EP0115925A1 - Kontrollbetätigungssysten mit direkt angetriebenem Stufenventil und Fehlerkontrolle - Google Patents

Kontrollbetätigungssysten mit direkt angetriebenem Stufenventil und Fehlerkontrolle Download PDF

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Publication number
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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EP
European Patent Office
Prior art keywords
piston
pressure
valve
control valve
control
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Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Ceased
Application number
EP84300329A
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English (en)
French (fr)
Inventor
Robert D. Vanderlaan
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Pneumo Abex Corp
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Pneumo Corp
Pneumo Abex Corp
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Filing date
Publication date
Application filed by Pneumo Corp, Pneumo Abex Corp filed Critical Pneumo Corp
Publication of EP0115925A1 publication Critical patent/EP0115925A1/de
Ceased legal-status Critical Current

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Classifications

    • 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.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Fluid-Pressure Circuits (AREA)
  • Servomotors (AREA)
EP84300329A 1983-02-03 1984-01-19 Kontrollbetätigungssysten mit direkt angetriebenem Stufenventil und Fehlerkontrolle Ceased EP0115925A1 (de)

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 (de) 1984-08-15

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Application Number Title Priority Date Filing Date
EP84300329A Ceased EP0115925A1 (de) 1983-02-03 1984-01-19 Kontrollbetätigungssysten mit direkt angetriebenem Stufenventil und Fehlerkontrolle

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EP (1) EP0115925A1 (de)
JP (1) JPS59175603A (de)
CA (1) CA1203145A (de)
IL (1) IL70970A0 (de)

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP3335987A1 (de) * 2016-12-16 2018-06-20 Microtecnica S.r.l. Integriertes stabilitäts- und steuerungsverstärkungssystem
CN110844046A (zh) * 2018-08-20 2020-02-28 古德里奇驱动系统有限公司 致动器系统

Citations (5)

* Cited by examiner, † Cited by third party
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

Patent Citations (5)

* Cited by examiner, † Cited by third party
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 (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP3335987A1 (de) * 2016-12-16 2018-06-20 Microtecnica S.r.l. Integriertes stabilitäts- und steuerungsverstärkungssystem
US10836469B2 (en) 2016-12-16 2020-11-17 Microtecnica S.R.L. Integrated stability and control augmentation system
CN110844046A (zh) * 2018-08-20 2020-02-28 古德里奇驱动系统有限公司 致动器系统
CN110844046B (zh) * 2018-08-20 2024-05-28 古德里奇驱动系统有限公司 致动器系统

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Publication number Publication date
JPS59175603A (ja) 1984-10-04
CA1203145A (en) 1986-04-15
IL70970A0 (en) 1984-05-31

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