|Publication number||US3741073 A|
|Publication date||Jun 26, 1973|
|Filing date||Jan 29, 1971|
|Priority date||Jan 29, 1971|
|Also published as||DE2203848A1, DE2203848B2, DE2203848C3|
|Publication number||US 3741073 A, US 3741073A, US-A-3741073, US3741073 A, US3741073A|
|Original Assignee||Moog Inc|
|Export Citation||BiBTeX, EndNote, RefMan|
|Referenced by (12), Classifications (10)|
|External Links: USPTO, USPTO Assignment, Espacenet|
United States Patent [191 Garnjost June 26, 1973 1 HYSTERETIC EQUALIZATION IN REDUNDANT ELECTRICALLY OPERATED FLUID POWERED SERVOPOSITIONING APPARATUS  Inventor: Kenneth D. Garnjost, East Aurora,
 Assignee: Moog lnc., East Aurora, NY.
 Filed: Jan. 29, 1971  Appl. No.: 110,927
 US. Cl. 91/363 A, 91/411 B, 91/433  Int. Cl. FlSb 9/03, F15b 9/09  Field of Search 91/363 A, 384, 433, 91/411 B, 363 R, 411 A, 411 R [56 References Cited UNITED STATES PATENTS 2,995,014 8/1961 Horky et al 91/433 3,190,185 6/1965 Rasmussen 91/363 A 3,272,062 9/1966 Flippo et al 91/384 3,398,647 8/1968 Baltus et a1. 91/384 3,487,750 1/1970 Borgeson 91/384 Primary Examiner-Paul E. Maslousky Attrney--Sommer & Weber 7] ABSTRACT Hysteretic equalization means are disclosed for equal- 12 Claims, 16 Drawing Figures CHANNEL A CHANNEL B 2: CHANNEL C i 26:: 2s 26 b 26b 45c 25c SERVOVALVE ze SERVOVALVE SERVOg ALVE 4 28c A a HYSTERETIC 42 EQUALIZER 2 HYSTERETIC C EOUALIZER EQUALlZATION .HYSTERETIC B FEEDBACK EQUALIZER EQUALIZAT|ON A FEEDBACK 32c EQUALIZATION FEEDBACK -ww 4| POSITION ACTUATOR FEEDBACK P33 POSITION AC1 FEEDBACK r 30b 35b POSITION A T 35 M FEEDBACK 30 A I as LOAD DRIVE PRESSURE LOAD DRIVE PREssuRE LOAD DRIVE PRESSURE PATENTEU I973 SIIEEIIIIF4 CHANNEL 1 0 NULL ACTUATOR I *rPosmoN I /\CHANNEL 2 ERROR NULL :5 l'
s MNET ACTUATOR PRESSURE GAIN CT-IANNEL 3 NuLL ACTUATOR POSITION ERROR CHANNEL 2 NULL NET ACTUATOR PRESSURE GAIN LOAD FORCE REQUIRED A ACTUATOR POSITION ovERALL FORCE GAIN CHANNEL 1 FORCE GAIN CHANNEL 2 FORCE GAIN AcTuAToR POSITION ERRoR g2 F1 0. I-O 5E I-m 22 COMMAND INPUT LOAD DETE NT ATTORNEYS PAIENIEDJuues ms SHEEIJIYd INVENTOR.
AT TOR NE YS nvm nmw
PATENTEUJUN 2 6 I975 SHEH SBFII INVENTOR.
Kenneth D. Garn BY jost ATTORNE YS IIYSTERETIC EQUALIZATION IN REDUNDANT ELECTRICALLY OPERATED FLUID POWERED SERVOPOSITIONING APPARATUS This invention relates to apparatus for equalizing the outputs of two or more servovalves in a redundant electrically operated fluid powered servopositioning apparatus. This is achieved by providing, for each channel, equalizer means having a hysteresis characteristic for generating and feeding back to an electrical signal summer for a given channel an equalization signal responsive to the output of the servovalve of such given channel.
A number of outstanding advantages attend the present invention. One advantage is that no electrical nor hydraulic interchannel connections are required since these involve a common power supply that introduces a potential common failure mode.
Another advantage of the present invention is that no change in logic such as through the use of switches, shuttle valves and the like, is required when shutting off one channel of a redundant system, such as going from four channels to three, and then three channels to two,
and so on.
A further advantage of the present invention is that the redundant arrangement is allowed to follow commands throughout its frequency range, as contrasted with some'prior art equalization schemes which do not work well for very low frequency cyclic commands.
A still further advantage of the present invention is that the high pressure gain associated with common twostage electrohydraulic flow control servovalves is retained, this being necessary for good system resolution and accuracy.
Yet another advantage of the present invention is that identical equalization devices or instrumentalities, regardless of type, can be used in each channel of the redundant system.
The present invention improves a redundant electrically operated fluid powered servopositioningapparatus having at least two channels each including a servovalve, signal summing means through which a command input signal'is fed, an actuator operatively associated with the output of the servovalve and arranged to drive a common load member and feedback means arranged to provide a load position feedback signal to the signal summing means, the output of said signal summing means being fed to said servovalve, by equalizing the outputs of the various servovalves achieved through the use, for each channel, of equalizer means having a hysteresis characteristic for generating and feeding back to the signal summing means of a given channel an equalization signal responsive to the output of the servovalve of such given channel.
Other objects andadvantages of the present invention will be apparent from the following description of two embodiments of the means for providing a hyster- Y etic equalization signal as shown in the accompanying drawings wherein:
FIG. 1 is a diagram indicating for a two-channel system the relationship between load drive pressure as a function of supply pressure P, and actuator position error and showing the net actuator pressure gain.
FIG. 2 is a similar diagram except for a three-channel system.
FIG. 3 is a diagram similar to FIG. 1 except showing the pressure gain characteristic for a two-channel system employing proportional pressure feedback servovalves.
FIG. 4 is a diagram indicating the relationship be tween load force required and actuator position for a system having a high force gradient in the region of force polarity reversal.
FIG. 5 is a diagram similar to FIG. 1 and illustrates the channel force as developed at error position HP, or EP required to breakout the load on'either side of the actuator null.
FIG. 6 is a diagram showing the relationship between command input and actuator position and depicting the width of the deadzone.
FIG. 7 is a diagram indicating the relationship between equalization force level and actuator position error and illustrating how an equalization signal will draw the null of one channel toward that of another.
FIG. 8 is a diagram showing the relationship between equalization signal and load differential pressure AP for a two-channel system having equalization in accordance with the present invention and indicating at points IE8; and 2ES the equal differential pressure outputs of the channels when the load is zero.
FIG. 9 is a diagram similar to FIG. 8 but indicating the condition of the actuator equalization signals at points lES and 2ES when a positive differential pressure is required to overcome the load.
FIG. 10 is a diagram similar to FIG. 9, but indicating the condition of the equalization signals at points lES and H58 when a reversal of the direction of fluid drive occurs.
FIG. 11 is a schematic of a redundant electrically operated fluid powered servopositioning apparatus having three channels A, B and C and arranged to drive a common load member shown as a shaft. 1
FIG. 12 is a sectional view of a hydraulic type hysteretic equalizer shown schematically in FIG. 8.
FIG. 13 isa diagram of the relationship of piston displacement to differential pressure for the two pistons shown in the hysteretic equalizer illustrated in FIG. 12 g and indicating the differential pressure breakout range for such pistons.
FIG. 14 is a diagram similar to FIG. 13 but also indicating the effect of the lost motion drive included in the mechanization depicted in FIG. 12.
FIG. 15 is a diagram similar to FIG. 14 but additionally showing the hysteresis' effect of friction means included in the mechanization depicted in FIG. 12.
FIG. 16 is a schematic of an electrical circuitfor providing a hysteretic equalization feedback signal including three components labeled as pressure transducer circuit, hysteresis circuit and threshold circuit, this overall circuit being an embodiment of the present invention in electrical means, there being a diagram indicating the relationship between input and output opposite each component circuit all of which collectively produce the composite diagram at the bottom of the figure indicating the relationship between equalization signal and load differential pressure.
It is very important in certain situations to provide electrical input-hydraulic output systems which are very reliable since life may depend upon their proper operation or they may be part of a vehicle such as aircraft or spacecraft which is .very expensive. Generally speaking, there are two ways of improving reliability. One is to make the elements stronger by overdesigning them. This is not feasible where the parts are inherently small and delicate, as here. In this latter situation the alternative is to use more than one device in the system so as to provide a redundant system in order to achieve reliability.
FIGS. 1-7 will be referred to in connection with the ensuing description to indicate the background for the present invention later to be discussed herein.
When two or more channels of a redundant electrohydraulic servoactuator are force summed, as with a tandem cylinder/piston or with torque summing shaft, the problem of synchronizing the channels arises. Without synchronization the channels may fight each other and so not work together to drive the load. Of even more concern, system deadzone may occur. Deadzone is that length of actuator travel where it can move in either direction since there is no net drive on the load. The problem of synchronization is especially acute when closed-center sliding-spool valves are used to drive each cylinder/piston. These valves have a very high pressure gain around null so that large differential pressures or high piston forces are created with very small spool displacement.
Consider the case of two servovalves and actuators in a position servosystem driving the same load to a force summed arrangement. A load drive force will be built up by each servovalve whenever a net position error exists. Net position error is the difference between the actuator position and command position, together with equivalent position offsets due to valve and amplifier null shifts and other disturbances. In effect, when the servovalve receives a command to move off null, it will create a differential pressure that attempts to move the actuator. The magnitude of this differential pressure is related to the magnitude of the position error.
The overall actuator pressure gain that is present with two force-summed servovalves when null shift is present is depicted in FIG. 1. The region of ,zero force gradient will occur whenever the actuator. stops. This region will cause actuator deadzone as a corresponding actuator position error must be developed whenever the actuator is to be moved.
With an odd number of servovalves and actuators operating simultaneously in a force summed arrangement, as depicted in FIG. 2, there is always a stiff region about null wherever the actuator is positioned. The staircase of actuator force gain would only be apparent under load.
Several design techniques are available to aid in servovalve synchronization. If the valves are physically just different sections of a tandem valve spool, the synchronization can be achieved by careful null cutting or grinding of the various flow metering lands. If the valves are physically separate but rigidly linked together, as in a rip-stop design, them it is possible to adjust or shim the valves for shychronization. Usually the mechanical valve synchronization techniques are practical for only the two-valve case.
Another approach taken by the prior art is to soften the pressure gain characteristic of the servovalves by overcutting the nulls to create an open-center valve configuration, or by using proportional pressure feedback servovalves. If nozzle or jet type single-stage servovalves are used, the pressure gain is inherently low due to the open-center configuration.
If the valve pressure gain is sufficiently reduced so that it remains relatively constant throughout the range wherein valve mismatch will occur, then load sharing is accomplished as depicted in FIG. 3. However, the actual pressure gain that results from either the opencenter or proportional pressure feedback technique is significantly lower than that which is achieved with closed-center servovalves. This gives correspondingly poorer actuator resolution capability.
When separate servovalves are used in multi-channel redundant servovalve actuation system, it is generally necessary to use equalization. Equalization involves the use of additional feedback loops to create signals that reduce interchannel mismatch.
Equalization is usually required for two reasons. One is to reduce the transient that may result when shuttingoff a failed channel, and the other is to achieve better synchronization or load sharing between various channels. Reduction of the shut-off transient involves the concept of slowly changing from the system null that existed prior to failure and shut-off, to that which exists after shut-off.
A number of different equalization schemes have been devised in the past with varying degrees of success. The limitations of a specific approach are usually not immediately apparent, and only appear after careful scrutiny of such characteristics as (a) performance with differing supply pressures, (b) dynamic stability of equalization loops, (0) reset provisions if integrators are used, (d) need for interchannel comparisons or connections, (e) if there are interchannel comparators, the need for redundancy of comparators and power supplies, and (f) the suitability for flightworthy implementation of the equalization concept. These characteristics will be brought out in the following discussion.
All equalization concepts or arrangements require a means for measuring interchannel mismatch. With force summed channels, this measurement may be ential pressure detent, then a signal is developed that can be used for equalization. I
Another concept or arrangement is similar butuses position detents that are displaced whenever interchannel mismatch occurs. These detents create electrical signals that can be used for equalization.
The range of the equalization feedback signals must be sufficient to offset the total anticipated interchannel mismatch, including input signal mismatch, null and tracking differences in position feedback transducers, servovalve null shifts and servoamplifier mismatch. These may be on the order of i 5 percent command input mismatch, i l percent-position feedback mismatch, :L 1 percent due to servovalve null shifts (the normal servovalve null uncertainties of 6'to i 10 percent are reduced by the gain of the electrical servoamplifier) and i 1 percent due to the servoamplifiers. Therefore the usual equalization range approaches 1t 10 percent of full command input. Equalization signals that exceed a preset limit may be usedto indicate channel failure.
The function of equalization feedback in synchronizing interchannel mismatch can be described by considering the two-channel case. Previously the deadzone problem was described for two-channel operation with null mismatch, that is, lack of synchronization. The deadzone resulted from zero force gradient in the re gion of force polarity reversal. This deadzone problem can be understood by the ability or inability of the actuator to move back and forth across a load that requires a force reversal, such as a detent or breakout load as represented in FIG. 4.
The actuator force scheme, without equalization, may be as depicted in FIG. 5. There, to breakout the load on either side of actuator null requires a channel force as developed at error position EP or EP in FIG. 5. The separation of these points will then correspond to the actuator deadzone, as shown in FIG. 6. The width of the deadzone will depend upon the actual null shift between the two channels.
Now with equalization present, an equalization feedback signal will be developed at one or both channels, depending upon the load present and the equalization breakout levels. For example, with equalization breakout levels as shown in FIG. 7, and driving the load corresponding to position EP, in FIG. 5, channel No. 2 will receive an equalization signal that draws its null toward channel No. 1. For such situation, the equalization signal has reduced the null separation between channels No. l and No. 2 by feeding back to channel No. 2 a signal related to the null mismatch. However, this equalization signal may or may not be effective depending upon the holding capability of the equalization feedback. If there is no holding ability in the equalization, then the equalization feedback will shift from channel No. 2 to channel No. l, and back, as the actuator is cycled across the detent load. It is apparent that this would not overcome the actuator deadzone problem.
According to the prior art, one way of providing the holding ability of the equalization feedback was by integrators, either electrical or hydromechanical. Ideally the equalization integrators would hold the correct values of feedback signals sufficient to offset individual channel null shifts. However, the use of pure integrators, unfortunately, will not work without some means of continuous integrator reset.
A hypothetical sequence that illustrates the need for integrator reset is the following. Assume one channel of a triple channel actuator having pure integral equalization 'is arbitrarily commanded percent. Its equalizer will introduce a 10 percent signal to offset the command, as the single command input appears as a null shift, or interchannel mismatch. Next, repeat the sequence for the second channel, and then the third. A situation at this point is that all commands are at 10 percent and yet the actuator has not moved. In practice this dilemma is manifested by saturation of the equalization integrators shortly after turning on the system.
While several techniques have been used to prevent saturation of the equalization integrators, most of these have involved some form of cross-channel comparison, together with integrator reset on a preferential basis. Such an arrangement is tremendously complex and has several undesirable aspects including the requirement of connections between the interchannel comparators and this comprises the separateness of the channels and may lead to mutual failure modes; also the crossconnections involve multitude of wires; further the reset logic must be changed for systems with multiple failure capability depending on how many and which channels are active; and also comparison elements and their power supplies may need to be redundant to achieve a reasonable failure probability.
Because of these limitations, other means for reset of equalization integrators is usually desirable. Such other means involves the use of a simple laggedintegrator with a long time constant. The time constant of the reset circuit must be sufficiently long such that the equalization signal does not follow the lowest frequency of command input. With quasi-static input signals that approach the corner frequency of the equalizer, the actuator position will fail to follow the command as the equalizer will continuously wash out the command signal. This is particularly troublesome with underdamped long-period vehicle resonant modes, such as the phugoid oscillation of an airplane or rocket. It is generally impractical to achieve sufficiently long time constants for lagged equalization integrators in a flightweight system.
FIGS. 8-15 The aforementioned deficiencies and disadvantage of the prior art equalization techniques have been overcome by the principle of the present invention which briefly may be characterized as providing a hysteretic equalization feedback signal. The present invention avoids the limitations of cross-channel comparators for equalization reset, and does not suffer the frequency sensitive problems of lagged integrating equalizers.
Before describing the inventive hysteretic equalizer means in detail, a sequence of problems that led to the need for such an equalizer can be reviewed as follows. First, the need for equalization in multi-channel operation in a majority voting or mid-value logic system arises to reduce the transient following shut-off of a failed channel. There is also the need for finite loaddrive pressure gain in the region around actuator null. This only becomes a problem when an even number of channels are active as, with an even number, any channel null shift'leads to a region of zero pressure gradient about null. This condition of zero pressure gradient will give actuator position deadzone. The techniques of open-center valves or proportional pressure feedback valves to soften the pressure gain characteristics and so avoid deadzone have been discarded in favor of a technique for equalizing the channel nulls.
Accordingly, an approach for equalization pursuant to the present inventive concept involves providing a means for measuring interchannel mismatch by mechanical or hydraulic detents. The magnitude of mismatch is fed back through a holding-type of equalization that provides a signal to reduce the interchannel mismatch. Reset of equalization holding circuits by interchannel comparisons has been discarded in favor of a self-contained channel reset approach. This led to the consideration of lagged equalization integrators and the frequency bandwidth problem associated with the reset time constant.
Referring to FIG. 8, let us consider an equalization sensor which has a hysteresis characteristic together with a threshold characteristic. With only two channels, the force or differential pressure output of one channel equals that of the other, providing the load is zero. This situation is indicated by points lES and 2ES in FIG. 8 where the right side of the diagram represents the equalization of one channel, say channel No. l, and the left side indicates the equalization in the other, say channel No. 2. The sum of the two equalization signals represents the total null mismatch. These equalization signals will remain fixed until the actuator is required to move a load.
To overcome a load, the actuator must develop a net differential pressure of one polarity or the other. Assume this polarity is positive in FIG. 8. The change in actuator equalization signals will then be as shown in FIG. 9. There, channel No. 2 contains a constant equalization signal, as represented by point 2ES while its servovalve develops a differential pressure essentially equivalent to driving the load. Channel No. 1 can develop very little positive differential pressure as its equalization signal, indicated at point lES in FIG. 9, effectively offsets any error signal tending to increase the pressure.
Reversal of the direction of load drive causes the opposite effect as represented by the points 1ES and 2ES in FIG. 10. The result is that a net equalization signal is obtained in each channel that establishes an operating point for the differential pressures which allow both valves to work together to drive the load.
It is significant to note that the total actuator pressure gain in the region around these valve operating points is effectively double that of just one valve operating alone. Ideal load synchronization is, therefore, present.
The magnitude of the equalization signals must be sufficient to offset the anticipated interchannel null mismatch. Excessive equalization signal may be used as a criterion of channel failure. The maximumdifferential pressure available within the equalization operating range must be sufficient to drive the load for the lowest number of channels active. The width of the hysteresis range should be as wide as practical without losing the region about null wherein zero differential pressure always corresponds to zero equalization signal. If this condition is not retained, then a problem equivalent to integrator reset of the non-hysteretic equalizer arises.
Turning now specifically to FIG. 11, a redundant electrically operated fluid powered servopositioning apparatus having at least three identical channels is depicted therein. These channels are designated, reading from left to right, as channel A, channel B, and channel C.
Each channel includes certain elements hereinafter described in detail in connection with channel A as illustrative of all channels, the similar elements for channel B being designated with the suffix b, and the similar ones for channel C being designated with the suffic c.
Channel A shows a command signal represented by line 20 being fed to signal summing means including a first summing point 21 connected by a line 22 to a second summing point 23, the output of this second summing point being transmitted via line 24 to an amplifier 25. The output of this amplifier is transmitted via line 26 to a flow control servovalve 28. This servovalve may be of any suitable construction and is intended as being of the electrohydraulic type such as shown in US. Pat. No. 3,023,782. This servovalve 28 has two actuating or control ports, one of which via conduit 29 leads to the left end of the cylinder 30 of an actuator 31, the other port being connected via conduit 32 to the right hand end of cylinder 30. In actuator cylinder 30 is shown a slidable piston 33 connected by a right piston rod 34 to the free or outer end of a lever arm 35 fast to a shaft represented by the broken lines 36. One end of this shaft is shown as being fast to another lever arm 38 suitably connected to a load such as a control surface of an aircraft. The other side of the actuator piston is shown as connected by a left piston rod 39 to a position feedback device indicated generally at 40. This device may be of any suitable construction, that illustrated being shown schematically as a linear variable differential transformer adapted to" generate an electrical signal proportional to displacement of piston rod 39, such signal being transmitted via line 41 to summing point 21.
In accordance with the present invention a hysteretic equalizer indicated generally at 42 is provided for generating and feeding back an equalization signal responsive to the output of servovalve 28. Accordingly, hysteresis equalizer 42 is shown as having a pair of conduits 43 and 44 connected to conduits 29 and 32, respectively. For present purposes, let it be assumed that hysteretic equalizer 42 moves a member 45 of a feedback device 46 adapted to generate an electrical signal transmitted via line 48 to summing point 23.
By reason of actuator positionfeedback signal returning via line 41 to summing point 21 to which the command signal is fed, there leaves this summing point the normal servovalve position error signal which is transmitted via line 22 as an input to the second summing point 23 to which the hysteretic equalization signal is fed back via line 48. As an output of this second summing pint 23, an equalized error signal is transmitted via line 24 to amplifier 25 from whence the amplified signal passes via line 26 as a valve signal to the torque motor of servovalve 28.
It will be noted that lever arms 35b and 35c connected to the corresponding right actuator piston rod are fast to shaft 36 so that this shaft represents a common load member for the various channels. It will further be noted that there are no interchannel connections shown in FIG. 1 1 between the various channels A, B and C. The only connection is through the common load member 36.
A hydromechanical form of hysteretic equalizer 42 is shown in detail in FIG. 12. This equalizer is there shown as comprising a cup-shaped body 50 having a stepped recess including a relatively deep and narrow inner cylindrical recess portion 51 and an outer cylindrical portion 52 connected thereto by an outwardly facing shoulder 53, the outer extremity of recess portion 52 being internally threaded as indicated at 54. A closure 55 having an externally threaded neck is screwed into the threads 54 and closes off the recess in body 50, thereby to provide a compartment defined in part at the right end by an end wall 56 and defined adjacent its other or left end by an end face 58 formed on the inner end of the extension of closure 55.
A first or right piston 59 is slidably arranged in recess portion 51. A second or left piston 60 is slidably arranged in recess portion 52. Operatively interposed between these pistons is a helical compression spring 61 which in the embodiment being described is in a precompressed condition so that pistons 59 and 60 are biased laterally outwardly or preloaded, the right end face 62 of right piston 59 bearing against end wall 56 and the left end face 63 of left piston 60 bearing against the surface 58.
Closure 55 is shown as being recessed from end face 58 so as to provide a chamber 64. This chamber is communicatively connected to a radial passage 65 which extends outwardly through body 50 into communication with conduit 43.
The right end portion of right piston 59 is shown as being reduced so that it in cooperation with recess portion 51 provides a chamber 66 which is communicatively connected via radial passage 68 in body 50 to conduit 44.
Left piston 60 is shown as having a reduced inner portion 69 which in cooperation with recess portion 52 and shoulder 53 provides a chamber 70. This chamber is connected with passage 68 via a passage 71 provided longitudinally in body 50.
Piston 60 is shown as having a central horizontal longitudinally extending through bore 72 which at its left end can communicate with chamber 64 although this is assured by reason of an angled passage 73, and at its right end communicates with a chamber 74 existing between pistons 59 and 60.
In this manner, it will be seen that a differential pressure between conduits 43 and 44 is applied to piston 59 to urge it to move in one direction, and the same differential pressure is applied to piston 60 to urge it to move in the same direction. The passage arrangement is such that the differential pressure is applied to each of pistons 59 and 60 so as to urge one to move toward its stop surface and the other to move away from its stop surface.
In FIG. 12, the device 46 is shown specifically as a linear variable differential transformer including a coil 75 and a movable member or probe which is designated 45, this being the movable member illustrated in FIG. 11.
Lost motion drive means are operatively interposed between pistons 59 and 60 and probe 45. Such means are shown as comprising a rod 76 connected suitably at its left end to probe 45 and extending through an opening 78 provided in the cover portion of closure 55, this rod continuing to extend through bore 72 in piston 60, past its inner end face 67' and past the opposing end face 77 of piston 59 into a recess 79 having an outer enlarged portion 80 leading to end face 77 and an inner enlarged portion 81.
Three annular collars or lugs 82, 83 and 84 are shown as fast to rod 76 at spaced intervals therealong. Lug 84 is shown as being fast to the extreme right end of rod 76 and in the position of the piston 59 illustrated in FIG. 12 the left end face of this lug engages the shoulder 85 between recess portions 79 and 81 in this piston. Left lug 82 is shown as having its right end face engaging end face 63 on piston 60. Intermediate lug 83 is shown as having its left end face 86 opposing end face 67 on left piston 60 and spaced therefrom a distance equal to that which exists between the right end face 88 of this lug and the opposing shoulder 89 which exists between recess portions 79 and 80 in right piston 59. In other words, the spacing between the various lugs 82-84 is such that intermediate lug 83 is spaced substantially equally from piston end faces 67 and 89 when piston end face 63 engages left'lug 82 and piston end face 85 engages right lug 84, this occuring when the pistons 59 and 60 engage their respective stop surfaces 56 and 58, respectively, as shown. The diameters of lugs 83 and 84 are sufficiently less than those of the wall surfaces of the corresponding recess portions 80 and 81 such that the fluid pressure is equal on opposite sides of lugs 83 and 84 at all times.
Friction means are provided for resisting yieldingly movement of the probe 45 or rod 76. In the embodiment shown in FIG. 12, such means are indicated generally at 90 and while capable of being constructed in any suitable manner are shown as comprising a body of suitable packing material 91 such as asbestos or the like arranged between two glands 92 and .93, this assembly of elements 91-93 being arranged in passage 78 in closure 55 and severally surrounding rod 76. The packing 91 holds rod 76 against movement while lost motion in the aforementioned drive means is being taken up and thereafter wipingly engages the periphery of this rod as it moves in or out of body 50.
The effect of the various elements of the hysteretic equalizer shown in FIG. 12 are depicted in FIGS. 13-15. A differential pressure breakout range is indicated in FIG. 13, that for piston 59 being the distance along the abscissa from point 0 to point B59 and that for piston 60 being the distance along the abscissa from point 0 to point B60.
Since each of pistons 59 and 60 contains a lost motion drive to the probe 45 of the electrical signal generating device 46, this is represented in FIG. 14 for piston 59 by the distance along the abscissa from point B59 to point L59 and for piston 60 by the distance along the abscissa from point B60 to point L60, and gives a wider deadzone for the probe position. The friction device 90 cooperates with the probe 45, 76 to hold the probe stationary while a piston moves to take up the lost motion.
Thereafter movement of piston 59 will move probe 45, 76 from point L59 to point P59, and movement of piston 60 will move this probe from point L60 to point P60, as depicted in FIG. 15.
The friction device 90 cooperates with the probe 45, 76 so that the probe will remain stationary when either piston thereafter backs off due to a decrease in differential pressure, as depicted in FIG. 15, for piston 59 by the horizontal distance from point P59 to point F59 and for piston 60 by the horizontal distance from point 4 P60 to point F60.
With these diagrams shown in FIGS. 13-15 in mind, the hysteretic equalizer 42 as shown in FIG. 12 operates in the following manner.
When there is no differential pressure between conduits 43 and 44, the various elements will be in the condition depicted in FIG. 12. Assume now that a differential pressure between conduits 43 and 44 is developed as a result of a valve signal to servovalve 28 such that the pressure in conduit 43 is higher than that in conduit 44. Such pressure differential applied to right piston 59 merely assures its extreme rightward position depicted wherein its right end face 62 engages end wall 56. On
left piston builds up to a level overcoming the preload thereon, indicated from point 0 to point B60, whereafter a further increase in differential pressure displaces piston 60 away from stop surface 58 against the urging of spring 61. Such rightward displacement of left piston 60 continues until its right end face 67 engages left end face 86 on intermediate lug 83, a total distance indicated from point B60 to point L60, whereupon this lug is moved along with the piston in a rightward direction. This shifts rod 76 and probe 45 in a rightward direction, causing the generation of an electrical signal by device 46 which is proportional to the displacement of this probe. This continues as depicted by the line from point L60 to point P60 until maximum differential pressure intended is reached, whereupon a reduction in this differential pressure will cause a leftward movement of piston 60 while the friction device the other hand, this differential pressure as applied to 90 holds rod 76 stationary, depicted as movement from point P60 to point F60, the flat top line on the curve shown in FIG. 15.
It will be noted that when piston 60 left its stop surface 59 and moved initially toward the end face 86 of intermediate lug 83, this piston moved away from left lug 82. Now it is returning toward engagement with this lug 82 while moving away from engagement with lug 83. This leftward movement of piston 60 continues until its end face 63 engages lug 82, terminating at point F60. When such engagement occurs, continued leftward movement of piston 60 will push lug 82 with it and thereby shift rod 76 and probe 45 leftwardly, thereby causing a change in the signal generated by device 46, depicted by the line from point F60 to point B60. This change in signal is proportionate to the unloading of now expanding spring 61 from its more fully compressed condition to its preloaded condition. This portion of the return travel of piston 60 continues until it engages stop surface 58. From this point on, any reduction in differential pressure is without the generation of any equalization signal, as represented by the horizontal line between point B60 and point in FIG. 15.
On the other hand, let us assume that the differential pressure between conduits 43 and 44 is such that the dominant pressure is in conduit 44 resulting in a change of polarity. This differential pressure across left piston 60 assures its retention against its stop surface 58. But this differential pressure applied to right piston 59 causes the same to build up to offset the preloading on this piston by spring 61, being represented by the horizontal line between point 0 and point B59 in FIG. 15. Thereafter an increase in this differential pressure of the same polarity will cause a yielding of spring 61 to allow piston 59 to move in the leftward direction until the lost motion or the spacing between surfaces 88 and 89 is taken up. This is represented by the horizontal line from point B59 to point L59 in FIG. 15. Thereafter an increase in differential pressure of the same polarity will continue to move piston 59 in the leftward direction, but in doing so it will push intermediate lug 83 and also shift rod 76 and probe 45 in a leftward direction. This causes generation of an electrical signal by device 46 which is proportionate to movement of probe 45, this electrical signal being represented by the line drawn from point L59 to point P59 in FIG. 15. This leftward movement continues until the maximum differential pressure intended has been achieved, represented at point P59. Thereafter, when the differential pressure reduces, the friction device 90 holds rod 76 and probe 45 stationary so that no change in the electrical signal generated is effected as the differential pressure subsides, represented in FIG. by the horizontal line between point P59 and point F59. This reduction in differential pressure allows the piston 59 to move rightwardly and back off away from engagement with intermediate lug 83. Such backing off continues until the end face or right lug 84 engages shoulder 85 on piston 59, occurring at point F59, whereupon continued movement of the piston picks up the rod 76 and shifts the probe 45 rightwardly causing the electrical signal to reduce along the line from point F59 to point B59 depicted in FIG. 15. This continues until piston 59 engages its stop surface-56, depicted at point B59. Thereafter the continued reduction and eventual elimination of the differential pressure is effective only to reassert the preload of the spring 61 on this piston as represented by the horizontal line from point B59 to point 0 in FIG. 15.
From the foregoing, it will be seen that means are provided for equalizing the outputs of the three servovalves, including, for each channel, equalizer means having a hysteresis characteristic for generating and feeding back to the signal summing means of a given channel a hysteretic equalization signal responsive to the output of the servovalve of such given channel.
While in the embodiment shown in FIG. 11 the hysteretic equalizer is shown as being responsive to the differential pressure output of the servovalve or the load actuator, it will be understood that another output of the servovalve such as the actuator force produced by its output or the flow of its output might be utilized as the output to which the hysteretic equalizer is responsive. In the case of the hysteretic equalizer being responsive to the actuator force output, a force transducer such as a strain gauge interposed in the mechanical load drive may be employed.
Instead of using electrical feedback for the actuator position or the equalization signal, an arrangement'can be provided whereby these one or more signals severally may be a mechanical force or displacement applied directly to a suitable element of the servovalve.
Another form of means for producing a hysteretic equalization signal is depicted in FIG. 16. Whereas the hysteretic equalizer 42 shown in FIG. 11 was of a hydromechanical type, that shown in FIG. 16 is of an electrical type. The hysteretic equalizer shown in FIG. 16 includes three main circuit components indicated generally at 95, 96 and 97. Component is provided by a circuit including a pressure transducer of any suitable conventional construction to convert a differential pressure such as detected from conduits 43 and 44 into an electrical signal proportional to the actuator differential pressure, represented by a for output to the left of pressure transducer 95. This pressure transducer produces a relationship between differential pressure and output or electrical signal as indicated by the diagram immediately below the pressure transducer in FIG. 16.
The central circuit component 96 receives the electrical signal from component 95 and passes it through a combination of computer analog elements that create a hysteresis characteristic depicted by the diagram in FIG. 16 below this component 96, input being represented by i and output by 0 in this diagram. These elements are shown as including amplifiers 98 and 99 and an integrator 100, arranged in a circuit possessing a hysteresis effect well known to those skilled in the art.
Circuit component 97 comprises amplifiers 102 and 103 and a biased diode circuit 104 to provide a threshold function generator circuit, also understood by those skilled in the art. A diagram indicating the relationship between input i and output 0 is for circuit component 97 illustrated below this circuit.
By cascading these three circuit components, a pressure transducer circuit 95, a hysteresis circuit 96 and a threshold function generator circuit 95, a hysteretic equalization signal e is created which is transmitted by line 48' to summing point 23 for the servoamplifier 25'. A composite diagram indicating the relationship of hysteretic equalization signal e to load differential pressure AP is represented in the lowermost diagram in FIG. 16, it being similar in configuration to that shown in FIG. 15.
What is claimed is:
1. In redundant electrically operated fluid powered servopositioning apparatus having at least two channels each including a servovalve, signal summing means to which a command input signal is fed, an actuator operatively associated with the output of said servovalve and arranged to drive a common load member and feedback means arranged to provide a load position feedback signal to said signal summing means, the output of said summing means being fed to said servovalve, the improvement which comprises means for equalizing the outputs of said servovalves including, for each channel, equalizer means having a hysteresis characteristic for generating and feeding back to said signal summing means of a given channel an equalization signal responsive to the output of said servovalve of such given channel.
2. Apparatus according to claim 1 wherein said output of said servovalve is a'differential pressure.
3. Apparatus according to claim 2 wherein said equalizer means comprises a piston, means for applying said differential pressure to said piston, means for pre-. loading said piston against movement induced by said differential pressure and for resisting increasingly but yielding such movement above such preloading effect, means including a movable member for generating an output electrical signal in response to movement of said member, friction means resisting yieldingly movement of said member, and lost motion drive means operatively interposed between said piston and member, said output electrical signal being said equalization signal.
4. Apparatus according to claim 2 wherein said equalizer means comprises a pair of pistons, means for applying said differential pressure to each of said pistons, common spring means operatively interposed between said pistons, means including a movable member for generating an output electrical signal in response to movement of said member, friction means resisting yieldingly movement of said member, and lost motion drive means operatively interposed between said pistons and member, said output electrical signal being said equalization signal.
5. Apparatus according to claim 2 wherein said equalizer means comprises a pair of pistons, common spring means operatively interposed between said pistons to bias them severally in opposite directions,
means limiting movement of said pistons in said opposite directions, means for applying said differential pressure to each of said pistons to urge one to move in one of said directions and the other to move in the other of said directions, means including a movable member for generating an output electrical signal in response to movement of said member, friction means resisting yieldingly movement of said member, and lost motion drive means operatively interposed between said pistons and member, said output electrical signal being said equalization signal.
6. Apparatus according to claim 2 wherein said equalizer means comprises a body having piston stop surfaces defining partly a compartment, a pair of pistons severally slidably arranged in said compartment, common spring means constantly urging said pistons severally toward said stop surfaces, means for applying said differential pressure of said servovalve to each of said pistons so as to urge one to move toward its stop surface and the other to move away from its stop surface, linear variable differential transformer means including a probe, friction means resisting yieldingly movement of said probe, and lost motion drive means operatively interposed between said pistons and probe, the output electrical signal of said linear variable differential transformer means being said equalization signal.
7. Apparatus according to claim 6 wherein one of said pistons has oppositely facing first and second surfaces, the other of said pistons has oppositely facing third and fourth surfaces, and said lost motion drive means comprises passage means in said one of said pistons connecting said first and second surfaces, passage means in said other of said pistons connecting said third and fourth surfaces, a rod connected to said probe and extending through said passage means, a first lug fast to said rod opposite and adapted to engage said first surface, a second lug fast to said rod intermediate and adapted to engage either of said second and third surfaces and a third lug fast to said rod opposite and adapted to engage said fourth surface, the spacing between said lugs being such that said second lug is spaced substantially equally from said second and third surfaces when said first surface engages said first lug and said fourth surface engages said third lug.
8. Apparatus according to claim 6 wherein said friction means comprises a packing surrounding and wipingly engaging said probe.
9. Apparatus according to claim 7 wherein said friction means comprises a packing surrounding and wipingly engaging said rod. I
10. Apparatus according to claim 2 wherein said equalizer means comprises pressure transducer means for providing a first output electrical signal related to said differential pressure, and electrical circuit means receiving said first output electrical signal as input and providing a second output electrical signal having a hysteresis characteristic, said second output electrical signal being said equalization signal.
11. Apparatus according to claim 10 wherein said electrical circuit means includes hysteresis circuit means and threshold circuit means.
12. Apparatus according to claim 11 wherein said hysteresis circuit means includes amplifier means, integrator means and biased diode means, and said threshold circuit means includes amplifier means and biased diode means.
* k k I!
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|U.S. Classification||91/363.00A, 91/512, 91/509, 91/433|
|International Classification||F15B18/00, F15B20/00|
|Cooperative Classification||F15B20/00, F15B18/00|
|European Classification||F15B20/00, F15B18/00|