EP0115925A1 - Control actuation system including staged direct drive valve with fault control - Google Patents
Control actuation system including staged direct drive valve with fault control Download PDFInfo
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- EP0115925A1 EP0115925A1 EP84300329A EP84300329A EP0115925A1 EP 0115925 A1 EP0115925 A1 EP 0115925A1 EP 84300329 A EP84300329 A EP 84300329A EP 84300329 A EP84300329 A EP 84300329A EP 0115925 A1 EP0115925 A1 EP 0115925A1
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- European Patent Office
- Prior art keywords
- piston
- pressure
- valve
- control valve
- control
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
- F15B18/00—Parallel arrangements of independent servomotor systems
Definitions
- 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.
- system redundancy is desired to achieve increased reliability in various modes of operation, such as in a control augmentation or electrical mode.
- plural redundant electro-hydraulic 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.
- the servo actuators operate on opposite ends of a linearly movable valve element of the main control valve and are controlled by the electro-hydraulic valves located elsewhere in the system housing.
- 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 additional electrical and hydraulic elements necessary to perform the various sensing, equalization, 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 electro-hydraulic control system is an electro-mechanical control system wherein a force motor is coupled directly and mechanically to the main control servo valve.
- 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 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 has a practical limitation in that direct drive force motors utilizing 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.
- valve sleeve 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.
- this has been done by using a centering spring device which 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 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.
- 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 required times stroke over which the force must act, the force level usually being established by specified valve chip shearing requirements in aircraft applications.
- 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 actuation system includes an electro-mechanically controlled, hydraulically powered actuator for driving the main control valve of a servo actuator control system.
- the actuator includes a tandem 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.
- 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.
- 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.
- the flexible quill has a ball bearing in which is engaged an eccentric pin on the 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.
- 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.
- 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 centering 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.
- 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.
- 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 hydraulically locked against high loads of short duration.
- 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 therewith.
- 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 particularly 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.
- 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-down or failure conditions.
- 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.
- control actuation system has high stiffness and is capable of supporting high loads.
- 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.
- the two servo actuators normally are operated simultaneously to effect position control of the actuator 16 and hence the flight control element 18.
- 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 actuators in the overall system provide a redundancy feature that increases safe operation of the aircraft.
- 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.
- 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.
- the hydraulic systems associated with the servo actuators 12 and 14 will respectively be referred to as the aft and forward hydraulic systems.
- a passage 24 connects the inlet port 20 to a servo valve 26.
- Another 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 longitudinally 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 I, 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 connecting 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 to a corresponding one of the cylinders at opposite sides of the piston 48 therein.
- anti-cavitation valves 50 and 52 respectively may be provided in the passages 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.
- 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.
- spool positioning may be effected through direct mechanical linkage to a control element in the aircraft cockpit.
- 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 aircraft cockpit control element.
- the manual control mode will remain passive unless a failure renders the electrical mode inoperable.
- 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 generall;' by reference numeral 70.
- the control actuation system 70 of the invention includes 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.
- 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 diametral slot 86 therein which may be engaged by a spring biased plunger 88 to lock the interconnected piston 74 and sleeve 34 against axial movement.
- 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 92.
- the piston section 90 has a cylinder pressure surface 94 and a source pressure surface 96 in opposition to the cylinder pressure surface 94.
- the piston section 92 has a cylinder pressure surface 98 and an opposed source pressure surface 100.
- the corresponding cylinder and source pressure surfaces of the piston sections are opposed and have equal effective pressure areas, respectively. This results in balanced forces acting on piston sections having matched characteristics.
- the source pressure surfaces 96 and 100 of the piston sections 90 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, respectively.
- the shut-down valves 106 and 108 may be conventional three-way, solenoid-operated valves which when energized respectively establish communication between the passages 102 and 104 and supply passages 110 and ll2 that connect the shut-down valves 106 and 108 to the forward and aft hydraulic system supplies associated with the actuators 14 and 12, respectively.
- the shut-down valves 106 and 108 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.
- the passages ll4 and ll6 have therein centering rate control or metering orifices 118 and 120, respectively.
- 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.
- 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 of the porting sleeve 136 to the right as seen in Figure 2.
- the cylindrical 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 ball 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. Accordingly, the indicated engagement of the plug insert 150 against the ball portion 148 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 portions.
- 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.
- connection of either or both ports 126 and 128 to respective sources of high pressure fluid will shift the fault control valve to the right relative to the porting sleeve 136 and to its control enabling 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. Accordingly, 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 disabling position of the fault control valve sleeve when butted thereagainst.
- the end wall 164 acts as an opposed stop to define the enabling position of the fault control valve sleeve when a cylindrical extension 170 on the fault control valve sleeve is butted thereagainst as seen in 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.
- 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 associated port 176 and the grooves 186 and 188 when the plunger 184 is in a null position.
- the metering land is operative to effect communication 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 enabling position of Figure 3. Accordingly, fluid pressure will be supplied to the groove 186 when the passage 122 is connected to the supply passage 110 by the shut-down valve 106. It is noted that at the same time, fluid pressure will be applied on the source pressure 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 via 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 as seen in Figure 1. Accordingly, the groove 188 is connected to the return of the respective or forward hydraulic system.
- the valving section of the pilot valve plunger 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 ll6 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 pressure areas whereby return pressure variations will not apply unbalanced forces and consequent inputs to the plunger.
- controlled selective movement of the valve plunger 184 may be effected by a force motor 218 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.
- 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.
- suitable failure monitoring circuitry is preferably provided to detect when and which channel has failed, and to uncouple or render passive the failed channel.
- 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 drivingly 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.
- 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 in 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.
- 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.
- the effective length of the quill may be adjusted at its threaded end 228 for adjusting the neutral or null position of the plunger relative to a null position of the force motor.
- the valve plunger 184 alternatively may be driven by a linear force motor.
- 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.
- the flexible quill With a linear force motor, the flexible quill will accommodate any misalignment between such drive member and the valve plunger without applying significant side loads to the valve plunger.
- valve plunger 184 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.
- a position transducer 246 may be operatively connected to the actuator piston 74 as schematically shown in Figure I.
- 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 have 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 disabling position seen in Figure 2. Such shifting will occur whenever the fluid pressure acting upon the differential pressure surfaces 158 and 160 of the valve sleeve 130 at the ports 132 and 134 therein is insufficient to overcome 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 disabling position whereat the inner end of the valve sleeve will be butted against the flange 168 on the porting member 136.
- the fault control valve sleeve 130 When the fault control valve sleeve 130 is in its disabling position, communication between the cylinder pressure 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 disabling 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.
- 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 disabling position blocks communication between the port 176 and port 172 while establishing 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 passage 210 is established via the metering orifice 252.
- 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.
- 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 and 100 of the piston sections 90 and 92, respectively.
- 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 132 and 134 via passages 122 and 124, respectively, whereupon the fault control valve sleeve 130 will be shifted from its disabling position of Figure 2 to its enabling position of Figure 3.
- controlled 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 relationships.
- the main control servo valve sleeve 34 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 locking engagement with the slot 86 in the sleeve extension 80.
- a relatively large reaction force is applied on the 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.
- a relatively large resistive 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.
- 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.
Abstract
A control actuation system (70) for an aircraft including an electro-mechanically controlled, hydraulically powered actuator (72) for driving a main control valve (26) of a dual hydraulic servo-actuator control system (10). The actuator includes a tandem piston (74) connected to the main control valve (26) which is controllably positioned by a staged valve (125) of relatively short stroke whereby a force motor (218) of minimum size and energy requirements may be used to directly drive the valve. The staged valve (125) includes a linearly movable valve plunger (184) for simultaneously controlling the differential application of fluid pressure from respective hydraulic system on opposed pressure surfaces (94, 96 and 98, 100) of respective piston sections (90, 92) to cause movement of the piston (74) in response to relatively short axial movement of the valve plunger as long as at least one hydraulic system remains operative. Also, the staged valve (125) includes a fault control valve sleeve (130) concentric with the valve plunger (184) 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 (94, 96 and 98, 100) of the tandem piston to respective returns therefor through respective centering rate control orifices (118,120, 250, 252) in the fault control valve sleeve (130) as the piston in moved to a neutral position by a centering spring device (254) acting on the main control valve.
Description
- 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-hydraulic 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 electro-hydraulic valves located elsewhere in the system housing. Although 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 additional electrical and hydraulic elements necessary to perform the various sensing, equalization, 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 electro-hydraulic control system is an electro-mechanical control system wherein a force motor is coupled 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 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 utilizing 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 which 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 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 required 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-mechanically controlled, hydraulically powered actuator for driving the main control valve of a servo actuator control system. The actuator includes a tandem 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 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 centering 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 hydraulically locked against 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 therewith.
- 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 particularly 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.
- 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-down 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 high loads.
- 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 1, a dual hydraulic servo system is designated generally by
reference numeral 10 and includes two similarhydraulic servo actuators tandem cylinder actuator 16. Theactuator 16 in turn is connected to a control member such as aflight 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 theactuator 16 and hence theflight 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 actuators in the overall system provide a redundancy feature that increases safe operation of the aircraft. - The servo actuators seen in Figure I 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 inlet port 20 for connection with a source of high pressure hydraulic fluid and areturn 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 theservo actuators - In each of the
servo actuators passage 24 connects theinlet port 20 to aservo valve 26. Anotherpassage 28 connects thereturn port 22 to thesame servo valve 26. Eachpassage 24 may be provided with acheck valve 30. - The main
control servo valve 26 includes aspool 32 which is longitudinally shiftable in asleeve 34 which in turn is longitudinally shiftable in thesystem housing 36. The spool and sleeve are divided into two fluidically isolated valving sections indicated generally at 38 and 40 in Figure I, which valving sections are associated respectively with theactuators passages passages passages 42 and 44 in the same servo actuator. - The
passages 42 and 44 of bothservo actuators cylinder tandem actuator 16 which includes a pair ofcylinders 46. Thepassages 42 and 44 of each servo actuator are connected to a corresponding one of the cylinders at opposite sides of thepiston 48 therein. If desired,anti-cavitation valves passages 42 and 44. Thepistons 48 and thecylinders 46 are interconnected by a connectingrod 54 and further are connected byoutput rod 56 to thecontrol element 18 throughlinkage 58. - From the foregoing, it will be apparent that selective relative movement of the
spool 32 andsleeve 34 simultaneously controls both valvingsections 38 and 40 which selectively connect one side of eachcylinder 46 to a high pressure hydraulic fluid source and the other side to fluid return for effecting controlled movement of theoutput rod 56 either to the right or left as seen in Figure 1. In the event one of theservo actuators - The relatively
shiftable spool 32 andsleeve 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 positioning may be effected through direct mechanical linkage to a control element in the aircraft cockpit. As seen in Figure 1, the spool may have acylindrical socket 58 which receives aball 60 at the end of a crank 62. The crank 62 may be connected by a suitable mechanical linkage system to the aircraft 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 thesleeve 34 is controllably shifted to effect the position control function by the hereinafter described control actuation system designated generall;' byreference numeral 70. - The
control actuation system 70 of the invention includes an electro-mechanically controlled, hydraulically poweredactuator 72 which is shown positioned generally in axial alignment with the maincontrol servo valve 26 as seen at the lower left in Figure 1. Theactuator 72 includes atandem piston 74 which is positioned for axial movement in a stepped cylinder bore 76 in thehousing 36. At its end nearest theservo valve 26, thepiston 74 has a piston extension 80 which extends axially in a cylindrical bore 82 of thehousing 36, which bore may be an axial continuation of the cylindrical housing bore 84 accommodating thesleeve 34 andspool 32. The sleeve extension 80 is connected by suitable means to thesleeve 34. The sleeve extension 80 also may have a diametral slot 86 therein which may be engaged by a springbiased plunger 88 to lock theinterconnected piston 74 andsleeve 34 against axial movement. As will be further discussed hereinafter, theplunger 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 92. 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 acylinder pressure surface 98 and an opposedsource pressure surface 100. Also, the corresponding cylinder and source pressure surfaces of the piston sections are opposed and have equal effective pressure areas, respectively. This results in balanced forces acting on piston sections having matched characteristics. - The source pressure surfaces 96 and 100 of the piston sections 90 and 92 respectively are in fluid communication with
passages valves valves passages valves actuators valves passages passages actuators metering orifices - Referring additionally to Figures 2 and 3, the
passages passages valve 125. More particularly, thepassages ports control valve sleeve 130, and theports annular groove 132 andport 134 in astatic porting sleeve 136. The faultcontrol valve sleeve 130 andstatic porting sleeve 136 are concentrically arranged in abore 138 of thesystem housing 36 with the fault control valve sleeve being axially shiftable relative to thehousing 36 and portingsleeve 136, and the porting sleeve being fixed to thehousing 36 against axial movement. - As seen at the right in Figure 2, the porting
sleeve 136 has acylindrical extension 140 which has anend piece 142 axially butted against astop plug 144 fixed in thehousing 36 to prevent axial movement of the portingsleeve 136 to the right as seen in Figure 2. Thecylindrical extension 140 also is diametrically slotted for receipt of a diametrically extendingpin 146 fixed in thehousing 36 which has acentral ball portion 148. Theball portion 148 serves as an axial stop against which bears aplug insert 150 that is fixed in thecylindrical extension 140 by means of apin 152. Accordingly, the indicated engagement of theplug insert 150 against theball portion 148 of thepin 146 prevents axial movement of the portingsleeve 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 portingsleeve 136, is radially stepped along its axial length to provide different thickness valve sleeve portions. As a result, the fault control valve sleeve has a slightly reduced thickness central portion 154 extending between theports port 126 thus providing two differential pressure surfaces 158 and 160 at the right side of each of theports ports sleeve 136 and to its control enabling position seen in Figure 3. - Such shifting of the fault
control valve sleeve 130 is opposed by the force exerted by aspring 162 which is positioned at the right end of thebore 138 and bears in opposition against theend wall 164 of thebore 138 and a shoulder 166 on thevalve sleeve 130. Accordingly, thespring 162 urges the faultcontrol valve sleeve 130 to the left as seen in Figure 2 and towards a radially outwardly extendingflange 168 on the portingsleeve 136 which acts as a stop to define the control disabling position of the fault control valve sleeve when butted thereagainst. On the other hand, theend wall 164 acts as an opposed stop to define the enabling position of the fault control valve sleeve when acylindrical extension 170 on the fault control valve sleeve is butted thereagainst as seen in Figure 3. - When the fault
control valve sleeve 130 is in its enabling position of Figure 3,ports ports sleeve 136 and thepassages ports valve plunger 184. - The
valve plunger 184 of the stagedvalve 125 is concentric with and constrained for axial movement in the portingsleeve 136. The valving section of the valve plunger associated with theport 176 consists ofannular grooves metering land 190. Themetering land 190 is operative to block communication between the associatedport 176 and thegrooves plunger 184 is in a null position. However, upon axial movement of the plunger relative to the portingsleeve 136 and out of its null position, the metering land is operative to effect communication between theport 176 and one or the other of thegrooves - The
groove 186 is in fluid communication with aport 192 in the portingsleeve 136 which in turn communicates with theport 126 when the faultcontrol valve sleeve 130 is in its enabling position of Figure 3. Accordingly, fluid pressure will be supplied to thegroove 186 when thepassage 122 is connected to the supply passage 110 by the shut-downvalve 106. It is noted that at the same time, fluid pressure will be applied on the source pressure surface 96 of the piston section 90. Theother groove 188 is in communication with aport 194 in the portingsleeve 136 which in turn communicates via aport 196 in the faultcontrol valve sleeve 130 with apassage 198 connected to thereturn passage 114 downstream of theorifice 118 as seen in Figure 1. Accordingly, thegroove 188 is connected to the return of the respective or forward hydraulic system. - Similarly, the valving section of the
pilot valve plunger 184 associated with theport 178 has a pair ofannular grooves metering land 204 which is operative in the same manner as themetering land 190 but in association with theport 178. Thegroove 200 is in fluid communication with theport 134 whereas theother groove 202 is in fluid communication with the return passage ll6 of the respective or aft hydraulic system via aport 206 in the porting sleeve,port 208 in the fault control valve sleeve and apassage 210 connected to thereturn passage 116 downstream of theorifice 120 as seen in Figure 1. - The staged
valve 125 also has apassage 212 which connects thepassage 210 to the right or outer end of thebore 138 as seen in Figure 3. Accordingly, the right end face of theplunger 184 will be exposed to return pressure of the aft hydraulic system. Also provided is apassage 214 which connects the right end of thebore 138 to the left end thereof so that the left end face of theplunger 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 pressure areas whereby return pressure variations will not apply unbalanced forces and consequent inputs to the plunger. - It should now be apparent that selective axial movement of the
plunger 184 relative to the portingsleeve 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 fromcylinder pressure surface 98 of piston section 92 to the aft hydraulic system return associated therewith. The resultant pressure imbalance will hydraulically power thepiston 74, and thus the main controlservo 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 thecylinder pressure surface 98 of the piston section 92 from the aft hydraulic system source associated therewith while fluid pressure is released from the cylinder pressure surface 94 of the piston section 90 to the forward hydraulic system return associated therewith. Under these conditions, the resultant pressure imbalance will hydraulically power thepiston 74 andvalve 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 pressure 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 movement of the
valve plunger 184 may be effected by aforce motor 218 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. - The
force motor 218 includes amotor housing 220 which is secured in thesystem housing 36 closely adjacent one end of thevalve plunger 184 with itsdrive shaft 222 extending perpendicularly to a plane through the longitudinal axis of the valve plunger. The drive shaft is shown drivingly 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 anaxial bore 226 through which the quill extends for connection at its threadedend 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 thedrive shaft 222, such other end being provided with aball bearing 232 engaged by aneccentric 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 atransverse bore 236 in the quill. - The
quill 224 may have acylindrical portion 238 and a reduced diameterflexible length portion 240. The cylindrical portion extends from the threadedend 228 of the quill about half way through theplunger 184 and is closely fitted in theaxial bore 226 whereby flexing of the quill is limited to the reduceddiameter portion 240. It will be appreciated that the flexingportion 240 of the quill accommodates the rise and fall of thebearing 232 without applying significant side loads to the tubular valve plunger as theeccentric pin 234 is driven by theforce motor 218 through a short arcuate stroke. It also is noted that the effective length of the quill may be adjusted at its threadedend 228 for adjusting the neutral or null position of the plunger relative to a null position of the force motor. - The
valve plunger 184 alternatively may be driven by a linear force motor. For this, thequill 224 is provided at its force motor connection end with a threadedaxial 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 misalignment between such drive member and the valve plunger without applying significant side loads to the valve plunger. - As indicated, controlled selective movement of the
valve plunger 184 is effected by theforce 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, aposition transducer 246 may be operatively connected to theactuator piston 74 as schematically shown in Figure I. 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 theactuator piston 74 which may be required to have a relatively long stroke for driving the maincontrol 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 disabling position seen in Figure 2. Such shifting will occur whenever the fluid pressure acting upon the differential pressure surfaces 158 and 160 of thevalve sleeve 130 at theports spring 162. This will occur upon failure of both independent hydraulic systems or upon shut-down of the electrical operational mode by the shut-downvalves spring 162 will shift the faultcontrol valve sleeve 130 to its disabling position whereat the inner end of the valve sleeve will be butted against theflange 168 on the portingmember 136. - When the fault
control valve sleeve 130 is in its disabling position, communication between the cylinder pressure surfaces 94 and 98 and therespective supply passages valve plunger 184. More particularly, the fault control valve sleeve in its disabling position blocks communication between thepassage 122 and theport 192 which otherwise cooperate to supply fluid pressure to thegroove 186 associated with the cylinder pressure surface 94. Also, the fault control valve sleeve blocks communication between thepassage 182 andport 178 associated with thecylinder pressure surface 98. - In addition, the fault
control valve sleeve 130 in its disabling position blocks communication between the cylinder pressure surfaces 94 and 98 and therespective return passages valve plunger 184, except through respective centering rate control ormetering orifices valve sleeve 130 to its disabling position blocks communication between theport 176 andport 172 while establishing communication between thepassage 180 andpassage 198 via themetering orifice 250. At the same time, communication between thepassage 182 andport 178 is blocked as indicated above while communication between thepassage 182 andpassage 210 is established via themetering orifice 252. - As a result, fluid pressure from the cylinder pressure surfaces 94 and 98 will be released to the
return passages metering orifices control valve sleeve 130 which control the rate at which fluid is ported from the cylinder pressure surfaces as the main controlservo valve sleeve 34 and thus thepiston 74 is moved to a centered or neutral position by aspring centering device 254 for system operation in the manual mode. Thespring centering device 254 can be seen at the right in Figure 1 and may be conventional. - During normal operation of the control actuation system in the electrical mode, each shut-down
valve biased plunger 256 seen in Figure 1. This moves theplunger 256 to the right against the biasing force to open thepassage 258 to fluid pressure for moving theplunger 88 out of locking engagement with the sleeve extension 80 thereby to permit axial movement of the maincontrol valve sleeve 34 and thepiston 74. - Fluid pressure also will be supplied to the
ports passages control valve sleeve 130 will be shifted from its disabling position of Figure 2 to its enabling position of Figure 3. With the fault control valve sleeve in its enabling position, controlled positioning of thepiston 74 and hence the maincontrol valve sleeve 34 may be effected by thevalve plunger 184 and forcemotor 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 relationships. - Position control of the
piston 74 and maincontrol valve sleeve 34 will be maintained even though one of the channels of the electrical mode fails or is rendered inoperative. However, if both channels fail or are rendered inoperative requiring reversion to the manual operational mode, both shut-downvalves control valve sleeve 130 to its disabling position shown in Figure 2. As the maincontrol valve sleeve 34 is urged towards its centered or neutral position by the centeringspring 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 centeringrate control orifices orifices piston 74 orvalve 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 centeringspring device 254 and also by the lockingplunger 88 which will then have moved into locking engagement with the slot 86 in the sleeve extension 80. In the unlikely event that a relatively large reaction force is applied on thevalve 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 resistive 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 certain 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.
Claims (10)
1. 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 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 axial movement of said piston in opposite directions in response to relatively short, directly driven linear movement of said valve plunger in opposite directions.
2. A control actuation system as set forth in claim 1, said opposed pressure surfaces of each piston section being opposed to corresponding pressure surfaces of the other piston section, further comprising respective means for supplying fluid pressure from such respective sources thereof to said actuator and staged valve 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, and means responsive to system shut-down for releasing fluid pressure acting on opposed corresponding pressure surfaces 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.
3. A system as set forth in claim I, wherein said opposed pressure surfaces of each piston section have unequal effective pressure areas, and means are provided for applying fluid pressure from such respective sources thereof normally only on the smaller area pressure 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, said smaller and larger area pressure surfaces of each piston section being axially opposed to and having effective pressure areas equal to corresponding pressure surfaces of the other piston section.
4. A system as set forth in claim 3, further comprising a fault control valve member axially 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.
5. A system as set forth in claim 4, further comprising shut-down means operable to release fluid pressure acting on said smaller area pressure surfaces of respective said piston sections to respective returns therefor, centering means for resiliently urging said piston to a neutral position upon operation of said shut-down means, and means for urging said fault control valve member to the disabling position thereof upon such release of fluid pressure by said shut-down means.
6. A system as set forth in claim 5, wherein said fault control valve member has porting means operative in the disabling position of said fault control valve member to release fluid pressure from said larger area pressure surfaces of said piston sections to respective returns therefor through respective centering rate control orifices to control the rate at which said piston is moved to the neutral position thereof by said centering means.
7. A system as set forth in claim 6, wherein said centering rate control orifices are located in said fault control valve member, said fault control valve member and valve plunger are concentrically arranged in said staged valve means, and said fault control valve member is in the form of a sleeve surrounding said valve plunger.
8. A system as set forth in claim 4, wherein said means for moving said fault control valve member includes two differential pressure areas on said fault control valve member, and means for communicating said differential pressure areas with fluid pressure applied to said smaller area pressure surfaces, respectively.
9. A control actuation system useful in a hydraulic servo actuator control system for operating a control valve element therein, comprising an actuator, a piston axially movable in said actuator and drivingly connectable to the control valve element, staged valve means operably connectable to control input means for effecting position control of said piston, said staged valve means including 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 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.
10. A system as set forth in claim 9, wherein said one side of said piston has a larger area pressure surface than the other side, and means are provided for normally applying such pressure fluid only on the smaller area pressure surface of said piston, said valve plunger being selectively movable either to admit fluid pressure to said larger area pressure surface or to release fluid pressure acting on said larger area pressure surface for pressure actuation of said piston in opposite directions, said means for releasing further including valve means responsive to such control means being rendered inoperative for precluding such normal application of fluid pressure on said smaller area pressure surface and for releasing fluid pressure on said smaller area pressure surface through a respective other of said metering orifices, said fault control valve member when in said position precluding such admission and release of fluid pressure and when in another position permits such admission and release, means for resiliently urging said fault control valve member to said position, and means responsive to such normal application of fluid pressure on said smaller area pressure surface for moving said fault control valve member to said another position thereof against said means for resiliently urging, said means for moving including opposed pressure surfaces on said fault control valve member of different effective pressure areas in fluid communication with said means for normally applying.
Applications Claiming Priority (2)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
US46363183A | 1983-02-03 | 1983-02-03 | |
US463631 | 1983-02-03 |
Publications (1)
Publication Number | Publication Date |
---|---|
EP0115925A1 true EP0115925A1 (en) | 1984-08-15 |
Family
ID=23840777
Family Applications (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
EP84300329A Ceased EP0115925A1 (en) | 1983-02-03 | 1984-01-19 | Control actuation system including staged direct drive valve with fault control |
Country Status (4)
Country | Link |
---|---|
EP (1) | EP0115925A1 (en) |
JP (1) | JPS59175603A (en) |
CA (1) | CA1203145A (en) |
IL (1) | IL70970A0 (en) |
Cited By (2)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
EP3335987A1 (en) * | 2016-12-16 | 2018-06-20 | Microtecnica S.r.l. | Integrated stability and control augmentation system |
CN110844046A (en) * | 2018-08-20 | 2020-02-28 | 古德里奇驱动系统有限公司 | Actuator system |
Citations (5)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
GB917808A (en) * | 1960-05-12 | 1963-02-06 | Fairey Eng | Improvements relating to servo systems |
US3253613A (en) * | 1963-07-01 | 1966-05-31 | Boeing Co | Fail safe servo valve |
US3543642A (en) * | 1969-05-28 | 1970-12-01 | Us Navy | Unitized control module for a hydraulic actuation apparatus |
US4090429A (en) * | 1975-07-21 | 1978-05-23 | Teijin Seiki Company Limited | Fail-safe fluid control valve |
WO1982000862A1 (en) * | 1980-09-02 | 1982-03-18 | Rockwell International Corp | Actuator system for a control surface of an aircraft |
-
1984
- 1984-01-18 CA CA000445533A patent/CA1203145A/en not_active Expired
- 1984-01-19 EP EP84300329A patent/EP0115925A1/en not_active Ceased
- 1984-02-02 JP JP59016175A patent/JPS59175603A/en active Pending
- 1984-02-15 IL IL70970A patent/IL70970A0/en unknown
Patent Citations (5)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
GB917808A (en) * | 1960-05-12 | 1963-02-06 | Fairey Eng | Improvements relating to servo systems |
US3253613A (en) * | 1963-07-01 | 1966-05-31 | Boeing Co | Fail safe servo valve |
US3543642A (en) * | 1969-05-28 | 1970-12-01 | Us Navy | Unitized control module for a hydraulic actuation apparatus |
US4090429A (en) * | 1975-07-21 | 1978-05-23 | Teijin Seiki Company Limited | Fail-safe fluid control valve |
WO1982000862A1 (en) * | 1980-09-02 | 1982-03-18 | Rockwell International Corp | Actuator system for a control surface of an aircraft |
Cited By (3)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
EP3335987A1 (en) * | 2016-12-16 | 2018-06-20 | Microtecnica S.r.l. | Integrated stability and control augmentation system |
US10836469B2 (en) | 2016-12-16 | 2020-11-17 | Microtecnica S.R.L. | Integrated stability and control augmentation system |
CN110844046A (en) * | 2018-08-20 | 2020-02-28 | 古德里奇驱动系统有限公司 | Actuator system |
Also Published As
Publication number | Publication date |
---|---|
CA1203145A (en) | 1986-04-15 |
IL70970A0 (en) | 1984-05-31 |
JPS59175603A (en) | 1984-10-04 |
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