orifices to control the rate at which said piston is urged to the
neutral position thereof by said centering means, said means for releasing
including a fault control valve member linearly movable in said staged valve
means to a position providing for the release of fluid pressure from one side
of said piston through a respective one of said metering orifices.
An embodiment of the invention will now be described, by way of
an example, with reference to the accompanying drawings, in which:
Figure 1 is a schematic illustration of a redundant servo system
embodying a preferred form of a control actuation system including staged
direct drive valve with fault control, according to the invention;
Figure 2 is an enlarged section through such staged direct drive
valve with fault control shown in its shut-down condition; and
Figure 3 is an enlarged section similar to Figure 2 but showing
the staged valve in its operational condition.
Referring now in detail to the drawings and initially to Figure lp
a dual hydraulic servo system is designated generally by reference numeral
10 and includes two similar hydraulic servo actuators 12 and 14 which are
connected to a common output device such as a dual tandem cylinder
actuator 16. The actuator 16 in turn is connected to a control member such
as a flight control element 18 of an aircraft. It will be seen below that the
two servo actuators normally are operated simultaneously to effect position
control of the actuator 16 and hence the flight control element 18. However,
each servo actuator preferably is capable of properly effecting such position
control independently of the other so that control is maintained even when
one of the servo actuators fails or is shut down. Accordingly, the two servo
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actuators in the overall system provide a redundancy feature that increases
safe operation of the aircraft.
The servo actuators seen in Figure 1 are similar and for ease in
description~ like reference numerals will be used to identify corresponding
like elements of the two servo actuators.
The servo actuators 12 and 14 each have an inlet port 20 for
connection with a source of high pressure hydraulic fluid and a return port
22 for connection with a hydraulic reservoir. Preferably, the respective
inlet and return ports of the servo actuators are connected to separate and
independent hydraulic systems in the aircraft, so that in the event one of
the hydraulic systems fails or is shut down, the servo actuator coupled to
the other still functioning hydraulic system may be operated to effect the
position control function. Hereinafter, the hydraulic systems associated
with the servo actuators 12 and 14 will respectively be referred to as the aft
and forward hydraulic systems.
In each of the servo actuators 12 and 14, a passage 24 connects
the inlet port 20 to a servo valve 26. ~nother passage 28 connects the
return port 22 to the same servo valve 26. Each passage 24 may be provided
with a check valve 30.
The main control servo valve 26 includes a spool 32 which is
longitudinaJly shiftable in a sleeve 34 which in turn is longitudinally
shiftable in the system housing 36. The spool and sleeve are divided into
two fluidically isolated valving sections indicated generally at 38 and 40 in
Figure 1, which valving sections are associated respectively with the
actuators 12 and 14 and the passages 24 and 28 thereof. Each valving section
of the spool and sleeve is provided with suitable lands, grooves and passages
such that either one of the spool or sleeve may be maintained at a neutral or
centered position, and the other selectively shifted for selectively con-
necting the passages 24 and 28 of each servo actuator to passages 42 and 44
in the same servo actuator.
The passages 42 and 44 of both servo actuators 12 and 14 are
connected to the dual cylinder tandem actuator 16 which includes a pair of
cylinders 46. The passages 42 and 44 of each servo actuator are connected
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lZ03:~15
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to a corresponding one of the cylinders at opposite sides of the piston 48
therein. If desired, anti-cavitation valves 50 and 52 respectively may be
provided in the passsges 42 and 44. The pistons 48 and the cylinders 46 are
interconnected by a connecting rod 54 and further are connected by output
rod 56 to the control element 18 through linkage 58.
From the fore~oillg, it will be apparent that selective relative
movement of the spool 32 and sleeve 34 simultaneously controls both valving
sections 38 and 40 which selectively connect one side of each cylinder 46 to
a high pressure hydraulic fluid source and the other side to fluid return for
effecting controlled movement of the output rod 56 either to the right or
left as seen in Figure 1. In the event one of the servo actuators 12, 14 fails
or is shut down, the other servo actuator will msintain control responsive to
selective relfltive movement of the spool and sleeve.
The relatively shiftable spool 32 and sleeve 34 provide for two
separate operational modes for effecting the position control function. The
spool, for example, may be operatively associated with a manual operational
mode while the sleeve is operatively associated with a control augmented or
electrical operational mode. In the manual operational mode, spool posi-
tioning may be effected through direct mechanical linkage to a control
element in the aircraft cockpit. As seen in Figure 1, the spool may have a
cylindrical socket 58 which receives a ball 60 at the end of a crank 62. The
crank 62 may be connected by a suitable mechanical linkage system to the
afrcraft cockpit control element. For a more detailed description of such a
mechanical linkage system, reference may be had to U.S. Patent No.
3,956,971 entitled "Stabilized Hydromechanical Servo System", issued May
18,1976.
Normally, the manual control mode will remain passive unless a
failure renders the electrical mode inoperable. During operation in the
electrical mode, the spool 32 is held in a neutral or centered position while
the sleeve 34 is controllably shifted to effect the position control function
by the hereinafter described control actuation system designated generally
by reference numeral 70.
i~33~5
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The control actuation system 70 of the invention incIudes an
electro-mechanically controlled, hydraulically powered actuator 72 which is
shown positioned generally in axial alignment with the main control servo
valve 26 as seen at the lower left in Figure 1. The actuator 72 includes a
tandem piston 74 which is positioned for axial movement in a stepped
cylinder bore 76 in the housing 36. At its end nearest the servo valve 26,
the piston 74 has a piston extension 80 which extends axially in a cylindrical
bore 82 of the housing 36, which bore may be an axial continuation of the
cylindrical housing bore 84 accommodating the sleeve 34 and spool 32~ The
sleeve extension 80 is connected by suitable means to the sleeve 34. The
sleeve extension 80 also may have a diametrRl slot 86 therein which may be
engaged by a spring biased plunger 88 to lock the interconnected piston 74
and sleeve 3~ against axial movement. As will be further discussed
hereinafter, the plunger 88 normally is maintained out of engagement with
the slot 86 during operation in the electrical mode by hydraulic system
pressure.
The tandem piston 74 includes two serially connected or
arranged piston sections 90 and 9?.. The piston section 90 has a cylinder
pressure surface 94 and a source pressure surface 96 in opposition to the
cylinder pressure surface 94. Similarly, the piston section 92 has a cylinder
pressure surIace 98 and an opposed source pressure surface 100. Also, the
co.~e3~0nding cylinder and source pressure surfaces of the piston sectio~s
are opposed and have equal effective pressure areas, respectively. This
results in ~alanced forces acting on piston sections having matched charac-
teristics.
The source pressure surfaces 96 and 100 of the piston sections 9D
and 92 respectively are in fluid communication with passages 102 and 104
which, as seen in Figure 1, lead to shut-down valves 106 and 108, respec-
tively. The shut-down valves 106 and 108 may be conventionsl three-way,
solenoid-operated valves which when energized respectively establish com-
munication between the passages 102 and 104 and supply passages 110 and 112
that connect the shut-down valves 106 Rnd 108 to the forward and aft
~lZ~)31~
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hydraulic system supplies associated with the actuators 14 and 12, respec-
tively. When de-energized, the shut-down valves 106 and 108 respectively
connect the passages 102 and 104 to return passages 114 and 116 which are
connected to the forward and aft hydraulic system returns associated with
the actuators 14 and 12, respectively. For a purpose that will become more
apparent below, the passages 114 and I16 have therein centering rate control
or metering orifices 118 and 120, respectively.
Referring additionally to Figures 2 and 3, the passages 102 and
104 also respectively are connected by passages 122 and 124 to a remotely
located pilot or staged valve 125. More particularly, the passages 122 and
124 are respectively connected to ports 126 and 128 in a fault control valve
sleeve 130, and the ports 126 and 128 in turn respectively are in fluid
communication with annular groove 132 and port 134 in a static porting
sleeve 136. The fault control valve sleeve 130 and static porting sleeve 136
are concentrically arranged in a bore 138 of the system housing 36 with the
fault control valve sleeve being axially shiftable relative to the housing 36
and porting sleeve 136, and the porting sleeve being fixed to the housing 36
against axial movement.
As seen at the right in Figure 2, the porting sleeve 136 has a
cylindrical extension 140 which has an end piece 142 axially butted against a
stop plug 144 fixed in the housing 36 to prevent axial movement o the
porting sleeve 136 to the right as seen in Figure 2. The cyIindrical extension
140 also is diametrically slotted for receipt of a diametrically extending pin
146 fixed in the housing 36 which has a central ball portion 148. The b~ll
portion 148 serves as an axial stop against which bears a plug insert 150 that
is fixed in the cylindrical extension 140 by means of a pin 152. Accordingl1y,
the indicated engagement of the plug insert lS0 against the ball portion 14~
of the pin 146 prevents axial movement of the porting sleeve 136 to the left
as seen in Figure 2.
The fault control valve sleeve 130 has a cylindrical outer surface
of constant diameter, whereas the radially inner surface thereof, and thus
the opposed radially outer surface of the porting sleeve 136, is radially
stepped along its axial length to provide different thickness valve sleeve
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portions. As a result, the fault control valve sleeve has a slightly reduced
thickness central portion 154 extending between the ports 126 and 128 and a
still further reduced thickness portion 156 extending to the left of the port
126 thus providing two differential pressure surfaces 158 and 160 at the right
side of each of the ports 126 and 128 as seen in Figure 2 and exposed to the
fluid pressure supplied to such ports. Thus, connection of either or both
ports 126 and 128 to respective sources of high pressure fluid will shift the
fault control v~lve to the right relative to the porting sleeve 136 and to its
control ~n~hling position seen in Figure 3.
Such shifting of the fault control valve sleeve 130 is opposed by
the force exerted by a spring 162 which is positioned at the right end of the
bore 138 and bears in opposition against the end wall 164 of the bore 138 and
a shoulder 166 on the valve sleeve 130. ~ccordingly, the spring 162 urges the
fault control valve sleeve 130 to the left as seen in Figure 2 and towards a
radially outwardly extending flange 168 on the porting sleeve 136 which acts
as a stop to define the control ~ hling position of the fault control valve
sleeve when butted thereagainst. On the other hand, the end wall 164 acts
as an opposed stop to define the ~n~hling position of the fault control valve
sleeve when a cylindrical extension 170 on the fault control valve sleeve is
butted thereagainst a~ seen in Figure 3.
When the fault control valve sleeve 130 is in its ~nP~hling position
of Figure 3, ports 172 and 174 in the fault control valve sleeve respectively
effect communication between ports 176 and 178 in the porting sleeve 136
and the passages 180 and 182 which in turn respectively communicate with
the cylinder pressure surfaces 94 and 98 as seen in Figure 1. In addition, the
ports 176 and 178 are associated with respective axially arranged valving
sections of a valve plunger 184.
The valve plunger 184 of the staged valve 125 is concentric with
and constrained for axial movement in the porting sleeve 136. The valving
section of the valve plunger associated with the port 176 consists of annular
grooves 186 and 188 which are axially separated by a metering land 190. The
metering land 190 is operative to block communication between the associ-
ated port 176 and the grooves 186 and 188 when the plunger 184 is in a null
lZ031'~5
position. HoweverJ upon axial movement of the plunger relative to the
porting sleeve 136 and out of its null position, the metering land is operative
to effect cornmunication between the port 176 and one or the other of the
grooves 186 and 188 depending on the direction of movement.
The groove 186 is in fluid communication with a port 192 in the
porting sleeve 136 which in turn communicates with the port 126 when the
fault control valve sleeve 130 is in its ~?n~hling position of Figure 3. Accord-ingly, fluid ~leO~u.e will be supplied to the groove 186 when the passage 122
is connected to the supply passage llû by the shut-down valve 106. It is
noted that at the same time, fluid pre~ule wiIl be applied on the source
pl~O~I,Le surface 96 of the piston section 90. The other groove 188 is in
communication with a port 194 in the porting sleeve 136 which in turn
communicates ViQ a port 196 in the fault control valve sleeve 130 with a
passage 198 connected to the return passage 114 downstream of the orifice
118 ~s seen in Figure 1. ~ccordinglyJ the groove 188 is connected to the
return of the respective or forward hydraulic system.
Similarly, the valving section of the pilot valve pllmger 184
associated with the port 178 has a pair of annular grooves 200 and 202 which
are axially separated by a metering land 204 which is operative in the same
manner as the metering land 190 but in association with the port 178. The
groove 200 is in fluid communication with the port 134 whereas the other
groove 202 is in fluid communication with the return passage 116 of the
respective or aft hydraulic system via a port 206 in the porting sleeve, port
208 in the fault control valve sleeve and a passage 210 connected to the
return passage 116 downstream of the orifice 120 as seen in Figure 1.
The staged valve 125 also has a passage 212 which connects the
passage 210 to the right or outer end of the bore 138 as seen in Figure 3.
Accordingly, the right end face of the plunger 184 will be exposed to return
pressure of the aft hydraulic system. Also provided is a passage 214 which
connects the right end of the bore 138 to the left end thereof so that the left
end face of the plunger 184 is exposed to the same fluid pressure as its right
end face. Moreover, the left and right end faces of the plunger have equal
effective pressu~e areas whereby return pressure variations will not apply
unbalanced forces and consequent inputs to the plunger.
1;~031'~
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It should now be apparent that selective axial movement of the
plunger 184 relative to the porting sleeve 136 simultaneously controls both
valving sections thereof which in turn control the differential application of
fluid pressure from respective independent hydraulic systems on the opposed
pressure surfaces of the piston sections 90 and 92. If the plunger is moved
to the right from its null position, fluid pressure is applied to the cylinder
pressure surface 94 of piston section 90 from the forward hydraulic system
source associated therewith while fluid pressure is released from cylinder
pressure surface 98 of piston section 92 to the aft hydraulic system return
associated therewith. The resultant pressure imbalance will hydraulically
power the piston 74, and thus the main control servo valve sleeve 34, to the
right as seen in Figure 1. Conversely, if the plunger is moved to the left
from its null position, fluid pressure is applied to the cylinder pressure
surface 98 of the piston section 92 from the aft hydraulic system source
associated therewith while fluid pressure is released from the cylinder
pre~au.e surf~ce 94 of the piston section 90 to the forward hydraulic system
return associated therewith. Under these conditions, the resultant pressure
imbalance will hydraulically power the piston 74 and valve sleeve 34 to the
left as seen in Figure 1. Accordingly, movement of the plunger in either
direction will control the differential application of fluid pre3~u.e on the
piston sections 90 and 92 to effect movement of the piston in opposite
directions. In addition, either piston section and associated valving section
of the plunger will maintain control of the piston in the event that the
hydraulic system associated with the other is shut down or otherwise lost.
With particular reference to Figure 3, controlled selective move-
ment of the valve plunger 184 may be effected by a force motor ~18 located
closely adjacent one end of the plunger. The force motor may be responsive
to command signals received from the aircraft cockpit whereby the force
motor serves as a control input to the plunger. Also, the force motor
preferably has redundant multiple parallel coils so that if one coil or its
associated electronics should fail, its counterpart channel will maintain
control. Moreover, suitable failure monitoring circuitry is preferably
provided to detect when and which channel has failed, and to uncouple or
render passive the failed channel.
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The force motor 218 includes a motor housing 220 which is
secured in the system housing 36 closely adjacent one end of the valve
plunger 184 with its drive shaft 222 extending perpendicularly to a plane
through the longitudinal axis of the valve plunger. The drive shaft is shown
d~iYill~ly connected to the valve plunger by a flexible link member or quill
224 which is connected at opposite ends to the valve plunger and drive shaft.
The valve plunger being tubular as shown has an axial bore 226 through
which the quill extends for connection at its threaded end 228 to the closed
end of the valve plunger furthest or opposite the force motor. At its other
end, the quill extends out of the bore for connection to the drive shaft 222,
such other end being provided with a ball bearing 232 engaged by an
eccentric pin 234 on the drive shaft. More particularly, the eccentric pin is
closely fitted in the inner race of the bearing which has its outer race
closely fitted in a transverse bore 236 ht the quill.
The quill 224 may have a cylindrical portion 238 and a reduced
diameter flexible length portion 240. The cylindrical portion extends from
the threaded end 228 of the quill about half way through the plunger 184 and
is closely fitted in the axial bore 226 whereby flexing of the quill is`limited
to the reduced diameter portion 240. It will be appreciated that the flexing
portion 240 of the quill accommodates the rise and fall of the bearing 232
without applying significant side loads to the tubular valve plunger as the
eccentric pin 234 is driven by the force motor 218 through a short arcuate
stroke. It also is noted that the effective length of the quill may be
adjusted at its threaded end 228 for adjusting the neutral or nllll position of
the plunger relative to a mlll position of the force motor.
The valve plunger 184 alternatively may be driven by a linear
force motor. For this, the quill 224 is provided at its force motor
connection end with a threaded axial bore 244 for connection to the linear
drive element of the linear force motor. With a linear force motor, the
flexible quill will accommodate any mi~ nment between such drive
member and the valve plunger without applying significant side loads to the
valve plunger.
~.~203145
-14-
As indicated, controlled selective movement of the valve plunger
184 is effected by the force motor 218 which is responsive to command
signals received, for example, from the aircraft cockpit. To provide proper
feedback information to the command system controlling the force motor, a
position transducer 246 may be operatively connected to the actuator piston
74 as schematically shown in Figure 1. With this arrangement, it will be
appreciated that the stroke of the plunger and force motor may be
relatively short in relation to that of the actuator piston 74 which may be
required to h~ve a relatively long stroke for driving the main control valve
sleeve 34. This permits reduction of plunger length, drive motor size and
energy capability, and the amount of space otherwise required to accomplish
the fault control function described hereinafter.
The fault control function is effected upon shifting of the fault
control valve sleeve 130 to its ~ hling position seen in Figure 2. Such
shifting will occur whenever the fluid pressure acting upon the difIerential
p~ u.e surfaces 158 and 160 of the valve sleeve 130 at the ports 132 and 134
therein is insufficient to overcorne the force exerted by the spring 162. This
will occur upon failure of both independent hydraulic systems or upon shut-
down of the electrical operational mode by the shut-down valves 106 and 108
after multiple failures have rendered such mode inoperative. Upon such
failure or shut-down, the spring 162 will shift the fault control valve sleeve
130 to its rli.~ahling position whereat the inner end of the valve sleeve will be
butted a~ainst the flange 168 on the porting member 136.
When the fault control valve sleeve 130 is in its ~icahling
position, communication between the cylinder pre~,~ule surfaces 94 and 98
and the respective supply passages 122 and 124 is blocked by the fault control
valve sleeve regardless of the position of the valve plunger 184. More
particularly, the fault control valve sleeve in its ~ hlin~ position blocks
communication between the passage 122 and the port 192 which otherwise
cooperate to supply fluid pressure to the groove 186 associated with the
cylinder pressure surface 94. Also, the fault control valve sleeve blocks
communication between the passage 182 and port 178 associated with the
cylinder pressure surface 98.
~03~5
-15-
ln addition, the fault control valve sleeve 130 in its disabling
position blocks communication between the cylinder pressure surfaces 94
and 98 and the respective return passages 198 and 210 regardless of the
position of the valve plunger 184, except through respective centering rate
control or metering orifices 250 and 252 provided in the fault control valve
sleeve. More particularly, shifting of the valve sleeve 130 to its ~ hling
position blocks communication between the port 176 and port 172 while
estahli~hing communication between the passage 180 and passage 198 via the
metering orifice 250. At the same time, communication between the
passage 182 and port 178 is blocked as indicated above while communication
between the passage 182 and p~ssage 210 is established via the metering
orifice 252~
As a result, fluid pre ~ule from the cylinder pressure surfaces 94
and 98 will be released to the return passages 198 and 210 through the
metering orifices 250 and 252 in the fault control valve sleeve 130 ~Ivhich
control the rate at which fluid is ported from the cylinder pressure surfaces
as the main control servo valve sleeve 34 and thus the piston 74 is moved to
a centered or neutral position by a spring centering device 254 for system
operation in the manual mode. The spring centering device 254 can be seen
at the right in Figure 1 and may be conventional.
Operation
During normal operation of the control actuation system in the
electrical mode, each shut-down valve 106, 108 is energized. This supplies
fluid pressure from the forward and aft hydraulic systems to the source
pressure surfaces 96 nnd 100 of the piston sections 90 and 92, respectively.
In addition, fluid pressure is supplied from the aft hydraulic system to the
end of a spring biased plunger 256 seen in Figure 1~ This moves the plunger
256 to the right against the biasing force to open the passage 258 to fluid
pressure for moving the plunger 88 out of locking engagement with the
sleeve extension 80 thereby to permit axial movement of the main control
valve sleeve 34 and the piston 74.
Fluid pressure also will be supplied to the ports 182 and 134 via
passages 122 and 124, respectively, whereupon the fault control valve sleeve
~2~3~5
--16-
130 will be shifted from its disabling position of Figure 2 to its en~3hling
position of ~igure 3. With the fault control valve sleeve in its enabling
position, controIled positioning of the piston 74 and hence the main control
valve sleeve 34 may be effected by the valve plunger 184 and force motor
218 in response to electrical command signals received from the aircraft
cockpit. It will be appreciated that simultaneous energization of the shut-
down valves will not cause large turn-on transients because the pressure
surfaces of the piston sections result in equal and opposite forces on the
piston by reason of their aforedescribed pressure area and porting relation-
ships.
Position control of the piston 74 and main control valve sleeve
34 will be maintained even though one of the channels of the electrical
mode fails or is rendered inoperative. ~lowever, if both chflnnels fail or are
rendered inoperative requiring reversion to the manual operational mode,
both shut-down valves 106 and 108 are de-energized. This connects the
source pressure surfaces 96 and 100 of the piston sections 90 and 92 to
return pressure and effects shifting of the fault control valve sleeve 130 to
its (1i.~hling position shown in Figure 2. As the main control valve sleeve 34
is urged towards its centered or neutral position by the centering spring
device 254, fluid will be pumped out of the actuator mechanism at a rate
controlled by the then existing pressures due to the centering spring force
and the centering rate control orifices 118,120, 250 and 252. Depending Ol~
the direction of centering movement, either the centering rate control
orifices 118 and 252 or the orifices 120 and 250 wiU act in concert to control
the rate of centering. As control orifices are provided for each piston
section, centering rate control is ensured even if fluid is totally lost from
one of the hydraulic systems. Moreover, centering rate control is effective
regardless of the position of the piston 74 or valve plunger 184.
When in the manual operational mode, the main control servo
valve sleeve 34 is held in its centered or neutral position by the centering
spring device 254 and also by the locking plunger 88 which will then have
moved into loclcing engagement with the slot 86 in the sleeve extension 80.
In the unlikely event that a relatively large reaction force is applied on the
1203~L~S
-17-
valve sleeve 34 which exceeds the holding capability of the centering spring
device and locking plunger, fluid pressure behind the opposing pressure
surfaces of the piston sections 90 and 92 would be built up. As a result, a
relatively large re~ iv~ force would be caused to act upon the piston
depending on the duration of the applied reaction force thereby to resist
back-driving of the piston. Of course, an extended relatively large reaction
force application time would, upon unseating of the locking plunger,
eventually move the piston from center upon the pumping of fluid through
the centering rate control orifices.
Although the invention has been shown and described with
respect to a cerWn preferred embodiment, it is obvious that equivalent
alterations and modifications will occur to others skilled in the art upon the
reading and understanding of the specification. The present invention
includes all such equivalent alterations and modifications, and is limited
only by the scope of the following claims.
