Abstract:
An arrangement for assuring synchronous motion of a pair of actuators ( 65, 67 ) includes two gear racks ( 99, 101 ) moved by cables ( 95, 97 ) fixed to remote hydraulic actuators ( 65, 67 ) to sense the motion of the actuators ( 65, 67 ). The racks ( 99, 101 ) drive pinions ( 111, 113 ) and attached bevel gears ( 113, 115 ) at identical rates and an idler bevel gear ( 119 ) that spins in-place on its own axis when actuator travel is synchronous. When the actuator ( 65, 67 ) travels are synchronous, no error signal is sent to a synchronizer servo valve ( 105 ), however, when one actuator&#39;s speed is too great, an output error signal from the idler gear ( 119 ) is sent. The output signal positions a servo valve ( 105 ) to release pressure which is holding open one of two throttling valves ( 107, 109 ). Releasing the pressure allows the throttling valve ( 107, 109 ) to be closed by a spring ( 129, 131 ) and restrict flow to the actuator ( 65, 67 ) which is traveling the fastest. A logic valve ( 103 ) which has been previously positioned by the motive flow to or from the actuators assures closure of only the correct throttling valve ( 107, 109 ) so that the speed of only the faster of the actuators is reduced. Thus the slower actuator is always assured application of full motive flow and pressure so that it can drive the system at its full rated power.

Description:
The present invention relates generally to hydraulic or pneumatic control systems and more particularly to methods and apparatus for hydro-mechanically synchronizing the operation of independent actuators or motors regardless of variations in imposed loads, travel, or actuator efficiencies. In particular, the present invention provides synchronization for the hydraulically powered movement of doors associated with aircraft engines. 
     It is desirable to provide a hydraulic control which assures synchronous motion of a pair of actuators. The present invention creates a hydro-mechanical means to provide reliable synchronous operation of hydraulically or pneumatically powered linear or rotary actuators or motors. It may be applied to systems in which the actuators or motors have equal or unequal displacements, efficiencies, and/or imposed loads, and may even be applied where the strokes of the actuators from stop-to-stop differ or the actuators act in differing directions. Synchronous operation is assured by means of a closed loop, mechanical feedback system and a controller which operates on the hydraulic or pneumatic motive fluid. 
     Some systems may require the total stop-to-stop stroke of each actuator to be different, or the orientation of each actuator to differ, yet operation to be synchronized so that each remains in step with the other at any given point in its travel range or to reach the stops simultaneously. In this scenario, a ratio other than 1 to 1 between the feedbacks from the individual actuators may be built into the differential sensor device, to provide an appropriate output error signal to the servo valve. 
     Another advantage inherent in the flexibility offered by the proposed approach is the availability of a number of places in which system gains can be optimized, to maximize the effectiveness of the synchronizer, and to minimize the risk of erratic behavior or unstable transient operation. For instance, over-activity of the controller in response to small travel perturbations can be restrained by inclusion of deadband in the valving, yet quick reaction to significant synchronous operation can be assured by high gains in the throttling loops. 
     In certain aircraft applications, each inner door, upper or lower, is independently positioned by its own pair of hydraulically powered ram actuators. The actuators are located at either end of the door, each physically isolated from the other. A controller package manages operation of the actuators on each door. In one particular application, each inner door actuator is sized to withstand a tensile stall load of nearly 37,000 pounds and to slew against dynamic loads of 20,200 pounds. Each has a 12.5″ stroke. Extreme loads, long strokes, the large physical separation between these actuators, and the light engine/airframe structure make impractical the use of a mechanical means to assure synchronization. Such a system would simply be too heavy for practical aircraft use. 
     The present invention provides synchronization while obviating these problems in which direction of actuator travel is detected, differential rate of motion is determined, and correction to restore synchronous operation is established by valves which use the motive flow powering the devices being driven. The valving restricts the supply of motive flow to the faster of the powered devices to limit its rate of travel. The need for the application of large braking forces at the output of the powered devices to control rate of travel is eliminated, as are the inherent heat generation and the deleterious impact on system efficiencies of braking systems. The present invention is a combination comprising: a pair of bidirectional fluid powered actuators; a source of pressurized fluid for operating the actuators; means monitoring the motion of each actuator and providing an indication of a dissimilarity between the motions of the two actuators; and means responsive to a dissimilarity indication for retarding the flow of actuating fluid from the source to the appropriate one of the actuators to diminish the dissimilarity between the motions of the two actuators. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a simplified partially schematic illustration of one actuator synchronizing system according to the invention in one form; 
     FIG. 2 is a detailed schematic of a hydraulic control valve assembly illustrating the invention in another form; and 
     FIG. 3 is a view in cross-section of a modification to the direction detection and logic valve of FIG.  2 . 
    
    
     Corresponding reference characters indicate corresponding parts throughout the several views of the drawings. 
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     In FIG. 1, a pair of fluid powered reciprocable pistons  11  and  13  are energized by pressurized fluid from a hydraulic or pneumatic fluid source  15  to move objects or loads  17  and  19  from initial locations  21  and  23  to final positions  25  and  27 . The distance d 1  which object  17  travels is greater than the distance d 2  which object  19  travels. The motion of each object is monitored by a pair of cables, wire ropes, tapes or other inelastic strands  29  and  31  which are fixed at one end to the ram portion of the pistons  11  and  13  and extend by way of idler pulleys such as  31  and  33  to dissimilar radius drums  35  and  37 . The drums  35  and  37  are fixed together for common rotation about a common axis. Also, the drums are preferably spring biased in a counterclockwise direction as viewed by a spiral rewind spring or other rewind tensioning arrangement (not shown). With the objects  17  and  19  in their retracted positions as illustrated, when the fluid source is enabled, fluid flows through the valves  53 ,  52  and  51  to extend the pistons  11  and  13 . It may happen that the motion of the two objects is in perfect synchronism with piston  13  moving at a fraction k=d 2 /d 1  of the speed of piston  11 . Under these conditions, if the objects commence their motion at the same time, they will arrive at their destinations at the same time. That is, their motion is perfectly synchronized. If drum  35  has a radius r 1  and drum  37  has a radius r 2  and if r 1 /r 2 =d 1 /d 2 , the common rotation of the two drums will allow the cables  29  and  31  to be paid out at exactly the correct rates and the lateral forces on the two drums will be equal. 
     The drums  35  and  37  are fixed to a lever arm  39  which is pivotable at  41 . Lever arm  39  is biased toward the neutral position shown in solid lines when the lateral forces on the two drums are equal. The opposite end of lever arm  39  is attached by way of a link  43  to the piston  45  of valve  52 . If object  19  is moving too rapidly, so as to arrive at its destination  27  before object  17  arrives at its destination  25 , the cable  31  will pull the drums rightwardly as toward the positions shown in dotted lines and with component positions indicated by primed reference numerals. Lever arm  39  will pivot counterclockwise and the spool or piston  45  of valve  52  will move leftwardly. This leftward spool motion diminishes or completely shuts off the fluid flow through conduit  47 . Depending on the valve  52  configuration, this leftward spool motion may enhance the flow from conduit (line or fluid duct)  49 , or may leave that flow unaffected. In either case, fluid flow to the piston  13  is reduced slowing the upward progress of object  19 , as desired. Of course, the lever arm  39  pivots clockwise effecting a similar control when the object  17  is moving too rapidly assuring that the objects reach their destinations at the same time. 
     As will be the case in FIG. 2, the pistons  11  and  13  could be fluid powered to return to their initial positions, however, return springs  54  and  58  are illustrated in FIG.  1 . To return the objects to their initial positions, valves  51  and  53  are actuated to interchange the fluid drain paths. Thus, the fluid from the cylinder of piston  13  flows under the urging of spring  58  through conduits  57 ,  49 ,  59 , and  61  to the return  55 . Similarly, fluid from the cylinder of piston  11  passes through conduits  63 ,  47 ,  59 ,  61  and into the return  55 . Valve  51  provides a direction correction so that the proportioning valve  52  continues to properly correct for object motion imbalances. 
     Assume that object  17  is returning too rapidly, that is, it will arrive at the initial point  21  prior to object  19  arriving at point  23 . Under this assumption, cable  29  is returning to the drum  35  too rapidly and the arm  39  again pivots counterclockwise toward the dotted line position. This motion moves piston  45  leftward restricting the conduit  47  and slowing the egress of fluid from beneath piston  11  thereby slowing the downward movement of object  17  as desired. Had valve  51  not been actuated, fluid flow from the cylinder of piston  13  would have been restricted slowing the wrong object. 
     Referring now to the door control valve assembly illustrated in FIG. 2, bidirectional inner door actuators  65  and  67  receive extend fluid flow from lines  69  and  71 . Retract fluid flows to the actuators from lines  73  and  75 . High pressure hydraulic fluid is supplied by line  77  and line  79  is a low pressure return line. Valve  81  controls the flow of pressurized fluid through line  83  to extend a pair of outer door actuators and through line  85  to retract those actuators. Electrical connections to the assembly is by connectors such as  80  and an electrical failure solenoid is illustrated at  82 . The outer door  66 ,  68  actuators are not synchronized, however, the inner door actuators  65  and  67  are synchronized. 
     The actuator controller package for each door includes an inner door electro-hydraulic servo valve  87  to receive electrical signals from the engine electronic controller which provides the command to position the door, and two RVDT (“rotary variable differential transformer”) position sensors  89  and  91  for feedback of door position to the engine controller. Each actuator controller package also contains a synchronizer device  93  for the inner door actuators. The synchronizer  93  receives actuator position feedback via mechanical linkages such as Bowden cables  95  and  97  and racks  99  and  101 . A differential gear arrangement detects any out-of-synchronous travel between the two actuators and synchronous operation is restored by hydraulic means. 
     The synchronizer or synchronous differential error sensor  93  is a mechanical differential device driven by mechanical feedback linkage from the actuators. It reduces large excursions of the actuators to a small amplitude relative motion indicative of synchronicity error between the travels of the two actuators. It acts through logic  103 , servo  105 , and throttling  109  and  107  valves to assure that positional synchronization is maintained between the pair of actuators throughout their full travel range. This error sensor  93  is a differential gear motion sensing device and includes the pair of racks  99  and  101  each of which is coupled to and linearly movable with a corresponding one of the cables  95  and  97 . There are a pair of gears  111  and  113  each engaging a corresponding rack and a pair of bevel gears  115  and  117  each coupled for joint rotation with a corresponding gear  111  or  113 . A third bevel gear  119  simultaneously engages each of the pair of bevel gears  115  and  117  and revolves about its axis ultimately resulting in a retardation of the flow of actuating fluid from the source line  77  to the appropriate one of the actuators  65  or  67  only when the two cables  95  and  97  move by different amounts. In response to this out-of-synchronous actuator travel condition, the error sensor  93  positions the hydraulic synchronizer servo valve  105  in the controller. For example, if actuator  65  which is coupled to cable  95  is leading the other actuator in the extend direction, the spool in valve  105  moves leftwardly as viewed. The resultant signal from the servo valve  105  is routed through the direction detector and logic valve  103  to one of two spring loaded, pressure opened hydraulic synchronizer throttling valve  109 . 
     The logic valve  103  is positioned by a differential pressure signal from one of the connecting hydraulic lines between the inner door valve  87  and the throttling valves  107  and  109 . The pressure signal which positions the logic valve  103  is created by check valves  121  and  123  within this line. The directional check valve  123  operates (opens) when flow in the line is extending the actuators, the other check valve  121  opens when flow is retracting the actuators  65  and  67 . Check valve cracking pressures are preset to achieve the desired minimum force margin to move the logic valve. Force margins and robustness as well as logic and check valve reliability may be enhanced by the logic and check valve modification shown in FIG. 3 wherein, as pictured, the logic valve  103 ′ must move before pressure can reach the check valve  121 ′ which only then relieves and permits flow to the actuators. The logic valve  103  determines which of the two throttle valves  107  or  109  is to be activated, depending upon whether the actuators  65  and  67  are extending or retracting, to correct an asynchronous condition. The activated throttle valve  107  or  109  then restricts the flow of fluid to or from both rod and head ends of the leading actuator of the pair. Synchronicity is quickly re-established, the servo valve  105  signal is turned off (nulled) by the error sensor  93 , and the throttle valve  107  or  109  returns to the open position for unrestricted flow to the actuator. 
     When the engine is started, hydraulic pressure from the engine pumping system will act on the end of each throttling valve  107  or  109  by way of lines  77  and  125  or  127  and compress the valve springs  129  and  131 . This opens flow passages to the head and rod ends of each actuator. When the valve  87  responds to engine commands to route pressure to and from the actuator lines  133  or  135 , flow in one of the two lines is initially blocked by the directional flow check valves  121  or  123 . The resulting pressure drop acts on one end of the logic valve  103  to shuttle it to the other end of its stroke. One check valve remains closed to the flow but, when pressure in this line reaches the check valve cracking pressure, the other check valve opens to permit full flow to or from the actuators. In essence, the force causing the spool of valve  103  to transition is the pressure differential across the check valve. Somewhat enhanced operation of the direction sensing logic valve  103  may be achieved with the modification of FIG.  3 . 
     In FIG. 3, primed reference numerals identify components analogous to components having unprimed reference numerals in FIG.  2 . If the spool of valve  103 ′ is in its uppermost position opposite the position shown, and extension pressure is commanded by valve  87 , nearly full supply pressure will be applied by way of line  133 ′ to the face of the spool driving it downwardly. It is only after the spool has moved some distance that the pressure driving the spool is the pressure differential across the check valve  123 ′. The valve  103 ′ behaves similarly to commanded actuator retraction. The arrangement of FIG. 3 differs from FIG. 2 only in the location of the check valve connections to the valve  103 ′. This modification provides a strong initial impetus to the spool and a more positive or robust response by valve  103 . 
     Returning to FIG. 2, the rate of travel of the actuators is proportional to command current to the inner door valve  87 . Given identical sizing, loading, efficiencies and seal drag, the two actuators  65  and  67  will proceed toward their final positions at equivalent rates. The feedback to the controller error sensor  93  will be the same from each actuator and relative motion in all elements of the sensor is such that no position discrepancy error output is indicated. Without a load equalizing member as part of the inner door structure, however, it is unlikely that the travel rate will be identical between the actuators at the two ends of the door. The valve  87  supplies the same flow and pressure to each of the two actuators. If unequally loaded, travel rates will differ slightly and the positions of the two actuators at any point in time will tend to be out-of-sync. Whenever the actuators are out-of-sync, feedback will cause the controller synchronous differential error sensor  93  to be offset by an amount proportional to the discrepancy. As shown schematically, one rack  99  or  101  will be traveling faster and will drive its pinion  111  or  113  and attached bevel gear  115  or  117  faster than the other does. The idler bevel gear  119  will try to spin at the faster rate but is constrained by the slower rack and pinion gear set. To stay in mesh, the idler bevel gear  119  will then impart a force normal to its axis that will cause the idler gear carrier and carrier axle or rotating spider shaft  137  to rotate on the centerline between the outer bevel gears. This allows the idler  119  to advance between the outer bevel gears at a rate equal to one-half the difference in speeds of the racks and outer bevel gears, in proportion to the amount of synchronous error between the actuators. Rotation of the carrier axle  137  drives torsional movement of the synchronizer servo valve lever  139 . Movement of the lever  139  offsets the synchronizer servo valve  105  from the null position to open a port which vents the hydraulic pressure from either line  125  or  127  that has been holding the throttling valves  107 ,  109 , open. 
     Either actuator could be the faster (leading) actuator of the pair, and the actuators could be traveling toward either extend or retract at the time the asynchronous condition is detected. The error sensor  93  alone has insufficient intelligence to determine which actuator is leading in which direction of travel. The logic valve  103  provides this function and permits venting of hold-open pressure only from the end of the throttling valve associated with the leading actuator, through the servo valve  105 , to return line  79 . The logic valve  103  determines which of the two throttle valves  107  or  109  is to be activated, depending upon whether the actuators are extending or retracting, to correct an asynchronous condition. The activated throttle valve then restricts the flow of fluid to or from both rod and head ends of the leading actuator of the pair. 
     An orifice  141  or  143  upstream of each throttle valve in the signal line  125  or  127  restricts the supply of new fluid to the end of the activated throttle valve and pressure decays to a lower level. Below the pressure level at which throttle valve force balance is reached a spring  129  or  131  closes the throttle valve in proportion to the decrease in pressure. The valve begins to restrict both the rod and head end flows of the leading actuator and slows it till its rate and position match that of the lagging actuator. As this occurs the servo valve  105  is returned toward null by the error sensor  93 , restoring signal pressure at the throttling valve and moving it open again against its spring till the valve is again hard over against the open stop in the positions illustrated. 
     With synchronous operation restored, the error sensor  93  and servo valve  105  will have returned to null. If unequal load conditions persist, the system will remain slightly off null, throttling flow to the lead actuator and matching the pace and position of the two actuators. The logic valve remains in its last position because only actuator synchronicity has changed, not the direction of flow to the actuators or their direction of travel. 
     In FIG. 1, the direction of actuator motion was determined by valve  53 . To achieve proper feedback, valve  51  needed to be set to the proper (extend or retract) position. This could be achieved manually or by ganging the two valves  51  together for simultaneous reversal. In FIG. 2, the direction of actuator motion is determined by valve  87  and valve  103  automatically senses the fluid flow to the actuators and conditions itself appropriately. Reversal of the command to the valve  87  and thus flow to the actuators, shuttles the logic valve  103  to the opposite end of its stroke where it permits synchronous actuator movement to be monitored and maintained in this direction of travel. With the reversed travel, a leading actuator generates an opposite torque at the error sensor  93  and strokes the servo valve to the opposite side of null, to again throttle flow only to the leading actuator. 
     A variety of options is available for implementation of the invention depending upon packaging constraints and optimal performance to engine inner door subsystem requirements or other application. For instance, mechanical feedback may take the form of a linkage, an aircraft control cable set (push-pull, loop, or tensioned pull cables, or a steel tape system), or may even be a rotary system with internal actuator helical ball splines driving flexible shafting, to provide position feedback to the controller. The feedback mechanism may even be encased in the operating fluid cooling flow return or head end supply lines if desired, for cooling, ambient protection, lubrication, or simply packaging benefits. The error sensor within the controller can simply be a multi-grooved pulley arranged to move a pivoted lever in response to an synchronous actuator position as illustrated in FIG. 1, or it may be any of a variety of rack or gear driven planetary differential schemes, with an output proportional to the synchronous error to drive the servo system. The throttling valves can be located at the individual actuators, rather than in the controller package. And the controller package itself can be remotely located relative to the actuators. 
     These and other conceivable options make for a robust system approach to provide the greatest possible flexibility in the final design, not only for the controller and actuator packages, but also for the system to which it is applied.