US3021674A - Normal fuel control with acceleration override control for gas turbine engine - Google Patents

Description

Feb. 20, 1962 H. c. zElsLol-T 3,021,674

NORMAL FUEL CONTROL WITH ACCELERATION VERRIDE CONTROL FOR GAS TURBINE ENGINE Filed Jan. 3l, 1955 I 2804?#230 j Gov ma@ www ZKM/MQW TMA/ff 3,921,674 NRMAL FUEL CUNTRL WillH ACCEEERATEN VER; TEE CNTRL EUR GAS TURBNE ENGNE Harry C. Zeisloft, Rochester, NX., assignor to The Bendix Corporation, a corporation of Delaware Fiied Jan. 31, 1955, Ser. No. 485,226 iti @latinsa (Ci. 6ft- $9.23)

This invention relates to a fuel feeding device for combustion engines and more particularly to speed governing mechanism therefor; it has been found to be most particularly adapted for use with combustion engines of the gas turbine type.

t is common practice to utilize centrifugal governing mechanism of the mechanical iiy-Weight type to control a fuel metering or throttle valve in fuel control systems for gas turbine engines, the governing mechanism including a governor spring which is normally manually adjustable by the pilot to accelerate and decelerate the engine to a selected speed.

The most widely used type of governing mechanism for gas turbine engines has been the so-called proportional type, wherein any deviation in engine speed from an equilibrium setting or operating condition results in a fuel now correcting opening or closing movement of the throttle valve which is proportional to the speed deviation. One of the primary advantages of this type of governor as compared, for example, with an isochronous or non-speed deviating type, relates to the fact that the above noted proportional action results in engine operating stability, since tl e valve does not travel to a fully open or fully closed position whenever a slight deviation in engine speed occurs, with subsequent hunting for an equilibrium location.

Heretofore, proportional type governors for gas turbine engines have been Classifiable in two main groups, viz. the direct acting kind in which the fly weights and govrnor spring act directly on the throttle valve, and the servo actuated type in which the governing mechanism functions through a power boost travel type servo device to control the position of the valve. The former type is well illustrated in the Mock Patent No. 2,689,606, led December 13, i945, and the latter type in the copending application of Andrew A. Kuzmitz Serial No. 446,335, tiled Iuiy 28, 1954 (common assignee), now abandoned.

One of the disadvantages of both f the above mentioned classes of proportional type governors is that maximum throttle valve travel is limited by the amount of displacement which can be imparted thereto by the movement of the centrifugal weights, either directly, as in the Mock patent, supra, or through servo mechanism, as in the Kuzmitz application, supra. Y

My invention overcomes this disadvantage by providing a servo powered proportional type governor in which position of the throttle valve is almost wholly independent of centrifugal weight position. This has been accomplished by utilizing what may be termed a force type servo governor mechanism in which centrifugal Weight force alone is converted into valve travel, as distinguished from a travel type servo governor mechanism in which centrifugal weight travel, resulting from a change in force output, is converted into valve travel.

This invention is illustrated in FGURES l and 2 ofthe drawings in a fuel control device of the so-called threedimensional cam type, which type of control is fully disclosed in my copending application Serial No. 248,402, filed September 26, 1951. One of the problems encountered in the use of such controls relates to excessive wear, over a period of time, of the contoured three-dimensional acceleration cam which results from excessive spring loads and the like imposed on said cam by the cam follower during engine accelerations. My invention provides mechanism associated with the governor for relieving any such excessive load on the cam, thereby greatly increasing the useful life thereof.

One of the primary objects of this invention is to provide an improved engine speed governor wherein a speed selecting force output is directly converted into a fuel now controlling travel.

Another object of this invention is to provide a force type servo governor for engines which is not subject to certain inherent limitations heretofore existent in direct acting and servo actuated proportional type governors.

A further object of this invention is to provide cam load limiting means in a combustion engine fuel control device.

Additional objects and advantages of this invention will become apparent to those skilled in the art in view of the following description taken in conjunction with the drawings, wherein;

FlGURE l is a sectional schematic view of a gas turbine engine and a diagrammatically illustrated .fuel control system connected thereto;

FIGURE 2 is a sectional schematic view of an engine' governor and associated mechanism adapted to be used in the fuel control device of FIGURE 1, in accordance with the invention; and

FIGURE 3 is a curve chart, illustrating the operational characteristics of the governor shown in FIGURE 2.

Referring now to FGURE l, a gas turbine engine is generally indicated at iti and includes a series of annularly disposed combustion chambers ll mounted in a casing having a header or air intake section l2, and a compressor i3, shown as of the axial flow type, whichis driven by means of a turbine 14 through a shaft 15. Each of the combustion chambers is provided witha burner nozzle 16 to which metered fuel is supplied under pressure by way of a conduit 17, fuel manifold 18 and individual fuel lines 19. rfhe conduit i7 receives metered fuel from a fuel control device, generally indicated at 2t) in FIGURE l, which includes the applicants governing mechanism as shown in FlGURE 2. A pump 22 supplies fuel under pressure to fuel control 20 through a conduit 24, a portion of which fuel may be by-passed back to the pump inlet through a conduit 26.

The fuel control 20 contains mechanism adapted to respond to compressor inlet temperature (Ta), as sensed by a temperature bulb 2S, compressor inlet pressure (Pa), picked up by a impact pressure pick up tube 30, compressor discharge pressure (Pc), picked up at a second impact pressure pick up tube 32, engine speed (N), which is transmitted to said control by means of a bevel gearing arrangement 34 and 36 and a governor drive shaft 38, and to the position of a pilots control lever 4t), which is mounted as a rotatable shaft 48 in a control quadrant 50, and which is connected to the fuel control device 20 by means of a link 42, a lever 44 and a shaft 46.

Referring now to FIGURE 2, the governing mechanism is shown contained within a housing of the fuel control device 20, which receives fuel at a pressure P1 from conduit 24 is aninlet passage 82 and an annular chamber 84 formed between the housing and a fixed cylindrical sleeve member 86, Vand which discharges fuel to the metered fuel conduit 17 at a pressure P2 by way of a port SS, formed in the surface of a reciprocable and rotatable hollow cylindrical metering valve member 9i? and adapted to Variably register with a port 92 formed in sleeve 86, a valve chamber 94, valve ports 96, sleeve ports 98, an annular chamber 16@ and a fuel outlet passage 102.

The angular position of metering valve is at all times controlled as a function of compressor discharge pressure by a geared rack member 104 and a gear sector 106 tixed on an extension 108 of valve 90 and having splined gear teeth 110 in mesh with the gear teeth of rack 104. The rack 104 is formed on a rod 105 which is tixedly secured to the movable end of an evacuated belvlows 111. The bellows 111 is mounted in a sealed chamber 113 which receives compressor discharge pressure Pc from the impact pressure pick up tube 32 via a conduit 112. The bellows 111 responds to the compressor discharge pressure Pc in chamber l113 and rotates valve 90 -through rack 104 and sector 106 such that a given angular -position of valve 90 is maintained for each value of compressor discharge pressure Pc.

The axial position of metering valve 90' within sleeve 86 is controlled either by a contoured three-dimensional acceleration cam 114, which is adapted to cooperate with a cam follower and valve rod 116 and 118 during an accelerationof the engine, or by all-speed governing mechanisrn contained within housing chamber 120 and connected to the metering valve by a power servo piston 122 and a walking beam 124. The walking beam 124 is adapted to be variously fulcrumed at a ball joint connection 126 on a servo piston rod 128, at a pivot connection 130 to a rod 132 of a cam load limiting piston 134, or at a ball joint connection 136 on the cam follower rod 118, in a manner to be described. The axial position of ,valve 98 is at all times determined by that mechanism which demands the least quantity of fuel, i.e. the acceleration cam and the governing mechanism are arranged in mutually overriding relation, as hereinafter described, so that that one which tends to fix lthe smallest area at metering ports 88, 92 controls.

The acceleration cam 114 is mounted on a shaft 138 which is rotatably and axially actuable as a function of certain engine operating parameters such as engine speed and compressor inlet temperature, respectively. The shaft 138 carries a pinion 139 lixed secured thereto, said pinion being engaged with and rotated by a rack 141 formed on a rod xedly secured to a pitson 143.V The pistou 143 is responsive to a servo pressure P5 which is controlled by a half ball servo valve 145. The half ball servo valve 145 is positioned as a function of engine speed by apivot lever 147 operatively connected between said half ball servo valve and the centrifugal weights 160. The shaft 138 is positioned axially by a bellows 149 which has an operative connection with the temperature bulb 28 lby means of a tube 151, the bellows and tube being lled with suitable fluid or material responsive to temperature changes registered by the bulb 28. The structural details and mode of operation of the above mentioned mechanism for axially and rotatably actuating the shaft 138 is conventional and will be easily understood by those skilled in the art. The acceleration fuel flow schedule may be controlled by the cam 114 and rack 104 to vary as a predetermined function of Pc, Ta and N, such that maximum allowable turbine inlet temperature is substantially maintained throughout a predetermined portion of the acceleration schedule, and the phenomenon known as compressor surge or stall is not encountered. The angular and axial position controls for valve 90 at all times cooperate to define a metering area 140 formed at the portion of registry between ports 88 and 92. For example, as compressor discharge pressure increases, rack 104 actuates gear sector 106 to move valve 90 in a clockwise direction, as viewed from the top thereof, increasing dimension x of the metering port in proportion to the increase in said pressure, and variation in any engine operating parameter which effects a change in the axial position of metering Valve 90 varies dimension y of port 140.

Preferably, a constant fuel pressure differential is at all times maintained across metering port 140 by a regulator valve means generally shown at 153 and preferably of the type disclosed in U.S. Patent No. 2,689,606, grantedA September 21, 1954, to Frank C. Mock and assigned to the present assignee, whereby the iiow through said port is always a predetermined function of only those engine operating parameters which control the angular and axial positions of valve 90.

The governor mechanism contained within chamber connects the pilot controlled levers 40 and 44 to the power servo piston 122 by way of the shaft 46, a governor setting cam 142 mounted on shaft 46 for rotation therewith and adapted to be axially actuable by compressor inlet pressure and/ or temperature responsive mechanism, a cam follower rod 144, a speed selecting governor spring 146 mounted between spring retainers which abut the rod 144 and a servo valve lever 148, fulcrumed at 150, a half-ball type servo valve 152 connected to the lever 148 and cooperating with a discharge oriiice 154 in a servo pressure passage 156, a governor feedback tension spring 158 resiliently connecting the servo piston rod 128 with the left end of servo lever 148, and a pair of rotatable centrifugal weights 160 mounted on brackets 162 at pivots 164 for rotation with a mounting plate 166 which is keyed to the governor drive shaft 38, said centrifugal weights having foot members 16S in abutment with a ange of a force transmitting rod 172, which in turn abuis a lever 174 pivoted at 176 for transmitting a moment of the force output of weights 160 to servo lever 148 at a governor slope adjusting nut 178. The compressor inlet pressure responsive mechanism may take the form of an evacuated bellows 155 suitably mounted in a sealed chamber 157 which receives compressor inlet pressure from impact pressure pick up tube 30 via a conduit 159. A rod 161 extends from the movable end of bellows 155 into engagement with the shaft 46 such that a given axial position of the cam 142 is maintained for each value of the compressor inlet pressure. The bellows 155 may be replaced by temperature sensing mechanism similar to `the afore-mentioned temperature sensing bulb 28, tube 151 and bellows 149 in which case the shaft 46 could be actuated as a function of compressor inlet temperature Ta. The compressor inlet pressure and temperature re- .sponsive mechanism may be combined to form a conventional density sensitive apparatus, not shown, such that the shaft 46 is positioned as a function of compressor inlet pressure and temperature.

The servo pressure (P5) passage 156 connects the P1 pressure chamber 84 to chamber 120, which is connected to pump by-pass conduit 26 by a passage, not shown, at pump inlet pressure lo. A discharge pressure regulator valve in passage 156 is adapted to maintain a constant pressure in said passage upstream of a calibrated restriction 182, the pressure Ps beneath power piston 122 therefore varying solely as a function of the area ratio between restriction 181. and orifice 154. A power return spring 184 and fuel at. pump inlet pressure in a chamber 186 opposes Ps pressure and spring 158 across piston 122.

A servo balance spring 188, adjustable by means of a screw 190, is utilized to act on lever 148 and 192 for the purpose of compensating for variations in manufacturing tolerances between the designed rates of springs 146 and 158, and for insuring that said latter springs will sustain a pre-load even when the engine is at rest; this latter function insures that neither spring 146 nor 158 will ever be extended to a free height or no-load condition. If either of said springs were ever to reach a 11o-load condition they might become disconnected from their respective retaining means. The effective force output of the balance spring 188 on the servo system remains substantially constant, irrespective of variations in engine operating conditions, as the result of an arrangement which effects a total range of movement of servo valve 152 amounting to only a very few thousandths of an inch.

During all conditions of engine governing and equilibrium operation the fulcrum 130 of Walking beam 124 is maintained in the position shown by a cam load limiting spring 196, which is contained in a cylindrical chamber 198 formed on one side of piston 134. The spring 196 normally maintains piston 134 in abutment with a stop 200 formed in the housing 8%, and the chamber 198 is vented to P pressure in chamber 1211 by an opening 202 in piston 134, a chamber 264, and an aperture 266 in stop 2130.

Operation Referring now to FEGURE 3, sea level acceleration, governing, and steady state engine characteristics are qualitatively illustrated by the curves 221i, 222. and 224, respectively, on the curve chart which is plotted on the coordinates of fuel ow in pounds per hour versus engine speed in rpm. The basic contour of the characteristic seal level acceleration curve 220 is fixed by the contour of the acceleration cam 114, which is actuable as a function of certain engine operating parameters, as hereinbefore described whereas the height or level of said curve above the engine speed coordinate is primarily controlled by the compressor discharge pressure responsive mechanism. The governor fuel cut-off curve 222 illustrates the operation of my force servo type proportional governor from the acceleration curve to a selected speed point on the engine steady state operating curve. The steady state curve 224 illustrates engine fuel iiow demand at all available engine speeds during sea level operation.

Assume that the engine has been started and accelerated to the steady state operating point a in the mid-speed range. In this condition of operation the pilots control lever 4t) will be positioned approximately mid-way between the ends of quadrant Sti, in which position link and lever 42 and 44 position the governor setting cam 142 on shaft 46 so as to impose a speed selecting load on governor spring 146 through cam follower rod 144 which demands that engine operating speed existent at point a. In this condition of operation the half-ball servo Valve 152, being maintained in fixed position by an existent force balance on lever 148, controls a servo pressure Ps in passage 156 which imposes a force on servo piston 122; this force is balanced by the opposing forces of the governor feedback spring 158, the power return spring 1S4, and the pump inlet pressure in chamber 186, whereby the position of walking beam 124 is iixed thereby controlling the axial position of metering valve 9% such that the area of metering port 149 controls that quantity of fuel flow to the nozzle 16 which is necessary to maintain engine speed as set at point a.

The forces acting on servo lever 148 in a direction which tends to move servo valve 152 in an opening direction, i.e. in a direction which would result in an accelerating fuel flow to the engine, consist of the moment forces of governor spring 146, feedback spring 15S, and the fuel pressure force on servo valve 152 at orifice 154, all acting about fulcrum 15G. These moment forces are opposed and balanced by the moment force of the centrifugal weights 169, which varies as the square of engine speed, and the substantially constant moment force of the servo balance spring 183. At any given speed, the engine driven centrifugal weights 161B` generate an effective force output on lever 14S which is equal to the Weight force output times the ratio of the lever arm between rod 172 and fulcrum 176 to the lever arm between nut 17S and fulcrum 176. The force output of weights 160 is therefore diminished at lever 148 by the lever ratio factor, said factor being adjustable by nut 173 to adjust the governor cut-on slope as will be hereinafter described.

The design of the area of servo piston 122, of the lever ratio of walking beam 124 across fulcrum 131i?I between pivots 126 and 136, of the area relation between restriction 182 and orifice 154, and of the rate of feedback spring 158, with a given configuration of metering ports 8S, 92 enables full governor cut-olir action to be effected at any speed with extremely small movement of servo light spring 196.

valve 152, the maximum movement of which will not exceed .004 or .005 inch. This arrangement enables the governing mechanism to actuate metering valve through its total range of movement with substantially no change in the radius of rotation of governor weights 16?, whereby the governor mechanism inherently lends itself to a large degree of design versatility with respect to meeting the operating characteristics of engines having widely different fuel requirements, inasmuch as metering valve 99 may be controlled throughout its axial travel rance with a negligible change in position of the weights 169. From this it is also apparent that a substantially constant force is imposed on lever 148 by the balance spring 188, irrespective of variations in loading on governor spring 146 at different selected speeds.

Assume now that the pilot actuates control lever 40 in a counterclockwise direction to select, say, a maximum operating speed for the engine, as denoted by point d on curve 224, and as illustrated by the position of lever 40 in FIGURE l. As the pilot rotates lever 40 to the indicated maximum speed position, the fuel control unit reacts substantially instantaneously as follows: Governor setting cam 142 is rotated on shaft 46 to a position at which the cam rise imposes maximum compression on governork spring 146 to select that speed at which the governor weight force and other forces acting on servo lever 143 will be in equilibrium, as illustrated at point d; servo lever 14, being no longer in balance, rotates a minute amount in a clockwise direction to increase the area ratio between orice 154 and restriction 182, thereby instantaneously decreasing servo pressure Ps and unbalancing the forces across servo piston 122; the force differential across piston 122 actuates the piston downwardly, and the walking beam 124 is rotated about fulcrum in a clockwise direction until cam follower 116 comes into Contact with acceleration cam 114, in which position the metering area has increased to effect an increasein fuel ow from point a on curve 224 to point b on curve 220.

The amount of compression imparted to governor spring 146 which is in excess of that necessary to effect an acceleration to point g on curve 224 results in a continued downward movement of piston 122, following contact of cam follower 116 with acceleration cam 114, about anew fulcrum 136, as pivot 130 and piston 134 move downwardly from the normallyV fixed position thereof aga-inst spring 196 until piston 122 reaches a position at which the decreased loadings on springs 134 and 158 plus the 10W pressure in chamber 186 again balances the force of pressure P5; the loading force imposed on the acceleration cam through follower 116 is thereby held to a relatively small amount, which results only from the degree of compression imparted to the The increase of fuel flow from point a to point b results in an excess over that required to run the engine at the existing speed, and acceleration proceeds along curve 220, as determined by the contour of cam 114 which effects an opening movement of metering valve 90 in an axial direction; at the same time compressor discharge pressure responsive rack 104 eects an yopening rotational movement of valve 96. The combined axial and rotational movement of Valve 9G' results in a fuel flow versus speed characteristic such as is illustrated by the curve segment bc. As acceleration of the engine proceeds from point b, and the rate of opening movement of valve 96 is determined by the rotating cam. 114, it is apparent that the fulcrum of walking beam 124 will shift from pivot 136 to pivot 126, as spring 196 actuates piston 134 and pivot 13@ upwardly until the piston is again in contact with the stop 20G. During such upward movement of piston 134, beam 124 is rotated in a clockwise direction about pivot 126, contact between follower 116 and the decreasing rise of cam 114 being thereby positively maintained, while the engine is accelerating, with a cam loadspenen ing force which varies within chosen design limits as a function of the rate of spring 196. It is apparent that selection of a low rate cam loading spring will insure against exceesive loads being placed on the cam 114 during acceleration of the engine, with resultant long cam life.

At the speed indicated at point c, the centrifugal Weights v160 have generated a force output which is suihcient to begin to overcome the force moments of governor spring 145 and feedback spring 158, said spring 158 being, at this time, in a relatively relaxed condition. Servo valve 152 therefore moves slightly upwardly to reduce the area ratio between orifice 154 and restriction `182, thereby producing an immediate increase in pressure Ps which unbalances servo piston 122 in an upward direction, resulting in a counterclockwise rotation of walking beam 124 about fulcrum 130 and a movement of follower 116 away from acceleration cam 114 to reduce the dimension y of metering port 140. As piston 122 moves upwardly, feedback spring 158 elongates and produces an increasing moment of force on servo lever 148 in the same direction as that produced by governor spring 146. This action results in continued acceleration of the engine as fuel flow decreases along the governor break curve 222, inasmuch as the increasing force output of spring 158 necessitates a proportional increase in engine speed, so that the effective force output of weights 160 plus the substantially constant force output of balance spring 188 may balance the effective force outputs of spring 146 and 158 at equilibrium operation. In other words, neglecting the substantially constant forces of springs 146 and 188, and also the small hydraulic force acting on valve 152, the increasing load on feedback spring 158 feeds into the governor mechanism a demand increase in engine speed which is proportional to the decrease in fuel ilow along the governor break curve 222. This proportional governing characteristic imparts a negative slope to the fuel cut-off curve which is afunction of the rate of the feedback spring, and permits the engine to reach a stable equilibrium point of operation.

An unstable condition which is generally known as hunting in the governor art, would result if the feedback spring were not included in the governor mechanism; i.e. without spring 158 a substantially vertical or isochronous governing characteristic, as illustrated by curve edf, would result due to the relatively constant force output of governor spring 146 at any given selected speed. Again, without spring 158 a slight variation in engine speed above or below that selected by the pilot would result in movement of servo valve 152 to cause potentially maximum movement of servo piston 122 and metering valve 90 in seeking a condition of equilibrium.

When the fuel flow required to run the engine is being metered at point d, no accelerating engine torque exists, and the governing mechanism is in a state of equilibrium; i.e. the moments of forces acting on lever 148 in a clockwise direction exactly balances the opposing counterclockwise moments, and the positions of servo piston 122 and metering valve 90 are fixed. If, for any reason, the engine should tend to increase or decrease speed from point d, as with a decrease or increase in the density of the surrounding atmosphere, respectively, fuel now ould tend to vary along curve ,222 to reestablish equilibrium on the new engine operating curve.

With any given xed design rate of spring 153, the slope of the governor break curve may be manually adjusted as desired by nut 178, inward adjustment on lever 174 effecting an increase in the slope of the governor cut-off curve as a result of the increased effective force output of weights 160 with a given increase in engine r.p.m., and outward movement of nut 178 on lever 174 effecting a decreased slope of said curve.

The relative arrangement of parts, as shown in Fl"- URES l and 2, approximates that which would exist if an engine were being controlled by Ymy device to operate at point d. It should be remembered that the fulcrum of the walking beam 124 maintains a fixed position, as shown, throughout governor cut-off and equilibrium operation at any selected speed, the fulcrum of beam 124 shifting to pivot 136 only during downward motion of piston 134 against spring 195 following initial contact of cam follower 116 with the acceleration cam, and that said fulcrum shifts from pivot 136 to pivot 126 only during upward movement of piston 134 to the position indicated during an acceleration of the engine to the beginning of governor cut-off action.

A deceleration of the engine is effected by relaxing the loading of spring 146 by counterclockwise movement ofpilots lever 4i), which imbalances the forces acting on lever 14g and effects an upward movement of piston 122 until piston rod 123 contacts a deceleration stop 226; as beam 124 rotates about fulcrum 130 in a counterclockwise direction to the position determined by stop 226, the axial position of metering valve 90 and the y dimension of metering area 140 are xed. Fuel ilow immediately decreases along the curve df, and the engine then begins to decelerate along the curve fh as metering area dimension x decreases with decreasing compressor discharge pressure. At point h the force output of Weights has decreased to such an extent that spring 146 begins to overcome said weight force output and move servo valve 152 slightly open to cause an opening movement of the metering valve 96, which increases fuel flow as the engine speed decreases along the proportional governor break curve ha. At point a the control again reaches a condition of equilibrium.

From the above description, it is now apparent that I have provided a force type governor servo mechanism having a built in proportional governor break characteristic with means for adjusting the slope thereof as desired, plus means for minimizing the loading to which an acceleration cam is subject during an acceleration of the engine.

It will be apparent to those skilled in the art that various changes in the structure and relative arrangement of parts may be made without departing from the scope of my invention.

I claim:

l. in a fuel feed and power control system for a gas turbine engine having a burner, a fuel conduit for conducting fuel to the burner, valve means for regulating the flow of fuel through said conduit, a movable maximum flow stop operatively connected to said valve means for limiting the rate of valve opening movement during an acceleration of the engine, and a force type servo governor mechanism operatively connected to said valve means for controlling the fiow regulating function of said Valve means as an inverse function of engine speed inciuding servo valve means, resilient means operatively connected to said servo valve means for imposing an engine speed selecting force thereon, the force output of said resilient means being substantially constant during governor cutoff action to any given selected engine speed, an engine speed responsive element operatively connected to said servo valve means for transmitting a force thereto which opposes said speed selecting force and which varies as a function of existing engine speed, the force output of said speed responsive element varying substantially only with variations in engine speed during governor cut-ofi action to a selected speed, and a yielding means operatively connecting said servo valve means to said governor valve means for transmitting a force to said servo valve means which varies as a function of the position of said governor valve means during governor cut-off action to a selected Vengine speed. i

2. In a fuel feed and power control system for a gas turbine engine having a burner, a fuel conduit for conducting fuel to the burner, valve means for regulating the iow of fuel through said conduit, a movable maximum ow stop for limiting the rate of valve opening movement during an acceleration of the engine, and engine speed governor mechanism for controlling the flow regulating fulcrum being shiftable to said iirst pivot means following the initiation of control of said valve means by said movable stop.

3. A fuel feed and power control system as claimed in claim 2 wherein said fulcrum is shiftable from said first pivot means to said third pivot means during valve controlling movement of said stop.

4. A fuel feed and power control system as claimed in claim 3 wherein said fulcrum is shiftabie from said third pivot means to said second pivot means during actuation of said valve means away from said stop by said engine speed responsive means.

5. In a fuel feed and power control system for a gas turbine engine having a burner, a fuel conduit for conducting fuel to the burner, valve means for regulating the ow of fuel through said conduit, valve stop means operatively connected to said valve means during an acceleration of the engine, and engine speed governing means for controlling the ow regulating position of said valve means as an inverse function of engine speed during governor cut-off operation following an acceleration of the engine including a walking beam, first pivot means connecting said valve means to said beam, second pivot means operatively connecting yielding means to said beam, and third pivot means operatively connecting engine speed responsive means to said beam, said second pivot means being adapted to function as a beam fulcrum during governor cut-off operation, and means for adjusting said engine speed responsive means to select an operating speed for the engine, said fulcrum being shiftable to said first pivot means following selection of a new engine operating speed by said adjusting means.

6. A fuel feed and power control system as claimed in claim 5 wherein said fulcrum is shiftable from said first pivot means to said third pivot means during an acceleration of the engine.

7. In a fuel feed and power control system for a gas turbine engine having a burner, a fuel conduit for conducting fuel to the burner, vaive means for regulating the ow of fuel through said conduit, a contoured acceleration cam adapted to be operatively connected to said valve means during an acceleration of the engine, and an engine speed responsive governor mechanism operatively connected to said valve means for controlling the flow regulating position of said valve means during governor cut-olf and equilibrium operation of the engine, said last mentioned operative connection including lever means having a movable fulcrum, said fulcrum being movable from a rst to a second position along said lever means l following actuation of said valve means into an operative connection with said cam means.

8. A fuel feed and power control system as claimed in claim 7 wherein said fulcrum moves from said second position to a third position Whenever the position of said valve means varies with the contour of said cam.

9. An engine speed governor comprisingV governor valve means for controlling the flow of motive fluid to the engine, engine speed responsive means operatively connected to said valve means for controlling the flow regulating position thereof as an inverse function of engine speed, and means operatively connected to said speed responsive means for selecting an operating speed for the engine, said rst mentioned operative connection including a walking beam having a fulcrum, rst pivot means operatively connected to Walking beam and said valve means, second pivot means operatively connected to said Walking beam and said engine speed responsive means, and third pivot means operatively connected to said walking beam intermediate said frst and second pivot means and supported by resilient means, said fulcrum being shiftable between said first, second and third pivot means during an acceleration of ythe engine from a first to a second selected engine speed.

10. In a fuel fed and power control system for a gas turbine engine having a burner, a fuel conduit for conducting fuel to the burner, valve means for regulating the ow of fuel through said conduit, a maximum flow stop adapted to be operatively connected to said valve means during an acceleration of the engine, engine speed responsive means operatively connected to said valve means for controlling the flow regulating function thereof as an inverse function of engine speed, said operative connection including lever means connected to said engine speed responsive means, means operatively connected to said speed responsive means for selecting an operating speed for the engine, said levermeans having a fulcrum, and resilient means operatively connected to said fulcrum for maintaining a fixed position thereof during control of said valve means by said governor means, said resilient means being adapted to permit a change in the position of said fulcrum during control of said valve means by said maximum flow stop, following the initiation of an acceleration of the engine by said speed selecting means for limiting the force loading on said stop.

References Cited in the iile of this patent UNITED STATES PATENTS 2,253,963 Van Nest Aug. 26, 1941 2,576,352 Neal NOV. 27, 1951 2,604,756 Greenland July 29, 1952 2,638,992 Lindquist May 19, 1953 2,668,414 Lee Feb. 9, 1954 2,675,674 Lee Apr. 20, 1954 2,712,219 Warne July 5, 1955 2,720,752 Chandler et al Oct. 18, 1955 2,759,549 Best Aug. 21, 1956 2,764,868 Watson et al. Oct. 2, 1956

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