EP0256648A2 - Système de commande hydraulique - Google Patents

Système de commande hydraulique Download PDF

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Publication number
EP0256648A2
EP0256648A2 EP87305960A EP87305960A EP0256648A2 EP 0256648 A2 EP0256648 A2 EP 0256648A2 EP 87305960 A EP87305960 A EP 87305960A EP 87305960 A EP87305960 A EP 87305960A EP 0256648 A2 EP0256648 A2 EP 0256648A2
Authority
EP
European Patent Office
Prior art keywords
valve
control
hydraulic
control system
electrical
Prior art date
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.)
Withdrawn
Application number
EP87305960A
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German (de)
English (en)
Other versions
EP0256648A3 (fr
Inventor
Frederick James Fuell
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.)
Claverham Holdings Ltd
Original Assignee
Fairey Hydraulics Ltd
Priority date (The priority date 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 date listed.)
Filing date
Publication date
Priority claimed from GB868619412A external-priority patent/GB8619412D0/en
Application filed by Fairey Hydraulics Ltd filed Critical Fairey Hydraulics Ltd
Publication of EP0256648A2 publication Critical patent/EP0256648A2/fr
Publication of EP0256648A3 publication Critical patent/EP0256648A3/fr
Withdrawn legal-status Critical Current

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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

  • the present invention relates to a hydraulic control system.
  • the invention is particularly concerned with the control of movable members provided with first and second actuators which may be arranged, for example, in parallel or in tandem.
  • the movable member may be a movable load of substantial value which requires to be positioned and moved with a high degree of accuracy and reliability.
  • the movable member may be a hydraulic control valve.
  • a particular but not exclusive application of such hydraulic control valves is in the control of actuators for use in moving control surfaces of aircraft.
  • the methods and systems according to the invention are applicable for the control of high performance aircraft and the systems may be constructed to be of high integrity and to provide multi-redundant electro­hydraulic actuation.
  • hydraulic actuators In many applications of hydraulic actuators, it is desirable to position a movable load of some several tonnes with a high degree of accuracy while maintaining a high degree of protection against failure in the hydraulic system or its control system. Many such actuators are required to be controlled remotely by way of electrical signals from a remote control point and it is necessary to provide redundancy to accommodate the failure of various components in the hydraulic system itself or the electrical control system and hydraulic valves associated with it, so that control of the actuator may be maintained in the event of such a failure.
  • One particular, but not exclusive example of such an actuator is a hydraulic actuator used to effect the movement of an aircraft control surface, particularly a high speed aircraft.
  • the usual design philosophy in such multi-­redundant system is to provide an arrangement which can survive at least two failures, one of which may be hydraulic. This requires at least three electrical lanes and duplex hydraulic systems. Two electrical lanes are insufficient, because it is desirable to be able to identify a faulty lane by comparing it with the remaining good lanes. With a total of only two lanes, the faulty lane could not be eliminated in this way. Two hydraulic systems are sufficient because a hydraulic failure will simply lead to loss of system pressure and no advantage is gained by comparing one hydraulic lane with another.
  • a potential disadvantage with multi-redundant systems of this type is the difficulty of correctly matching all the electrical and hydraulic lanes with each other to prevent "force-fighting" and parasitic loss as will be described hereinafter.
  • the primary, potentially catastrophic type is where one electrical lane receives a large faulty signal and completely overpowers the remaining lanes.
  • the secondary type may arise from natural differences which will exist between the control lanes arising from tolerances of manufacture and assembly.
  • Parasitic loss may arise where two hydraulic control valves are connected in parallel between a source of hydraulic pressure and an actuator. If the zero or no flow positions of the valves are not exactly matched, one valve may be slightly open while the other is shut. This would lead to undesired actuator movement. In practice, because position feed-back is employed, the system sets itself so that the two valves are each slightly open in opposite senses. This results in a small flow of hydraulic fluid through the two valves to the return line. This is known as parasitic flow and represents a power loss.
  • a known control system is illustrated by European Patent Application No. EP-A-0092972 of the present applicants.
  • This system proposes that, between the main valve which is to be controlled and the four electrical control lanes conventionally provided in a high performance aircraft, duplex hydraulic control systems are provided comprising first and second actuators for moving the main valve, each of which actuators is controlled by a pair of hydraulically parallel-connected electrohydraulic spool valves.
  • Parasitic flow between the valves of the respective pairs is avoided by providing one valve of each pair with a significant overlap at the zero point, so that no flow is provided for a significant range of spool movement either side of the zero point.
  • Each electrohydraulic valve comprises a so-called “flapper” or “jet pipe” which in response to an electrical input moves between a pair of orifices or receivers respectively and thus controls the flow through these orifices.
  • This flow control is used to vary the pressure conditions at each end of the spool and thus controls the spool movement.
  • the valve therefore requires a source of hydraulic fluid pressure, and commercially available valves are arranged also to control hydraulic flow, by means of the spool, from the same source as that required for valve operation.
  • Such an electrohydraulic valve will operate on and control a single hydraulic supply. Therefore, with only two hydraulic supplies available, it clearly is best to connect each supply only to two valves. If one hydraulic supply should fail, two valves would still be operational.
  • each valve would automatically eliminate the effectiveness of two lanes, even although the electrical signals on the lanes might be functioning correctly.
  • each lane is connected to two valves supplied by respective hydraulic supplies. To permit this, each valve needs two operating windings, resulting in a complex circuit arrangement.
  • a disadvantage of this arrangement is that a fault in one electrical lane adversely affects two valves, so that, in the worst case, only two electrical lane failures could cause all four valves to malfunction.
  • a further disadvantage with this arrangement is that if one hydraulic supply should fail so that the system is forced to rely on the other hydraulic supply, only one pair of electro-hydraulic valves would be operational. Thus, failure of one hydraulic supply automatically eliminates one pair of valves and their corresponding driving mechanisms. Therefore, should one of the two hydraulic supplies fail, the entire burden of controlling the fluid supplied to the actuators is borne by the driving mechanisms of only one of the two pairs of valves. This is clearly disadvantageous, since it is likely to result in a substantial degradation of performance.
  • a tandem spool valve directly driven by several high power electrical torque motors is particularly susceptible to this force-fighting problem owing to the fact that the output of each torque motor is not limited in any particular way and will increase in dependence upon the size of an input control current.
  • an abnormally high control current is supplied as a result of a system fault, one torque motor would, in the worst case, overpower the remaining three torque motors and lead to complete system failure.
  • This type of failure is particularly associated with systems directly controlled by means of electrical current.
  • a hydraulic control system comprising first and second actuators operating in parallel or tandem and being separately supplied with fluid along first and second fluid paths, and control valve means including at least three independent electrical actuating means, the control valve means being arranged for operation, on actuation by any one or more of the electrical actuating means, for the common control of the fluid respectively in the first and second paths.
  • the electrical actuating means comprise at least three electrical coils, or other windings.
  • the control valve means may comprise a single valve having at least three coils for operation thereof; alternatively, it may comprise at least two control valves, each having at least two coils for operation thereof; alternatively again, it may comprise at least three control valves, each having exactly one coil for operation thereof.
  • the control valve means may comprise one or more direct drive valves, either driven by an electrical torque motor or an electrical force motor. Alternatively, the control valve means may comprise at least one electro-hydraulic valve.
  • control valve means comprises at least one spool valve, the or each valve having a spool extending between the first and second paths.
  • the actuators may comprise a further fluid valve which may be, for example, a spool valve.
  • This further fluid valve may be arranged to control fluid within two further independent fluid paths. The fluid in these two further paths may be arranged to control a further actuator or spool.
  • control valve means is arranged for operation by exactly four electrical coils, each coil being independently energised via a separate electrical pathway.
  • failure of one hydraulic supply does not automatically eliminate any part of the valve control means, for example any one of the individual valves which may make up the valve control means, since the two hydraulic supplies are each connected to all of the valves. In other words, the two hydraulic supplies are connected in parallel.
  • the driving mechanism for example the torque motor
  • direct drive valves driven by torque motors or force motors
  • the disadvantages of electro-hydraulic valves are avoided.
  • the present invention also avoids the use of special fault detecting equipment, which is a feature of many present arrangements.
  • a valve assembly 1 comprises a main valve 2 arranged to be driven by first and second actuators 12a and 12b. Each actuator is connected to each of four control valves 3a, 3b, 3c and 3d having respective electrical torque motors 4a, 4b, 4c and 4d connected for direct drive of the respective control valves.
  • the main valve 2 is a spool valve and connected to its spool are first and second position feed back transducer assemblies 6a and 6b each of which preferably comprises a pair of linear variable differential transformers (hereinafter LVDT).
  • LVDT linear variable differential transformers
  • FIG. 2 is a cut-away version of Figure 1 thus enabling the spool 8 of main valve 2 to be seen and also permitting the individual LVDT's 11a, 11b, 11c and 11d to be seen.
  • each control valve 3a, 3b, 3c and 3d has a respective valve spool 9 and that each valve spool 9 is directly connected to the shaft of a respective one of the torque motors 4a to 4d which have respective coils 10a to 10d, and rotary feed-back transducers 20a to 20d for closed loop servo control of position.
  • Each torque motor operates through a limited angle in the range of 5 to 30° and thereby causes linear motion of the respective valve spool by means of a respective spherical ball joint 21a,21b,21c or 21d between the motor shaft and the spool which is offset from the axis of rotation of the motor.
  • the spherical ball is not illustrated in the Figures.
  • each valve spool 9 is provided with a return spring and in addition or as an alternative may have multi-redundant electrical positional feedback for closed loop servo control.
  • FIG. 3A and 4 the interconnection of the various components of the valve assembly may be seen schematically.
  • these figures also illustrate a second stage or main actuator 13 provided with quadruplex feedback transducers 17, preferably LDVT's, for closed loop servo control of position.
  • each of the four first stage valves 3a to 3d is connected to control the first stage actuator 12a via control lines C1 and is also connected to control the second actuator 12b via control lines C2 which are independent of control lines C1.
  • the first stage actuators 12a and 12b directly control the second stage valves 8, which may be referred to as the main valves, which in turn control via two independent control lines C3 and C4 two independent hydraulic piston and cylinder assemblies of the second stage main actuator 13.
  • Figures 3B and 3C show, respectively, an arrangement in which two first stage valves are provided, each having two coils (and, possibly, two separate corresponding force motors or torque motors), and an arrangement in which a single first stage valve is provided, this valve having four coils (and, possibly, four independent torque or force motors).
  • FIG 5 corresponds to Figures 3A and 4 and which shows further detail of the construction of the torque motors 4a to 4b, further detail of the connection of the hydraulic lines and further internal detail of the first stage valves, second stage valves and first and second stage actuators.
  • the hydraulic fluid pressure is preferably 27 MN/m2 (4000 psi nominal).
  • each of the control valves 3a to 3d provides two independently controllable hydraulic porting arrangements on a common spool.
  • Each porting arrangement is connected to a respective one of the hydraulic supplies P1 and P2.
  • each of the supplies P1 and P2 is connected to one side of a respective one of the first stage actuators 12a and 12b.
  • the other side of each of the first stage actuators 12a and 12b is connected to a respective one of the hydraulic porting arrangements of each of the valves 3a to 3d.
  • each hydraulic porting arrangement of each of the first stage valve 3a to 3d is such as to reduce the system pressure by approximately half and to supply this to one side of each of the first stage actuators 12a and 12b when the spool 9 is in its undisplaced or central position.
  • each of the actuators 12a and 12b is provided with system pressure on one side and 50% of system pressure on the other side in the neutral position.
  • the actuators are balanced by arranging for the unequal pressures to be applied to unequal areas in the ratio of approximately 2:1.
  • each first stage valve moves such that the pressure to the larger area (to which it is connected) is either increased or reduced thus providing a net force to move the main valve spool 8.
  • each of the control valves 3a to 3d requires two three-port configurations. It is equally feasible to use two four-port arrangements and in this case the first stage actuators will have equal piston areas and the two active chambers will be controlled differentially.
  • the main valves are arranged on a common tandem spool 8 and are each arranged to control a respective hydraulic piston 14a or 14b of the main actuator 13.
  • a conventional 4-port arrangement is employed and as the spool 8 displaces pressure on one side of each piston 14a and 14b tends to increase whilst it tends to reduce on the other side.
  • the pistons 14a and 14b are connected on a common hollow shaft 15 in a housing 16.
  • Quadruplex feedback transducers 17, preferably LDVT's, are provided within the shaft 15 for position feedback control.
  • each first stage valve is a duplex arrangement and a multi-redundant system is obtained by the addition of several such duplex valves by flow summation to control the first stage actuators 12a and 12b.
  • each of the torque motors 4a to 4d drives a tandem spool arrangement.
  • mismatch between the several first stage valves 3a to 3d will be minimized by accurate mechanical adjustment of the hydraulic and electrical datums to ensure that these are closely coincident. Furthermore, any residual electrical mismatch between the lanes may be minimized by an equalisation technique which reduces or eliminates the level of the steady state motor current.
  • Hydraulic integrity is provided by a duplex tandem arrangement throughout.
  • Integrity of the system is enhanced by the fact that the multi redundant electrical control systems are totally separated at the motors and the motors themselves are also physically separated. An electrical hardover of one motor leading to a hardover of the associated control valve cannot overpower the remaining motors because the motors are not connected to be force summing.
  • Mechanical integrity is provided at the first stage by a similar philosophy to that applied to the electrical integrity. If a first stage servo valve is mechanically jammed, the flow summation technique employed ensures that the remaining valves can overpower the jammed valve and that the system as a whole can continue to operate. Thus, the system can tolerate a single valve electrical or mechanical hardover without immediate corrective action being required. It is clearly a prerequisite for this advantage to be achieved at least three first stage valves are provided.
  • Integrity of the second stage valve is ensured by the provision of a sufficiently large first stage actuator area and force to overcome any definable jam condition.
  • direct-drive techniques have been employed in a way which leads to no loss of system integrity or reliability and no loss of redundancy.
  • the advantages of the direct-drive technique may therefore be achieved without suffering the disadvantages previously associated with this approach.
  • any electrically operated drive means may be used for the first stage valves 3a to 3d.
  • direct-drive torque motors and electro­hydraulic valves providing indirect drive may be employed, but also linear motors or solenoid type actuation systems acting directly or indirectly on the valve spools.

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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)
EP87305960A 1986-08-08 1987-07-06 Système de commande hydraulique Withdrawn EP0256648A3 (fr)

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
GB8619412 1986-08-08
GB868619412A GB8619412D0 (en) 1986-08-08 1986-08-08 Controlling a movable member
GB8705071 1987-03-04
GB878705071A GB8705071D0 (en) 1986-08-08 1987-03-04 Hydraulic control system

Publications (2)

Publication Number Publication Date
EP0256648A2 true EP0256648A2 (fr) 1988-02-24
EP0256648A3 EP0256648A3 (fr) 1989-11-02

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Family Applications (1)

Application Number Title Priority Date Filing Date
EP87305960A Withdrawn EP0256648A3 (fr) 1986-08-08 1987-07-06 Système de commande hydraulique

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EP (1) EP0256648A3 (fr)

Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0639499A1 (fr) * 1993-08-20 1995-02-22 Lucas France Dispositif de servocommande d'un organe de commande de vol d'aéronef
ES2156497A1 (es) * 1998-06-23 2001-06-16 Turbo Propulsores Ind Mejoras en el objeto de la patente 9801320, relativa a "sistema principal de servo-actuacion de piston con deteccion hidromecanica de fallos autocontenida.
KR20160138674A (ko) * 2015-05-26 2016-12-06 한국항공대학교산학협력단 비상 안전 밸브
US12435740B2 (en) 2021-11-12 2025-10-07 Woodward, Inc. Hydraulic force fight mitigation

Family Cites Families (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE1904534A1 (de) * 1969-01-30 1970-09-10 Messerschmitt Boelkow Blohm Einrichtung zur elektro-hydraulischen Steuerung eines hydraulischen Arbeitskolbens
GB1272859A (en) * 1969-11-25 1972-05-03 Hobson Ltd H M Improvements in position control servo systems
GB2082799B (en) * 1980-08-27 1983-11-23 Elliott Brothers London Ltd Hydraulic actuator systems

Cited By (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0639499A1 (fr) * 1993-08-20 1995-02-22 Lucas France Dispositif de servocommande d'un organe de commande de vol d'aéronef
FR2709110A1 (fr) * 1993-08-20 1995-02-24 Lucas Air Equipement Dispositif de servocommande d'un organe de commande de vol d'aéronef.
US5600220A (en) * 1993-08-20 1997-02-04 Lucas France System for servo-controlling an aircraft flight control member
ES2156497A1 (es) * 1998-06-23 2001-06-16 Turbo Propulsores Ind Mejoras en el objeto de la patente 9801320, relativa a "sistema principal de servo-actuacion de piston con deteccion hidromecanica de fallos autocontenida.
KR20160138674A (ko) * 2015-05-26 2016-12-06 한국항공대학교산학협력단 비상 안전 밸브
KR101711419B1 (ko) * 2015-05-26 2017-03-13 한국항공대학교산학협력단 비상 안전 밸브
US12435740B2 (en) 2021-11-12 2025-10-07 Woodward, Inc. Hydraulic force fight mitigation

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Publication number Publication date
EP0256648A3 (fr) 1989-11-02

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Inventor name: FUELL, FREDERICK JAMES