CA1215550A - Pressure responsive actuator mechanism - Google Patents
Pressure responsive actuator mechanismInfo
- Publication number
- CA1215550A CA1215550A CA000501747A CA501747A CA1215550A CA 1215550 A CA1215550 A CA 1215550A CA 000501747 A CA000501747 A CA 000501747A CA 501747 A CA501747 A CA 501747A CA 1215550 A CA1215550 A CA 1215550A
- Authority
- CA
- Canada
- Prior art keywords
- diaphragm
- actuator
- valve
- housing
- pressure
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired
Links
Landscapes
- Fuel-Injection Apparatus (AREA)
Abstract
Abstract A pressure responsive actuator mechanism consists of a housing and a diaphragm dividing the interior of the housing into first and second chambers. An actuator support is located in the first chamber and is carried by the diaphragm. An actuator is fixed at one end to this support to extend therefrom through one end of the housing to the exterior thereof. The actuator support and the diaphragm are biased towards the opposite end of the housing. A fluid is admitted at a pressure that may vary into the second chamber to impose on the diaphragm a force opposing the biasing force. As the fluid pressure varies, the actuator will be shifted in position relative to the housing.
Description
~1555~
PRESSURE RESPONSIVE ACTUATOR MECHANISM
TECHNICAL FIELD
This invention relates to a pressure responsive actuator mechanism, the application being a division of Canadian patent application Serial No. 421,309 filed February 10, 1983.
BACRGROUND ART
Many gas turbine fuel controls have heretofore been proposed. Among these are devices disclosed in U.S.
Patents Nos. 2,697,909, issued December 28, 1954, to Chandler; 2,796,733 issued June 25, 1957, to Pearl et al;
PRESSURE RESPONSIVE ACTUATOR MECHANISM
TECHNICAL FIELD
This invention relates to a pressure responsive actuator mechanism, the application being a division of Canadian patent application Serial No. 421,309 filed February 10, 1983.
BACRGROUND ART
Many gas turbine fuel controls have heretofore been proposed. Among these are devices disclosed in U.S.
Patents Nos. 2,697,909, issued December 28, 1954, to Chandler; 2,796,733 issued June 25, 1957, to Pearl et al;
2,822,666 issued February 11, 1958, to Best; 2,917,061 ' issued December 15, 1959, to ~ongfellow; 2,941,601 issued June 21~ 1960, to Best; 2,957,488 issued October 25, 1960, to Farkas; 3,052,095 issued September 4, 1962, to Prachar;
3,139,727 issued July 7, 1964, to Torell; 3,156,291 issued November 10, 1964, to Cornell; 3,164,161 issued January 5, 1965, to Tyler; 3,427,804 issued February 18, 1969, to Lawrence; 3,469,397 issued September 30, 1969, to Parker;
2Q 3,492,814 issued February 3, 1970, to Urban; 3,606,754 issued September 21~ 1971, to White; 3,712,055 issued January 23, 1973, tc~ McCabe; 3,878,676 issued April 22, 1975, to Hitzelberger; 3,879,936 issued April 29, 1975, to Stoltman; and 3,939,649 issued February 24, 1976, to McCabe.
The novel liquid fuel controllers disclosed herein are nonetheless unique and possess a combination o~
advantages not available in any known liquid fuel controller heretofore proposed.
- 1 - `' lZ~5SS0 ~
DISCLOSURE OF THE INVENTION
, In one aspect of the present invention, there is provided a pressure responsive actuator mechanism which comprises: a housing, a diaphragm ~ividing the interior of said housing into first and second chambers; an actuator support.located in said first chamber and carried by said diaphragm; an actuator fixed at one end thereof to said support, said actuator extending therefrom through one end of said housing to the exterior thereof; means biasing said actuator support and said diaphragm toward the opposite end of said housing; and means for admitting a fluid at a pressure which may vary into said second chamber to impose on said diaphragm a force opposing that exerted by said biasing means whereby, as the fluid pressure varies, said actuator member will be shifted in position relative to said housing.
In another aspect, the invention provides a differential pressure actuated regulator comprising: a housing with a chamber therein; a diaphragm in and spanning said housing; a fluid inlet passage in said housing and communicating with said chamber on one side of said diaphragm; a fluid outlet passage communica.ting witn said chamber on said one side of said diaphragm and extending through said housing to its exterior; a bypass passage communicating with said chamber on said one side of said diaphragm; and extending through said housing to the exterior thereof; a valve seat in said bypass passage at the end thereof which communicates with said chamber; a mount fixed to said diaphragm; a valve swivel carried at one end by ~Z1555~
said mount; a valve member fixed to the other end of said swivel in alignment with said valvé seat; spring means in said chamber on the other side of said diaphragm for biasing said diaphragm toward said valve seat to thereby seat said valve against said seat and keep fluid from flowing from said inlet passage through said chamber into said bypass passage; and means for admitting fluid at a second pressure into said other side o~ said diaphragm at a pressure which is different from, and lower than~ that of the fluid admitted into said one side of said diaphragm through said fluid inlet passage whereby, when the differential in the forces attributable to said fluid pressures exceeds the force exerted by said biasing means, said valve will open and fluid will be discharged through said bypass passage at a rate that will keep the pressure on the fluid discharged through said outlet passage essentially constant.
Other important features and advantages of the invention will become apparent from the appended claims and the detailed description in conjunction with the : accompanying drawingsO
BRIEF DESCRIPTION OF THE DRAWINGS
FigO lA and lB, taken together, constitute a partially sectioned side view of one gas turbine engine which may be equipped with a fuel flow scheduling controller;
Fig. 2 is a schematic view of a gas turbine engine liquid fuel supply system;
~, ~
~a2~
Figs. 3-6 are schematic views of a differen-tial linkage assembly employed in the flow controller of Fig. 2 showing how the assembly functions at different stages in the operating cycle of the turbine engine.
Fig. 7 is a section through a compressor dis-charye pressure responsive actuator employed in the con-troller;
Fig. 8 is an elevation of an embodiment of a fuelvalve assembly employed in the controller with part of the casing of the assembly being broken away to show a fuel flow metering valve incorporated in the assembly;
Fig. 9 is a section through the fuel valve assembly showing a differential pressure valve incor-porated therein;
Fig. 10 is a partial section through the assembly showing a two-stage relief valve incorporated therein;
Fig. 11 is a similar view showing a pressurizing valve incorporated in the assembly~
Fig. 12 is a graph showing the fuel schedule maintained by the controller;
Fig. 13 is a plan view of the differential link-age mechanism shown diagrammatically in Figs. 3-6; part of the assembly's dust cover has been broken away to show the nternal, working parts of the assembly; and Fig. 14 is a graph included to illustrate the operation of the mechanical differential linkage.
BEST MODE FOR CARRYING OUT THE INVENTION
Referrlng to the drawings, Fig. lA and lB depict a two-shaft, gas turbine engine 16 equipped with a fuel .
s~o supply system 18 which includes a fuel scheduling con-- troller 30.
Engine 16 has, among its major components, a fifteen-stage axial flow compressor 22 with a radialaxial inlet 24, inlet guide vanes 25, stators 28, and a fifteen-stage rotor 30. Th~ inlet guide vanes 26 and stators 28 are supported from the compressor housing-32 with the guide vanes and stators 23-1 through 28-5 of the first five stages being pivotally mounted so that they can be adjusted to control the flow of air through the compressor.
Each of the fifteen stages of the compressor rotor 30 consists of a disc 34 with radially extending blades 36 fixed to the periphery of the disc. The stages are integrated into a unitary structure as by electron beam welding.
The high pressure air discharged from compressor 22 flows through a diverging diffuser 38 and an enlarged dump plenum 40 to an annular combustor 42 supported in an insulated combustor case 44.
Combustor 42, ~hich is of the annula-r type, includes inner and outer liners 46 and 48 concentric with the axial centerline S0 of the engine and an annular com-bustor dome 52 spanning the gap between the liners at the forward or upstream end of the combustor.
Injectors 56 slidably mounted in dome 52 at generally equidistantly spaced intervals therearound dis-charge Euel into the annular combustion zone 58 betweeninner and outer liners 46 and 48. The fuel flows from fuel supply system 18 to injectors 56 through holders 60 which extend outwardly from combustor don~e 52 through com-bustor case 44.
~2~ S~
rhe compressor discharge air heated by combustor ~ and the combustion products generated in the combustor are expanded through a two-stage gas producer turbine 62 and then through a two-stage power turbine 64. The tur-bines are rotatably supported in a nozzle case 66 mounted in an ann~lar t~rbine housing or case 67.
The gas producer turbine 62 has a two-stage rotor 68 and sta~ionary, internally cooled, first and second stage nozzles 70 and 72. First stage nozzles 70 are assembled into an annual array or ring as are the second stage nozzles 72.
The two stages 78 and 86 of the gas producer tur-bine rotor 68 are bolted to each other and, in cantilever fashion, to the rear end of a forwardly extending shaft ~6. Shaft 96 is coupled through rear compressor hub 98 to compressor rotor 30, thereby drive-connecting gas producer turbine 62 to the compressor.
T~e compressor and qas producer turbines are rotatably supported by a thrust bearing 100 and by tapered land bearings 102, 104, and 106. Bearings 100 and 102 engage the ~ront compressor hub 108 which is bolted to 20 rotor 30 and is drive-connected to an accessory drive 110.
Power turbine 64 includes first and second stage nozzles 112 and 114, also supported from nozzle case 66, and a rotor 116 having a first, bladed stage 118 and a second, bladed stage 120. The first and second stage 25 nozzles 112 and 114 of power turbine 6~ are assembled into stationary annular arrays or rings.
Power turbine rotor stages 118 and 120 are bolted together for concomitant rotation. Rotor 116 is bolted to ~21SS5(~
a power turbine shaft assembly 128 rotatably supported by tapered land bearings 130 and 132 and a thrust bearing 134. rrhe shaft assembly is connected through a coupling 136 to an output shaft assembly 138 which furnishes the input for a generator, booster compressor, mechanical drive, or other driven unit (not shown).
The final major component of turbine engine 16 shown in Figs. lA and lB is an exhaust duct 140 for the gases discharged from power turbine 64.
Referring to Fig. 2, the fuel supply system 18 includes among its major components, an electronic control module 142 which converts a number of inputs including the load upon gas turbine engine 16 into an electrical fuel demand signal and a hydroelectric actuator 143 which con-verts the electrical fuel demand signal into a mechanical input to fuel flow controller 20.
Also, fuel supply system 18 includes a pump 144 for effecting a flow of liquid fue~ from a liquid fuel supply 146~through a filter 148 first to liquid fuel flow controller 2~ and then to the fuel injectors 55 of gas turbine engine 16.
Components of fuel supply system 18 such as hydroelectric actuator 143, pump 144, and filter 148 can be of any suitable character and are readily available.
Consequently, they will not be described further.
Nor is it considered necessary to describe elec-tronic control module 142 in detail. Any suitable system generating an appropriate fuel demand signal can be employed in fuel supply system 18.
12~555~
The electronic control module 142 processes - signals representative of the load upon turbine engine 16, the speeds of power turbine 64 (NpT) and gas producer turbine 62 (NGp), the temperature of the hot gases supplied to the gas producer turbine (T5), and the power turbine temperature (TpT) into output signals which can be employed to control the adjustable inlet guide vanes 26 of turbine 16 (GV) and the turbine bleed valve which is not shown in the drawings (BV). The module 142 also fur nishes the ~uel demand signal which actuator 143 converts to a mechanical input to fuel supply system flow con-lO troller 20.
Referring still to ~ig. 2, the major components of fuel flow controller 20 include a PCD actuator 150 which produces a mechanical output indicative of the pres-sure of the air discharged from the compressor 22.
15 Another major component is a differential linkage assembly 152 which mechanically multiplies the variable inputs frolT
hydroelectric actuator 143 and PCD actuator 150. The output of the differential linkage assembly serves as an actuator for a fuel metering valve 154 through whicn the , 20 fuel flows on its way from the discharge side of pump 144 to the fuel injectors 56 of gas turbine engine 16.
Metering of the liquid fuel to gas turbine engine 16 by the product of the variable fuel demand and com-pressor discharge pressure inputs generated in 25 differential linkage assembly 152 permits the engine to be rapidly started up and operated under varying loads without the danger of compressor surge. It also allows the engine to be rapidly decelerated without flameout.
~2~SSS~
Referring still to Fig. 2, fuel flow controller 20 also includes a ~P regulator 156 to maintain a con-trolled pressure differential across metering valve 154, a pressurizing valve 158 which maintains sufficient back pressure to enable the 4P regulator to function properly, and a relief valve 160 which prevents excess pressure build-up in fluid supply system 18.
Referring to Figs. 3-6, differential linkage assembly 152 has been shown in a somewhat diagrammatic form to simplify explanation. In actual practice, the links of that assembly are con~igured and arranged in a slightly different manner and, in some cases drilled to lighten the components and to reduce dynamic loads.
Movable components are suppo~ted by precision ball bearings to insure accurate response of metering valve 154 to the inputs from hydroelectric actuator 143 and PCD
actuator 150, to prevent metering valve position errors ; which might be caused by reaction of friction loads on PCD actuator 150, and to otherwise insu~e a stable, closed loop operation. A cover will typically be provided to protect the moving components of the linkage assembly from dust and other conditions.
One actual preferred differen~ial linkage assembly is illustrated in Fig. 13, the dust cover being dentified in that figure by reference character 162.
Referring again to Figs. 3-6, differential link-age assembly includes a PCD actua~or input link 164, ahydroelectric actuator input link 165, and an output link of fuel lever 168 which controls metering valve 154.
These links are supported from a base 170 by pivots 172, 174, (see Figure 6), and 176, respectively.
_g_ ~2~S5S~
Also included in the assembly are three links - 178, 180, and 182 which multiply the displacements of links 164 and 166 and impart to output link 168 a movement having a magnitude indicative of the product of the two inputs. Links 178 and 180 are pivotally connected to each other by pivot member 184 and to input and output links 164 and 168, respectively, by pivot members 185 and 188.
Link 1~2 is connected to links 178 and 180 by pivot member 189 and to hydroelectric actuator input link 166 by pivot member 189.
Stops 19û and 192 on base 170 limit the travel of PCD actuator input link 164. Similarly, stops 194 and 195 mounted on base 170 limit the travel of hydroelectric actuator input link 166, and a stop 198 on input link 166 limits the travel of the displacement multiplying links 178, 180, and 182. The stops mechanically limit the range of operation of the liquid fuel controller and provide reference points for assembly and c(alibration of the differential linkage assembly.
As discussed above, differential linkage assembly 152 multiplies the variable fuel demand input from hydro-electric actuator 143 and the variable compressor dis-charge pressure input from PCD actuator 150. By virtue of the relationship between actuator 143 and electronic control module 142, the differential linkage assembly therefore has a fuel flow rate controlling output which:
(a) is responsive to the temperature, speed, and load inputs to the control module; and ~b) is trimmed by varia-tions in compressor discharge pressure.
The control philosophy embodied in differential `~ linkage assembly 152 is shown in Fig. 14 in which "Z"
represents the positions taken by hydroelectric actuator 143 as engine 16 is accelerated to full load (Z=l.0) and to various part loads down to Z=0.4.
There is a straight line relationship between the position of the actuatGr and the rate of fuel flow WF
(and differential linkage assembly). This is modified or trimmed by PCD. For example, if the engine is operating under full load (differential linkage assembly fuel lever at "A") and the PCD drops from pressure "5" to pressure "4", the linkage assembly will move the fuel lever to "B", decreasing the 10w of fuel to engine 16 (WF) from rate "C" to rate "D".
A reverse case is acceleration from cold start (fuel demand maximum and PCD near zero). In this case (see, also, Fig. 4), WF is initially held to rate "E", despite a maximum demand for fuel by actuator 142, and increased to rate "C" as PCD increases to pressure "A".
The foregoing is accomplished in differential linkage assembly 1~2 by variations in the distance between 20 pivot members 174 and 184~
Fig. 13 shows the actual differential linkage assembly 170 depicted diagrammatically in Figures 3-6.
Aside from the previously identified dissimilarities, assembly 170 differs in that it has pivotably connected links l99a and l99b for transmitting the movement of hydroelectric actuator 142 to fuel demand input link 166 of assembly 152. It also has a pivotably mounted crank l99c for transmitting the movement of the PCv actuator : 30 ss~
output to the trim input link 164 of the assembly. These links merely facilitate connections between the interior and exterior of the differential linkage assembly housing. They do not affect the above described operation of the linkage assembly.
Fig. 7 depicts in detail the PCD actuator 150 which includes a housing 200 made up of end wall and adjacent castings 202 and 204 spaced from end wall casting 206 by cylindrical housing member 208. The housing com-ponents are bolted or otherwise fastened together.
Clamped between castings 202 and 204 is a diaphragm 210 which divides the interior of housing 200 into two chambers 212 and 214.
Fixed to diaphragm 210 by a circular clamp 216 and located in housing chamber 212 is an actuator mount 218. Threaded into one end of this mount is an elongated 15 actuator 220 extending through housing cha~ber 212 and through a fitting 222 threaded into end wall casting 206 o~ housing 200 to the exterior of that housing.
Actuator mount 218, along with the actuator and diaphragm 210, is biased toward that end of housing 200 20 defined by casting 202 by a coil spring 224. That spring extends between a spring seat 226 threaded into fitting ~22 and a ~lange 228 integral with and located toward the end of actuator mount 218 into which actuator 220 is threaded.
'rhe actuator, actuator mount, and diaphragm are, in circu~stances related below, also biased in the same direction as they are by spring 224 (i.e., toward end wall casting 202) by a coil spring 230. This spring extends SS~;;(l between a boss 232 at the inner end of a plunger 234 and the clamp 216 b~ which actuator mount 218 is fixed to diaphragm 210.
As shown in Fig. 7, the plunger and spring 230 are located in a cavity 235 in actuator mount 218 with the plunger extending through a bore 236 in clamp 216.
The end of the plunger facing casting 202 and terminating in head 237 is biased toward a stop 238 by a coil spring 2~0. The spring extends between the head 237 of the plunger and clamping plate ~15.
Springs 230 and 240 support valve plunger 234 in bore 236. Clearance between the plunger and clamp 216 is provided to eliminate friction and any possibility of the plunger sticking.
Stop 238 is fixed to the inner end of a shaft 241 threaded into housing member 202. A control knob 242 lS fixed to shaft 241 outside housing 200 allows stop 238 to be rotated through an angle lappro~imately 360) deter-mined by the engag~ment of a lug 2~4 on the stop with a second stop 246 fixed to casting 202.
Rotation of control knob 2~2 displaces stop 238 toward and away from casting 202, altering the force or bias which spring 2~0 is able to exert on diaphragm 210 via clamping plate 216. This is used, in fuel flow controller 30, to calibrate the PCD actuator for the altitude at which turbine engine 16 is operated.
Fluid (compressor discharge air in system 18) is admitted to the second chamber 21~ in the housing 200 of PCD act~ator 150 through an inlet 247 in end wall cast-ing 202. As is apparent from Fig. 7, the force generated by this fluid opposes that generated by springs 224 and 230 and displaces actuator 220 a distance proportional to the fluid pressure. In controlling the flow of fuel to the injectors of a gas turbine engine, actuator 220 fur-nishes to a differential linkage assembly 152 a mechanical input having a magnitude proportional to the engine's dis-charge pressure.
Fig. 7 shows the actuator with its moveable com-ponents positioned as they are with minimum pressure in chamber 214. At this stage of operation, spring 240 is biasing the head 237 of plunger 234 against stop 238.
Springs 224 and 230 are acting in parallel to displace actuator mount 218 toward the end wall casting 202 of the regulator, and actuator 220 is in its most retracted position.
As the pressure of the fluid chamber 214 increases, the resulting force exerted on diaphragm 210 displaces it, together with actuator mount 218 and actua-tor 220, toward the opposite end wall 205 of the actuator against the resistance offered by springs 224 and 230. At a specified pressure level determined by the relative 20 dimensioning of springs 224, 230, and 250, spring 240 becomes fully extended; and the head 237 o~ plunger 234 moves out of contact with stop 238. At this juncture, spring 230 becomes ineffective; thereafter only spring 224 opposes the fluid generated force. In this stage of operation actuator 220 is displaced a greater distance for each unit increase (or decrease) in fluid pressure.
Thus, PCD actuator 150 has two spring rates - a higher one when the fluid pressure in chamber 214 is below 5S~
a specified level and a lower one when the fluid pressure ~ is above that level.
Spring seat 226 is threaded into insert 222, which is kept from moving relative to end wall casting 206 by a snap ring 246a. The spring seat is held against rotation land thereby kept from moving longitudinally in housing 200 and changing the spring rate) by a pin 246b extending from it into end wall casting 206.
However, displacement of the spring seat toward and away from end wall casting 206 to vary the bias exerted by spring 224 on actuator mount 218 and diaphragm 210 can be effected by rotating internally threaded insert 222 ~hich is slotted to facilitate this operation. By doing so, one can vary the fluid pressure at which the break in the spring rate occurs to facilitate the initial calibration of the PCD actuator and permit operation of the actuator to be matched to the particular type of load being driven by t~rbine engine 16. f The metering valve 154 operated by the output from differential linkage assembly 152 is incorporated in a fuel valve assembly 248 (Fig. 8) which includes a ~P
20 regulator 156 and relief valve 160 lsee Fig. 10).
Fuel valve assembly 248 includes a casing 244.
Housed in casing 249 are a valve seat 250 and a metering valve plunger or member 252. The valve seat is threaded into a casting 254 of casing 249.
Valve plunger 252 is supported for rectilinear movement in a guide 256 mounted in casting 254 and clamped in place by a casting 258 which is part of the valve assembly casing or housing 249.
~z~s~
The valve plunger is biased closed (i.e., toward _ valve seat 250) by a coil spring 260 extending between valve guide 256 and a spring seat 262 fixed to the inner end of the valve plunger by a snap ring 263 making the valve fail-safe and eliminating backlash.
The valve design eliminates the customary close fitting spool and sleeve assembly which minimizes hysteresis and enables the metering valve to handle con-taminated fuels because the close tolerances needed in a spool and sleeve assembly are eliminated.
Liquid fuel delivered by pump 144 (See Fig. 2) flows through the external fuel line 264 in which filter 148 is incorporated and then seriatim through an internal passage 266 in casting 258 and an internal passage 268 in valve seat 250. From there the fuel can flow to a chamber 270 in the casting and then through an internal discharge passage 272 to pressurizing valve 158.
Referring to both Figs. 8 and 13, the rate (WF) at which fuel can flow into internal discharge passage 272 is determined by the distance between plunger 252 and valve seat 250 because WF is a product of that distance and the circumference of the passage 268 in valve seat 250, and the latter is a constant. The spacing (typically thousandths of an inch) between the valve plunger and valve seat is, in turn, controlled by the angular position of the fuel valve lever 168 in differential linkage assembly 152 as was discussed above.
More specifically, the position of lever 168 determines the angular position of the pivot shaft 176 on ::~2~
which the lever is mounted and, consequentially, the angu-- lar position of an eccentric 278 non-rotatably fixed to that shaft. Rotation of the eccentric, in turn, displaces a roller 280 carried by the eccentric toward and away from valve seat 250. The roller is confined between lands 282 and 283 on valve plunger 252 with essentially zero clear-ance between the roller and plunger. Consequently, the anyular movement of eccentric 278 is accurately translated into linear movement of valve plunger 252 toward and away ~rom valve seat 250.
It is believed that the operation of the fuel metering valve will be apparent from the foregoing.
Briefly, however, during the start-up of engine 16, fuel pump 14g is actuated. ~uel accordingly flows through line 264 and filter 148 and then through the passage 265 in fuel valve assembly casing 249 into the internal passage 26~ in valve seat 250. The rate of fuel flow is regulated by fuel valve lever 16~ and eccentric 278 in the manner just described. As discussed above in conjunction with the operation of differential linkage assembly 152, the position of the fuel valve lever - and, ultimately, the flow of fuel through metering valve 154 - is a function of the load on turbine engine 16 and the other inputs to electronic controller 14 as well as the compressor dis-charge pressure of the engine.
Operation of metering valve 154 requires that the pressure drop across that valve be accurately controlled because fuel pressure can vary as engine 16 is accele-rated, especially from a start. The ~P regulator, in association with the pressurizing valve 158 incorporated ~215556~
in fuel valve assembly 248, autornatically compensates kor ~- such variations, guaranteeing repeatability of operation.
The ~P regulator is housed in part in the cast-ing 254 of the fuel valve assembly housing and partly in a chamber 290 defined by that casting, a cooperating casting 292, and a cover plate 294.
Housed in casting 254 and clamped between that casting and casting 292 is a diaphragm 296 carrying a spring seat 298 on one side and a valve mount 300 on the other. Also housed in casting 254 are a valve seat 302, which is threaded into the casting, and a valve plunger or member 304.
The second (306) of the two chambers into which the interior of casing 24g is divided by diaphragm 296 communicates with the external fuel line 254 on the upstream side of fuel valve assembly 248 via an internal 15 passage 310 in casting 254. From chamber 306 the fuel flows through internal passage 266~to metering valve 154.
The chamber 305 of ~P regulator 156 also com-municates by way of valve 304, valve seat 302, and an internal passage 316 in casting 254 with an external fuel bypass or return line 318. That line is connected to the fuel supply line on the upstream side of f~el pump 144.
Finally, the chamber 290 in ~P regulator casting 292 on the opposite side of diaphragm 295 from chamber 305 is connerted through internal passage 322 Isee Figure 2) to the Euel passage 272 on the downstream side of metering valve 154. Passage 322 furnishes communication between the cavity 270 in casting 254 on the downstream side of ;iS5~
metering valve seat 250 and the chamber 290 in ~P regula-tor 156. This makes the pressure on the "downstream" side of the metering valve available in chamber 290.
Referring to Fig. 9, the valve plunger 304 is located in a fuel cavity 326 into whlch fuel flows from external fuel line 264 and internal fuel passage 327.
From cavity 326 the fuel flows around the valve plunger through internal passage 266 to metering valve 154. Valve plunger 304 has a conical tip 328 which is engaged with the valve seat 302 to keep fuel from flowing through the internal passage 330 in that seat into fuel bypass passage 316 when the valve is closed.
Plunger 304 is fixed to a valve stem 332 termina-ting, at its end opposite the valve member, in a swivel 334. The latter is supported in a generally hemispherical cavity 336 in valve mount 300 and retained therein by spring seat 298. This insures good alignment between valve member or plunger 304 and valve seat 302 without side loading to provide good seating and eliminate the friction and sticking common to heretofore available designs.
Valve plunger 304 is biased to the closed posi-tion illustrated in Fig. 9 by a coil spring 338. Spring 338 is centered on seat 298 and extends fronn that seat to the cover 394 of the fuel valve assembly casing 249.
As long as the pressure difference across meter-ing valve 154 is at the specified level, the force exerted by spring 338 and that generated by the fuel in chamber 290 at the pressure on the downstream side of metering valve 154 on diaphragm 296 will keep the valve closed.
5~S~
However, should this specified pressure differ-ence be exceeded, the force generated by the fuel at the pressure on the upstream side of the metering valve and applied to the diaphragm will unseat the valve, allowing fuel to flow through the passage 330 in valve seat 302, the passage 316 in casting 254, and external fuel return or bypass line 318 to the upstream side of fuel pump 144.
This bypassing of excess fuel reduces the pressure on the upstream side of the metering valve, bringing the pressure drop across that valve back to the specified level.
Referring to Figs. 2, 8, and 11, it was pointed out above that proper functioning of the ~P regulator ; requires that a minimum back pressure be maintained on the downstream side of metering valve 154 and that this is accomplished by pressurizing valve 158 which keeps the fuel pressure in the entire fuel supply system 18 from lS going below the minimum required for proper operation of the system. It will be remembered(that the pressurizing valve is incorporated in fuel valve assembly 248 (Fig~ 8) and occupies the upper right-hand portion of casting 254.
Referring specificially to Fig. 11, pressurizing 20 valve 158 includes a valve plunger 340 loosely fitted and rectilinearly moveable in a bore 342 formed in casting 254. 0-rings 344 and seals 346 keep fuel from leaking past the valve member. Valve member 340 is biased by a coil spring 348 toward the closed position, in which it : 25 blocks the flow of fuel in metering valve outlet passage 272. The spring is seated in the interior 350 of the valve plunger and extends from the head 352 of the latter into a spring seat 354 threaded into casting 254.
-~0- .
~Z~S55~
Like the other valves described above, and here-- inafter, pressurizing valve 158 has the advantage that close manufacturing tolerances are not required, making it capable of handling dirty fuel without sticking.
Pressurizing valve 158 is viscous damped for better stability by connecting the interior 350 of valve plunger 340 through the internal passage 356 in which orifice 358 is formed, internal passage 360, and a third internal passage in casting 254 (not shown) to internal fuel return passage 316. This also provides a fuel return pressure reference for operation o~ the pressurizing valve.
Especially upon start-up and deceleration or shut-down of engine 16, the back pressure in the fuel line 361 through which the fuel is delivered to the injectors 56 of engine 16 (see Fig. 2) is apt to be too low for proper operation of ~P regulator 156. In this case pres-surizing valve 158 remains closed until the pressure on the downstream side of metering va~ve 154 and in the fuel passage 272 and, ~onsequently, the pressure in chamber 290 of the ~P regulator, becomes high enough for proper operation lif the pressure drops below the specified mini-mum, and the ~P regulator consequently fails to operateproperly, unwanted and deleterious changes in the schedul-ing of fuel to the injectors 56 of engine 16 may occur).
With tne specified minimum pressure in ~P regu-lator chamber 290, the force generated by the fuel flowing into the passage 272 on the downstream side of metering valve 15~ land therefore in ~P regulator chamber 290) ~-becomes high enough to ~nseat and open valve 158, allowing the metered fuel to flow from the pressurizing valve inlet ~2~5~50 passage 272 to the valve outlet passage 351 and then to -- the fuel injectors 56 of gas turbine engine 16.
The pressure in chamber 306 and f-~el cavity 325 - is then sufficient for the excess fuel to flo~ from fuel cavity 326 into internal bypass passage 315 and, from there, into external fuel line 318, bypassing fuel from the upstream side of metering valve 154 at a rate which is determined by the spacing between ~ P valve plunger 304 and valve seat 302 to keep the pressure drop across seat 250 of metering valve 154 constant.
Should the pressure in passage 272 drop below that required by the ~P regulator during the operation of engine 15, spring 348 will close the pressurizing valve to the extent necessary to restore the pressure in passage 272 to the specified level.
Referring to Figs. 2, 9, and 10, the last o~ the above-discussed major components of fuel supply system 18 is relief valve assembly 160 which occupies the lower right-hand portion of fuel valve assembly casting 254.
The relief valve assembly includes a main relief valve 352 and a pilot valve 364. This two-stage arrange-ment affords superior regulation and is more stable than the usual poppet type of relief valve.
Referring to Fig. 10, the main relief valve is similar to the pressurizing valve and has the same advan-tages. Fuel reaches the valve via internal fuel passage 366 in casting 254 which communicates with external fuel return line 264 tsee Fig. 2). With the valve open the fuel can flow through an internal return passage 368 which communicates via return passsage 316 to previously dis-cussed external fuel return line 318 (Figs. 9 and 2).
:~2:~55~
Absent excessive fuel pressure in fuel supply ~~ system 18, the fuel is kept from flowing from internal passage 366 into return passage 368 by valve plunger 370.
Plunger 370 is loosely fitted and rectilinearly moveable in a bore 372 in casting 254 and is sealed against leakage by an O~ring 374 and seal 376.
Plunger 370 is biased to the closed position shown in Fig. 10 by a coil spring 378. The spring is seated in a cavity 380 in plunger 370 and extends from the head 382 of the plunger to a spring seat 384 threaded into casting 254.
~lain relief valve 362 differs from pressurizing valve 158 in that there is an orifice 386 in the head 382 - of the valve. Fuel can flow at a controlled rate from passage 366 through that orifice, the interior 380 of the valve, and a passage 388 in casting 254 to pilot valve 364.
As shown in Fig. 10, the pilot valve incl~des a valve seat 390 threaded into casting 254. Fuel can flow from passage 388, an orificed passage 392 in the valve seat, and an internal passage 394 in casting 254 into ~ internal fuel return ~assage 368 when the valve is opened.
:~ 20 Absent excessive pressure in fuel supply system18, however, such flow of fuel is prevented by valve member 396 which is biased against the valve seat by a coil spring 398. The latter extends between an annular ièdge 400 on the valve member and a spring seat 402 which is threaded into valve seat 390.
Should the pressure in fluid supply system 1~
exceed the specified maximum, the pressure of that fluid, which can flow from p~ssage 366 into valve seat 390 and ~%~SS50 against valve plunger 396 will exert sufficient force on the plunger to unseat it.
This allows the fluid in main relief valve plunger 370 to drain into return passage 36~, creating a pressure differential across the head 382 of the main relieE valve. This overcomes the valve closing force exerted by main relief valve spring 378 which then opens, returning fuel to the upstream side of fuel delivery pump 144 or keep components of fuel supply system 1~ from being damaged.
Operation of fuel supply system 18 will be clear by reference to Figs. 3-6 and 12.
~ s shown in Fig. 12, flow of fuel to engine 16 is carefully scheduled to keep the ratio of fuel flow WF to compressoe discharge pressure PCD between limits at which flameout would occur on one hand and surge on the other. As discussed above, this is accomplished by multi-plying fuel demand and compressor discharge inputs in differential lin~age mechanism 152 and employing the pro-duct of those two inputs to regulate the position of fuel metering valve 154.
` 20 Fig. 3 shows the differential linkage mechanism with its various links positioned as they are when engine 16 is started up. As the engine is brought up to the speed at which combustion can be effected and the engine operated under its own power, a maximum fuel demand signal is transmitted to the differential linkage mechanism as shown in Fig. 4; and metering valve 154 is opened, allow-ing fuel to be pumped to the injectors 56 of engine 16.
~155~ii0 However, as th~ compressor discharge pressure is rela-tively low in this stage of operation and as the PCD
actuator lS0 is operating under the higher of its two spring rates, the product of the two inputs is relatively low as indicated in Yigure 4; and fuel is supplied to the engine at a relatively low rate which, as shown in Figure 12, avoids engine surge.
At the specified breakover point, discharge pres~
sure will will have increased considerably as shown in Fig. 12. The PCD actuator will have shifted to opera-tion at the lower of its two spring rates. Thus, as is apparent from Figs. 5 and 12, metering valve 154 is rapidly opened wider at this point to rapidly increase the rate of flow of fuel to injectors 56 and to then pro-gressively open the valve until fully open, producing maximum acceleration of the engine.
As full speed is reached, the fuel demand decreases as shown in Fig. 12 and the engine operates along the steady state curve.
As shown in Fig. ~, shutdown of the engine results in the fuel demand dropping to zero with a corres-ponding reduction in the fuel supply to engine 16 as shownin Figure 12. However, because the compressor discharge pressure is still high at this stage, only a partial reduction in the flow of fuel is permitted with this being followed with a more gradual reduction as the compressor discharge pressure drops. This prevents flameout.
The invention may be embodied in other specific forms without departing from the spirit or essential characteristics thereof. For example, a three dimensional cam or a mechanical link supported on a linearly displace-able pivot member may be employed to effect the multiplication of the compressor discharge and fuel demand variables. Therefore, the embodiments described above are to be considered in all respects as illustrative and not restrictive.
~5
2Q 3,492,814 issued February 3, 1970, to Urban; 3,606,754 issued September 21~ 1971, to White; 3,712,055 issued January 23, 1973, tc~ McCabe; 3,878,676 issued April 22, 1975, to Hitzelberger; 3,879,936 issued April 29, 1975, to Stoltman; and 3,939,649 issued February 24, 1976, to McCabe.
The novel liquid fuel controllers disclosed herein are nonetheless unique and possess a combination o~
advantages not available in any known liquid fuel controller heretofore proposed.
- 1 - `' lZ~5SS0 ~
DISCLOSURE OF THE INVENTION
, In one aspect of the present invention, there is provided a pressure responsive actuator mechanism which comprises: a housing, a diaphragm ~ividing the interior of said housing into first and second chambers; an actuator support.located in said first chamber and carried by said diaphragm; an actuator fixed at one end thereof to said support, said actuator extending therefrom through one end of said housing to the exterior thereof; means biasing said actuator support and said diaphragm toward the opposite end of said housing; and means for admitting a fluid at a pressure which may vary into said second chamber to impose on said diaphragm a force opposing that exerted by said biasing means whereby, as the fluid pressure varies, said actuator member will be shifted in position relative to said housing.
In another aspect, the invention provides a differential pressure actuated regulator comprising: a housing with a chamber therein; a diaphragm in and spanning said housing; a fluid inlet passage in said housing and communicating with said chamber on one side of said diaphragm; a fluid outlet passage communica.ting witn said chamber on said one side of said diaphragm and extending through said housing to its exterior; a bypass passage communicating with said chamber on said one side of said diaphragm; and extending through said housing to the exterior thereof; a valve seat in said bypass passage at the end thereof which communicates with said chamber; a mount fixed to said diaphragm; a valve swivel carried at one end by ~Z1555~
said mount; a valve member fixed to the other end of said swivel in alignment with said valvé seat; spring means in said chamber on the other side of said diaphragm for biasing said diaphragm toward said valve seat to thereby seat said valve against said seat and keep fluid from flowing from said inlet passage through said chamber into said bypass passage; and means for admitting fluid at a second pressure into said other side o~ said diaphragm at a pressure which is different from, and lower than~ that of the fluid admitted into said one side of said diaphragm through said fluid inlet passage whereby, when the differential in the forces attributable to said fluid pressures exceeds the force exerted by said biasing means, said valve will open and fluid will be discharged through said bypass passage at a rate that will keep the pressure on the fluid discharged through said outlet passage essentially constant.
Other important features and advantages of the invention will become apparent from the appended claims and the detailed description in conjunction with the : accompanying drawingsO
BRIEF DESCRIPTION OF THE DRAWINGS
FigO lA and lB, taken together, constitute a partially sectioned side view of one gas turbine engine which may be equipped with a fuel flow scheduling controller;
Fig. 2 is a schematic view of a gas turbine engine liquid fuel supply system;
~, ~
~a2~
Figs. 3-6 are schematic views of a differen-tial linkage assembly employed in the flow controller of Fig. 2 showing how the assembly functions at different stages in the operating cycle of the turbine engine.
Fig. 7 is a section through a compressor dis-charye pressure responsive actuator employed in the con-troller;
Fig. 8 is an elevation of an embodiment of a fuelvalve assembly employed in the controller with part of the casing of the assembly being broken away to show a fuel flow metering valve incorporated in the assembly;
Fig. 9 is a section through the fuel valve assembly showing a differential pressure valve incor-porated therein;
Fig. 10 is a partial section through the assembly showing a two-stage relief valve incorporated therein;
Fig. 11 is a similar view showing a pressurizing valve incorporated in the assembly~
Fig. 12 is a graph showing the fuel schedule maintained by the controller;
Fig. 13 is a plan view of the differential link-age mechanism shown diagrammatically in Figs. 3-6; part of the assembly's dust cover has been broken away to show the nternal, working parts of the assembly; and Fig. 14 is a graph included to illustrate the operation of the mechanical differential linkage.
BEST MODE FOR CARRYING OUT THE INVENTION
Referrlng to the drawings, Fig. lA and lB depict a two-shaft, gas turbine engine 16 equipped with a fuel .
s~o supply system 18 which includes a fuel scheduling con-- troller 30.
Engine 16 has, among its major components, a fifteen-stage axial flow compressor 22 with a radialaxial inlet 24, inlet guide vanes 25, stators 28, and a fifteen-stage rotor 30. Th~ inlet guide vanes 26 and stators 28 are supported from the compressor housing-32 with the guide vanes and stators 23-1 through 28-5 of the first five stages being pivotally mounted so that they can be adjusted to control the flow of air through the compressor.
Each of the fifteen stages of the compressor rotor 30 consists of a disc 34 with radially extending blades 36 fixed to the periphery of the disc. The stages are integrated into a unitary structure as by electron beam welding.
The high pressure air discharged from compressor 22 flows through a diverging diffuser 38 and an enlarged dump plenum 40 to an annular combustor 42 supported in an insulated combustor case 44.
Combustor 42, ~hich is of the annula-r type, includes inner and outer liners 46 and 48 concentric with the axial centerline S0 of the engine and an annular com-bustor dome 52 spanning the gap between the liners at the forward or upstream end of the combustor.
Injectors 56 slidably mounted in dome 52 at generally equidistantly spaced intervals therearound dis-charge Euel into the annular combustion zone 58 betweeninner and outer liners 46 and 48. The fuel flows from fuel supply system 18 to injectors 56 through holders 60 which extend outwardly from combustor don~e 52 through com-bustor case 44.
~2~ S~
rhe compressor discharge air heated by combustor ~ and the combustion products generated in the combustor are expanded through a two-stage gas producer turbine 62 and then through a two-stage power turbine 64. The tur-bines are rotatably supported in a nozzle case 66 mounted in an ann~lar t~rbine housing or case 67.
The gas producer turbine 62 has a two-stage rotor 68 and sta~ionary, internally cooled, first and second stage nozzles 70 and 72. First stage nozzles 70 are assembled into an annual array or ring as are the second stage nozzles 72.
The two stages 78 and 86 of the gas producer tur-bine rotor 68 are bolted to each other and, in cantilever fashion, to the rear end of a forwardly extending shaft ~6. Shaft 96 is coupled through rear compressor hub 98 to compressor rotor 30, thereby drive-connecting gas producer turbine 62 to the compressor.
T~e compressor and qas producer turbines are rotatably supported by a thrust bearing 100 and by tapered land bearings 102, 104, and 106. Bearings 100 and 102 engage the ~ront compressor hub 108 which is bolted to 20 rotor 30 and is drive-connected to an accessory drive 110.
Power turbine 64 includes first and second stage nozzles 112 and 114, also supported from nozzle case 66, and a rotor 116 having a first, bladed stage 118 and a second, bladed stage 120. The first and second stage 25 nozzles 112 and 114 of power turbine 6~ are assembled into stationary annular arrays or rings.
Power turbine rotor stages 118 and 120 are bolted together for concomitant rotation. Rotor 116 is bolted to ~21SS5(~
a power turbine shaft assembly 128 rotatably supported by tapered land bearings 130 and 132 and a thrust bearing 134. rrhe shaft assembly is connected through a coupling 136 to an output shaft assembly 138 which furnishes the input for a generator, booster compressor, mechanical drive, or other driven unit (not shown).
The final major component of turbine engine 16 shown in Figs. lA and lB is an exhaust duct 140 for the gases discharged from power turbine 64.
Referring to Fig. 2, the fuel supply system 18 includes among its major components, an electronic control module 142 which converts a number of inputs including the load upon gas turbine engine 16 into an electrical fuel demand signal and a hydroelectric actuator 143 which con-verts the electrical fuel demand signal into a mechanical input to fuel flow controller 20.
Also, fuel supply system 18 includes a pump 144 for effecting a flow of liquid fue~ from a liquid fuel supply 146~through a filter 148 first to liquid fuel flow controller 2~ and then to the fuel injectors 55 of gas turbine engine 16.
Components of fuel supply system 18 such as hydroelectric actuator 143, pump 144, and filter 148 can be of any suitable character and are readily available.
Consequently, they will not be described further.
Nor is it considered necessary to describe elec-tronic control module 142 in detail. Any suitable system generating an appropriate fuel demand signal can be employed in fuel supply system 18.
12~555~
The electronic control module 142 processes - signals representative of the load upon turbine engine 16, the speeds of power turbine 64 (NpT) and gas producer turbine 62 (NGp), the temperature of the hot gases supplied to the gas producer turbine (T5), and the power turbine temperature (TpT) into output signals which can be employed to control the adjustable inlet guide vanes 26 of turbine 16 (GV) and the turbine bleed valve which is not shown in the drawings (BV). The module 142 also fur nishes the ~uel demand signal which actuator 143 converts to a mechanical input to fuel supply system flow con-lO troller 20.
Referring still to ~ig. 2, the major components of fuel flow controller 20 include a PCD actuator 150 which produces a mechanical output indicative of the pres-sure of the air discharged from the compressor 22.
15 Another major component is a differential linkage assembly 152 which mechanically multiplies the variable inputs frolT
hydroelectric actuator 143 and PCD actuator 150. The output of the differential linkage assembly serves as an actuator for a fuel metering valve 154 through whicn the , 20 fuel flows on its way from the discharge side of pump 144 to the fuel injectors 56 of gas turbine engine 16.
Metering of the liquid fuel to gas turbine engine 16 by the product of the variable fuel demand and com-pressor discharge pressure inputs generated in 25 differential linkage assembly 152 permits the engine to be rapidly started up and operated under varying loads without the danger of compressor surge. It also allows the engine to be rapidly decelerated without flameout.
~2~SSS~
Referring still to Fig. 2, fuel flow controller 20 also includes a ~P regulator 156 to maintain a con-trolled pressure differential across metering valve 154, a pressurizing valve 158 which maintains sufficient back pressure to enable the 4P regulator to function properly, and a relief valve 160 which prevents excess pressure build-up in fluid supply system 18.
Referring to Figs. 3-6, differential linkage assembly 152 has been shown in a somewhat diagrammatic form to simplify explanation. In actual practice, the links of that assembly are con~igured and arranged in a slightly different manner and, in some cases drilled to lighten the components and to reduce dynamic loads.
Movable components are suppo~ted by precision ball bearings to insure accurate response of metering valve 154 to the inputs from hydroelectric actuator 143 and PCD
actuator 150, to prevent metering valve position errors ; which might be caused by reaction of friction loads on PCD actuator 150, and to otherwise insu~e a stable, closed loop operation. A cover will typically be provided to protect the moving components of the linkage assembly from dust and other conditions.
One actual preferred differen~ial linkage assembly is illustrated in Fig. 13, the dust cover being dentified in that figure by reference character 162.
Referring again to Figs. 3-6, differential link-age assembly includes a PCD actua~or input link 164, ahydroelectric actuator input link 165, and an output link of fuel lever 168 which controls metering valve 154.
These links are supported from a base 170 by pivots 172, 174, (see Figure 6), and 176, respectively.
_g_ ~2~S5S~
Also included in the assembly are three links - 178, 180, and 182 which multiply the displacements of links 164 and 166 and impart to output link 168 a movement having a magnitude indicative of the product of the two inputs. Links 178 and 180 are pivotally connected to each other by pivot member 184 and to input and output links 164 and 168, respectively, by pivot members 185 and 188.
Link 1~2 is connected to links 178 and 180 by pivot member 189 and to hydroelectric actuator input link 166 by pivot member 189.
Stops 19û and 192 on base 170 limit the travel of PCD actuator input link 164. Similarly, stops 194 and 195 mounted on base 170 limit the travel of hydroelectric actuator input link 166, and a stop 198 on input link 166 limits the travel of the displacement multiplying links 178, 180, and 182. The stops mechanically limit the range of operation of the liquid fuel controller and provide reference points for assembly and c(alibration of the differential linkage assembly.
As discussed above, differential linkage assembly 152 multiplies the variable fuel demand input from hydro-electric actuator 143 and the variable compressor dis-charge pressure input from PCD actuator 150. By virtue of the relationship between actuator 143 and electronic control module 142, the differential linkage assembly therefore has a fuel flow rate controlling output which:
(a) is responsive to the temperature, speed, and load inputs to the control module; and ~b) is trimmed by varia-tions in compressor discharge pressure.
The control philosophy embodied in differential `~ linkage assembly 152 is shown in Fig. 14 in which "Z"
represents the positions taken by hydroelectric actuator 143 as engine 16 is accelerated to full load (Z=l.0) and to various part loads down to Z=0.4.
There is a straight line relationship between the position of the actuatGr and the rate of fuel flow WF
(and differential linkage assembly). This is modified or trimmed by PCD. For example, if the engine is operating under full load (differential linkage assembly fuel lever at "A") and the PCD drops from pressure "5" to pressure "4", the linkage assembly will move the fuel lever to "B", decreasing the 10w of fuel to engine 16 (WF) from rate "C" to rate "D".
A reverse case is acceleration from cold start (fuel demand maximum and PCD near zero). In this case (see, also, Fig. 4), WF is initially held to rate "E", despite a maximum demand for fuel by actuator 142, and increased to rate "C" as PCD increases to pressure "A".
The foregoing is accomplished in differential linkage assembly 1~2 by variations in the distance between 20 pivot members 174 and 184~
Fig. 13 shows the actual differential linkage assembly 170 depicted diagrammatically in Figures 3-6.
Aside from the previously identified dissimilarities, assembly 170 differs in that it has pivotably connected links l99a and l99b for transmitting the movement of hydroelectric actuator 142 to fuel demand input link 166 of assembly 152. It also has a pivotably mounted crank l99c for transmitting the movement of the PCv actuator : 30 ss~
output to the trim input link 164 of the assembly. These links merely facilitate connections between the interior and exterior of the differential linkage assembly housing. They do not affect the above described operation of the linkage assembly.
Fig. 7 depicts in detail the PCD actuator 150 which includes a housing 200 made up of end wall and adjacent castings 202 and 204 spaced from end wall casting 206 by cylindrical housing member 208. The housing com-ponents are bolted or otherwise fastened together.
Clamped between castings 202 and 204 is a diaphragm 210 which divides the interior of housing 200 into two chambers 212 and 214.
Fixed to diaphragm 210 by a circular clamp 216 and located in housing chamber 212 is an actuator mount 218. Threaded into one end of this mount is an elongated 15 actuator 220 extending through housing cha~ber 212 and through a fitting 222 threaded into end wall casting 206 o~ housing 200 to the exterior of that housing.
Actuator mount 218, along with the actuator and diaphragm 210, is biased toward that end of housing 200 20 defined by casting 202 by a coil spring 224. That spring extends between a spring seat 226 threaded into fitting ~22 and a ~lange 228 integral with and located toward the end of actuator mount 218 into which actuator 220 is threaded.
'rhe actuator, actuator mount, and diaphragm are, in circu~stances related below, also biased in the same direction as they are by spring 224 (i.e., toward end wall casting 202) by a coil spring 230. This spring extends SS~;;(l between a boss 232 at the inner end of a plunger 234 and the clamp 216 b~ which actuator mount 218 is fixed to diaphragm 210.
As shown in Fig. 7, the plunger and spring 230 are located in a cavity 235 in actuator mount 218 with the plunger extending through a bore 236 in clamp 216.
The end of the plunger facing casting 202 and terminating in head 237 is biased toward a stop 238 by a coil spring 2~0. The spring extends between the head 237 of the plunger and clamping plate ~15.
Springs 230 and 240 support valve plunger 234 in bore 236. Clearance between the plunger and clamp 216 is provided to eliminate friction and any possibility of the plunger sticking.
Stop 238 is fixed to the inner end of a shaft 241 threaded into housing member 202. A control knob 242 lS fixed to shaft 241 outside housing 200 allows stop 238 to be rotated through an angle lappro~imately 360) deter-mined by the engag~ment of a lug 2~4 on the stop with a second stop 246 fixed to casting 202.
Rotation of control knob 2~2 displaces stop 238 toward and away from casting 202, altering the force or bias which spring 2~0 is able to exert on diaphragm 210 via clamping plate 216. This is used, in fuel flow controller 30, to calibrate the PCD actuator for the altitude at which turbine engine 16 is operated.
Fluid (compressor discharge air in system 18) is admitted to the second chamber 21~ in the housing 200 of PCD act~ator 150 through an inlet 247 in end wall cast-ing 202. As is apparent from Fig. 7, the force generated by this fluid opposes that generated by springs 224 and 230 and displaces actuator 220 a distance proportional to the fluid pressure. In controlling the flow of fuel to the injectors of a gas turbine engine, actuator 220 fur-nishes to a differential linkage assembly 152 a mechanical input having a magnitude proportional to the engine's dis-charge pressure.
Fig. 7 shows the actuator with its moveable com-ponents positioned as they are with minimum pressure in chamber 214. At this stage of operation, spring 240 is biasing the head 237 of plunger 234 against stop 238.
Springs 224 and 230 are acting in parallel to displace actuator mount 218 toward the end wall casting 202 of the regulator, and actuator 220 is in its most retracted position.
As the pressure of the fluid chamber 214 increases, the resulting force exerted on diaphragm 210 displaces it, together with actuator mount 218 and actua-tor 220, toward the opposite end wall 205 of the actuator against the resistance offered by springs 224 and 230. At a specified pressure level determined by the relative 20 dimensioning of springs 224, 230, and 250, spring 240 becomes fully extended; and the head 237 o~ plunger 234 moves out of contact with stop 238. At this juncture, spring 230 becomes ineffective; thereafter only spring 224 opposes the fluid generated force. In this stage of operation actuator 220 is displaced a greater distance for each unit increase (or decrease) in fluid pressure.
Thus, PCD actuator 150 has two spring rates - a higher one when the fluid pressure in chamber 214 is below 5S~
a specified level and a lower one when the fluid pressure ~ is above that level.
Spring seat 226 is threaded into insert 222, which is kept from moving relative to end wall casting 206 by a snap ring 246a. The spring seat is held against rotation land thereby kept from moving longitudinally in housing 200 and changing the spring rate) by a pin 246b extending from it into end wall casting 206.
However, displacement of the spring seat toward and away from end wall casting 206 to vary the bias exerted by spring 224 on actuator mount 218 and diaphragm 210 can be effected by rotating internally threaded insert 222 ~hich is slotted to facilitate this operation. By doing so, one can vary the fluid pressure at which the break in the spring rate occurs to facilitate the initial calibration of the PCD actuator and permit operation of the actuator to be matched to the particular type of load being driven by t~rbine engine 16. f The metering valve 154 operated by the output from differential linkage assembly 152 is incorporated in a fuel valve assembly 248 (Fig. 8) which includes a ~P
20 regulator 156 and relief valve 160 lsee Fig. 10).
Fuel valve assembly 248 includes a casing 244.
Housed in casing 249 are a valve seat 250 and a metering valve plunger or member 252. The valve seat is threaded into a casting 254 of casing 249.
Valve plunger 252 is supported for rectilinear movement in a guide 256 mounted in casting 254 and clamped in place by a casting 258 which is part of the valve assembly casing or housing 249.
~z~s~
The valve plunger is biased closed (i.e., toward _ valve seat 250) by a coil spring 260 extending between valve guide 256 and a spring seat 262 fixed to the inner end of the valve plunger by a snap ring 263 making the valve fail-safe and eliminating backlash.
The valve design eliminates the customary close fitting spool and sleeve assembly which minimizes hysteresis and enables the metering valve to handle con-taminated fuels because the close tolerances needed in a spool and sleeve assembly are eliminated.
Liquid fuel delivered by pump 144 (See Fig. 2) flows through the external fuel line 264 in which filter 148 is incorporated and then seriatim through an internal passage 266 in casting 258 and an internal passage 268 in valve seat 250. From there the fuel can flow to a chamber 270 in the casting and then through an internal discharge passage 272 to pressurizing valve 158.
Referring to both Figs. 8 and 13, the rate (WF) at which fuel can flow into internal discharge passage 272 is determined by the distance between plunger 252 and valve seat 250 because WF is a product of that distance and the circumference of the passage 268 in valve seat 250, and the latter is a constant. The spacing (typically thousandths of an inch) between the valve plunger and valve seat is, in turn, controlled by the angular position of the fuel valve lever 168 in differential linkage assembly 152 as was discussed above.
More specifically, the position of lever 168 determines the angular position of the pivot shaft 176 on ::~2~
which the lever is mounted and, consequentially, the angu-- lar position of an eccentric 278 non-rotatably fixed to that shaft. Rotation of the eccentric, in turn, displaces a roller 280 carried by the eccentric toward and away from valve seat 250. The roller is confined between lands 282 and 283 on valve plunger 252 with essentially zero clear-ance between the roller and plunger. Consequently, the anyular movement of eccentric 278 is accurately translated into linear movement of valve plunger 252 toward and away ~rom valve seat 250.
It is believed that the operation of the fuel metering valve will be apparent from the foregoing.
Briefly, however, during the start-up of engine 16, fuel pump 14g is actuated. ~uel accordingly flows through line 264 and filter 148 and then through the passage 265 in fuel valve assembly casing 249 into the internal passage 26~ in valve seat 250. The rate of fuel flow is regulated by fuel valve lever 16~ and eccentric 278 in the manner just described. As discussed above in conjunction with the operation of differential linkage assembly 152, the position of the fuel valve lever - and, ultimately, the flow of fuel through metering valve 154 - is a function of the load on turbine engine 16 and the other inputs to electronic controller 14 as well as the compressor dis-charge pressure of the engine.
Operation of metering valve 154 requires that the pressure drop across that valve be accurately controlled because fuel pressure can vary as engine 16 is accele-rated, especially from a start. The ~P regulator, in association with the pressurizing valve 158 incorporated ~215556~
in fuel valve assembly 248, autornatically compensates kor ~- such variations, guaranteeing repeatability of operation.
The ~P regulator is housed in part in the cast-ing 254 of the fuel valve assembly housing and partly in a chamber 290 defined by that casting, a cooperating casting 292, and a cover plate 294.
Housed in casting 254 and clamped between that casting and casting 292 is a diaphragm 296 carrying a spring seat 298 on one side and a valve mount 300 on the other. Also housed in casting 254 are a valve seat 302, which is threaded into the casting, and a valve plunger or member 304.
The second (306) of the two chambers into which the interior of casing 24g is divided by diaphragm 296 communicates with the external fuel line 254 on the upstream side of fuel valve assembly 248 via an internal 15 passage 310 in casting 254. From chamber 306 the fuel flows through internal passage 266~to metering valve 154.
The chamber 305 of ~P regulator 156 also com-municates by way of valve 304, valve seat 302, and an internal passage 316 in casting 254 with an external fuel bypass or return line 318. That line is connected to the fuel supply line on the upstream side of f~el pump 144.
Finally, the chamber 290 in ~P regulator casting 292 on the opposite side of diaphragm 295 from chamber 305 is connerted through internal passage 322 Isee Figure 2) to the Euel passage 272 on the downstream side of metering valve 154. Passage 322 furnishes communication between the cavity 270 in casting 254 on the downstream side of ;iS5~
metering valve seat 250 and the chamber 290 in ~P regula-tor 156. This makes the pressure on the "downstream" side of the metering valve available in chamber 290.
Referring to Fig. 9, the valve plunger 304 is located in a fuel cavity 326 into whlch fuel flows from external fuel line 264 and internal fuel passage 327.
From cavity 326 the fuel flows around the valve plunger through internal passage 266 to metering valve 154. Valve plunger 304 has a conical tip 328 which is engaged with the valve seat 302 to keep fuel from flowing through the internal passage 330 in that seat into fuel bypass passage 316 when the valve is closed.
Plunger 304 is fixed to a valve stem 332 termina-ting, at its end opposite the valve member, in a swivel 334. The latter is supported in a generally hemispherical cavity 336 in valve mount 300 and retained therein by spring seat 298. This insures good alignment between valve member or plunger 304 and valve seat 302 without side loading to provide good seating and eliminate the friction and sticking common to heretofore available designs.
Valve plunger 304 is biased to the closed posi-tion illustrated in Fig. 9 by a coil spring 338. Spring 338 is centered on seat 298 and extends fronn that seat to the cover 394 of the fuel valve assembly casing 249.
As long as the pressure difference across meter-ing valve 154 is at the specified level, the force exerted by spring 338 and that generated by the fuel in chamber 290 at the pressure on the downstream side of metering valve 154 on diaphragm 296 will keep the valve closed.
5~S~
However, should this specified pressure differ-ence be exceeded, the force generated by the fuel at the pressure on the upstream side of the metering valve and applied to the diaphragm will unseat the valve, allowing fuel to flow through the passage 330 in valve seat 302, the passage 316 in casting 254, and external fuel return or bypass line 318 to the upstream side of fuel pump 144.
This bypassing of excess fuel reduces the pressure on the upstream side of the metering valve, bringing the pressure drop across that valve back to the specified level.
Referring to Figs. 2, 8, and 11, it was pointed out above that proper functioning of the ~P regulator ; requires that a minimum back pressure be maintained on the downstream side of metering valve 154 and that this is accomplished by pressurizing valve 158 which keeps the fuel pressure in the entire fuel supply system 18 from lS going below the minimum required for proper operation of the system. It will be remembered(that the pressurizing valve is incorporated in fuel valve assembly 248 (Fig~ 8) and occupies the upper right-hand portion of casting 254.
Referring specificially to Fig. 11, pressurizing 20 valve 158 includes a valve plunger 340 loosely fitted and rectilinearly moveable in a bore 342 formed in casting 254. 0-rings 344 and seals 346 keep fuel from leaking past the valve member. Valve member 340 is biased by a coil spring 348 toward the closed position, in which it : 25 blocks the flow of fuel in metering valve outlet passage 272. The spring is seated in the interior 350 of the valve plunger and extends from the head 352 of the latter into a spring seat 354 threaded into casting 254.
-~0- .
~Z~S55~
Like the other valves described above, and here-- inafter, pressurizing valve 158 has the advantage that close manufacturing tolerances are not required, making it capable of handling dirty fuel without sticking.
Pressurizing valve 158 is viscous damped for better stability by connecting the interior 350 of valve plunger 340 through the internal passage 356 in which orifice 358 is formed, internal passage 360, and a third internal passage in casting 254 (not shown) to internal fuel return passage 316. This also provides a fuel return pressure reference for operation o~ the pressurizing valve.
Especially upon start-up and deceleration or shut-down of engine 16, the back pressure in the fuel line 361 through which the fuel is delivered to the injectors 56 of engine 16 (see Fig. 2) is apt to be too low for proper operation of ~P regulator 156. In this case pres-surizing valve 158 remains closed until the pressure on the downstream side of metering va~ve 154 and in the fuel passage 272 and, ~onsequently, the pressure in chamber 290 of the ~P regulator, becomes high enough for proper operation lif the pressure drops below the specified mini-mum, and the ~P regulator consequently fails to operateproperly, unwanted and deleterious changes in the schedul-ing of fuel to the injectors 56 of engine 16 may occur).
With tne specified minimum pressure in ~P regu-lator chamber 290, the force generated by the fuel flowing into the passage 272 on the downstream side of metering valve 15~ land therefore in ~P regulator chamber 290) ~-becomes high enough to ~nseat and open valve 158, allowing the metered fuel to flow from the pressurizing valve inlet ~2~5~50 passage 272 to the valve outlet passage 351 and then to -- the fuel injectors 56 of gas turbine engine 16.
The pressure in chamber 306 and f-~el cavity 325 - is then sufficient for the excess fuel to flo~ from fuel cavity 326 into internal bypass passage 315 and, from there, into external fuel line 318, bypassing fuel from the upstream side of metering valve 154 at a rate which is determined by the spacing between ~ P valve plunger 304 and valve seat 302 to keep the pressure drop across seat 250 of metering valve 154 constant.
Should the pressure in passage 272 drop below that required by the ~P regulator during the operation of engine 15, spring 348 will close the pressurizing valve to the extent necessary to restore the pressure in passage 272 to the specified level.
Referring to Figs. 2, 9, and 10, the last o~ the above-discussed major components of fuel supply system 18 is relief valve assembly 160 which occupies the lower right-hand portion of fuel valve assembly casting 254.
The relief valve assembly includes a main relief valve 352 and a pilot valve 364. This two-stage arrange-ment affords superior regulation and is more stable than the usual poppet type of relief valve.
Referring to Fig. 10, the main relief valve is similar to the pressurizing valve and has the same advan-tages. Fuel reaches the valve via internal fuel passage 366 in casting 254 which communicates with external fuel return line 264 tsee Fig. 2). With the valve open the fuel can flow through an internal return passage 368 which communicates via return passsage 316 to previously dis-cussed external fuel return line 318 (Figs. 9 and 2).
:~2:~55~
Absent excessive fuel pressure in fuel supply ~~ system 18, the fuel is kept from flowing from internal passage 366 into return passage 368 by valve plunger 370.
Plunger 370 is loosely fitted and rectilinearly moveable in a bore 372 in casting 254 and is sealed against leakage by an O~ring 374 and seal 376.
Plunger 370 is biased to the closed position shown in Fig. 10 by a coil spring 378. The spring is seated in a cavity 380 in plunger 370 and extends from the head 382 of the plunger to a spring seat 384 threaded into casting 254.
~lain relief valve 362 differs from pressurizing valve 158 in that there is an orifice 386 in the head 382 - of the valve. Fuel can flow at a controlled rate from passage 366 through that orifice, the interior 380 of the valve, and a passage 388 in casting 254 to pilot valve 364.
As shown in Fig. 10, the pilot valve incl~des a valve seat 390 threaded into casting 254. Fuel can flow from passage 388, an orificed passage 392 in the valve seat, and an internal passage 394 in casting 254 into ~ internal fuel return ~assage 368 when the valve is opened.
:~ 20 Absent excessive pressure in fuel supply system18, however, such flow of fuel is prevented by valve member 396 which is biased against the valve seat by a coil spring 398. The latter extends between an annular ièdge 400 on the valve member and a spring seat 402 which is threaded into valve seat 390.
Should the pressure in fluid supply system 1~
exceed the specified maximum, the pressure of that fluid, which can flow from p~ssage 366 into valve seat 390 and ~%~SS50 against valve plunger 396 will exert sufficient force on the plunger to unseat it.
This allows the fluid in main relief valve plunger 370 to drain into return passage 36~, creating a pressure differential across the head 382 of the main relieE valve. This overcomes the valve closing force exerted by main relief valve spring 378 which then opens, returning fuel to the upstream side of fuel delivery pump 144 or keep components of fuel supply system 1~ from being damaged.
Operation of fuel supply system 18 will be clear by reference to Figs. 3-6 and 12.
~ s shown in Fig. 12, flow of fuel to engine 16 is carefully scheduled to keep the ratio of fuel flow WF to compressoe discharge pressure PCD between limits at which flameout would occur on one hand and surge on the other. As discussed above, this is accomplished by multi-plying fuel demand and compressor discharge inputs in differential lin~age mechanism 152 and employing the pro-duct of those two inputs to regulate the position of fuel metering valve 154.
` 20 Fig. 3 shows the differential linkage mechanism with its various links positioned as they are when engine 16 is started up. As the engine is brought up to the speed at which combustion can be effected and the engine operated under its own power, a maximum fuel demand signal is transmitted to the differential linkage mechanism as shown in Fig. 4; and metering valve 154 is opened, allow-ing fuel to be pumped to the injectors 56 of engine 16.
~155~ii0 However, as th~ compressor discharge pressure is rela-tively low in this stage of operation and as the PCD
actuator lS0 is operating under the higher of its two spring rates, the product of the two inputs is relatively low as indicated in Yigure 4; and fuel is supplied to the engine at a relatively low rate which, as shown in Figure 12, avoids engine surge.
At the specified breakover point, discharge pres~
sure will will have increased considerably as shown in Fig. 12. The PCD actuator will have shifted to opera-tion at the lower of its two spring rates. Thus, as is apparent from Figs. 5 and 12, metering valve 154 is rapidly opened wider at this point to rapidly increase the rate of flow of fuel to injectors 56 and to then pro-gressively open the valve until fully open, producing maximum acceleration of the engine.
As full speed is reached, the fuel demand decreases as shown in Fig. 12 and the engine operates along the steady state curve.
As shown in Fig. ~, shutdown of the engine results in the fuel demand dropping to zero with a corres-ponding reduction in the fuel supply to engine 16 as shownin Figure 12. However, because the compressor discharge pressure is still high at this stage, only a partial reduction in the flow of fuel is permitted with this being followed with a more gradual reduction as the compressor discharge pressure drops. This prevents flameout.
The invention may be embodied in other specific forms without departing from the spirit or essential characteristics thereof. For example, a three dimensional cam or a mechanical link supported on a linearly displace-able pivot member may be employed to effect the multiplication of the compressor discharge and fuel demand variables. Therefore, the embodiments described above are to be considered in all respects as illustrative and not restrictive.
~5
Claims (6)
1. A differential pressure actuated regulator comprising:
a housing with a chamber therein; a diaphragm in and spanning said housing; a fluid inlet passage in said hous-ing and communicating with said chamber on one side of said diaphragm; a fluid outlet passage communicating with said chamber on said one side of said diaphragm and extending through said housing to its exterior; a bypass passage communicating with said chamber on said one side of said diaphragm; and extending through said housing to the exterior thereof; a valve seat in said bypass passage at the end thereof which communicates with said chamber; a mount fixed to said diaphragm; a valve swivel carried at one end by said mount; a valve member fixed to the other end of said swivel in alignment with said valve seat; spring means in said chamber on the other side of said diaphragm for biasing said diaphragm toward said valve seat to thereby seat said valve against said seat and keep fluid from flowing from said inlet passage through said chamber into said bypass passage; and means for admitting fluid at a second pressure into said other side of said diaphragm at a pressure which is different from, and lower than, that of the fluid admitted into said one side of said diaphragm through said fluid inlet passage whereby, when the differential in the forces attributable to said fluid pressures exceeds the force exerted by said biasing means, said valve will open and fluid will be discharged through said bypass passage at a rate that will keep the pressure on the fluid discharged through said outlet passage essentially constant.
a housing with a chamber therein; a diaphragm in and spanning said housing; a fluid inlet passage in said hous-ing and communicating with said chamber on one side of said diaphragm; a fluid outlet passage communicating with said chamber on said one side of said diaphragm and extending through said housing to its exterior; a bypass passage communicating with said chamber on said one side of said diaphragm; and extending through said housing to the exterior thereof; a valve seat in said bypass passage at the end thereof which communicates with said chamber; a mount fixed to said diaphragm; a valve swivel carried at one end by said mount; a valve member fixed to the other end of said swivel in alignment with said valve seat; spring means in said chamber on the other side of said diaphragm for biasing said diaphragm toward said valve seat to thereby seat said valve against said seat and keep fluid from flowing from said inlet passage through said chamber into said bypass passage; and means for admitting fluid at a second pressure into said other side of said diaphragm at a pressure which is different from, and lower than, that of the fluid admitted into said one side of said diaphragm through said fluid inlet passage whereby, when the differential in the forces attributable to said fluid pressures exceeds the force exerted by said biasing means, said valve will open and fluid will be discharged through said bypass passage at a rate that will keep the pressure on the fluid discharged through said outlet passage essentially constant.
2. A pressure responsive actuator mechanism which comprises: a housing; a diaphragm dividing the interior of said housing into first and second chambers; an actuator support located in said first chamber and carried by said diaphragm; an actuator fixed at one end thereof to said support, said actuator extending therefrom through one end of said housing to the exterior thereof; means biasing said actuator support and said diaphragm toward the opposite end of said housing; and means for admitting a fluid at a pressure which may vary into said second chamber to impose on said diaphragm a force opposing that exerted by said biasing means whereby, as the fluid pressure varies, said actuator member will be shifted in position relative to said housing.
3. A pressure responsive actuator mechanism as defined in claim 2 which has adjusting means for altering the force exerted on said actuator support and said diaphragm by said biasing means, said biasing means extending through said housing toward said one end thereof and being seated on said adjusting means and said adjusting means being displaceable toward and away from said one end of said housing.
4. A pressure responsive actuator mechanism as defined in claim 2 which also includes a component supported by and protruding from said actuator support and rectilinearly displaceable relative thereto along a path extending in the same direction as the path of movement of said actuator; a stop fixedly positionable relative to said opposite end of said housing; a second biasing means extending between said actuator support and a protruding portion of said component and biasing the protruding end of the latter against said stop; and a third biasing means disposed in said actuator member support and extending between said support and said component for exertinq on said support and said diaphragm a force acting in the same direction as said first biasing means.
5. A pressure responsive actuator mechanism as defined in claim 2 which includes means for displacing said stop toward and away from said opposite end of said housing to thereby alter the force exerted by said second biasing means.
6. A pressure responsive actuator mechanism as defined in claim 2 wherein the limits of extension of said first and third biasing means are so related that, when the fluid pressure in said second chamber is above a specified level, one of said biasing means will be fully extended, whereby only the other of said biasing means will oppose the fluid generated force, thereby providing at pressures above and below said specified level different ratios of fluid pressure to actuator movement.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CA000501747A CA1215550A (en) | 1982-04-30 | 1986-02-12 | Pressure responsive actuator mechanism |
Applications Claiming Priority (4)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| US06/373,805 US4506503A (en) | 1982-04-30 | 1982-04-30 | Gas turbine engine fuel controller |
| US373,805 | 1982-04-30 | ||
| CA000421309A CA1204601A (en) | 1982-04-30 | 1983-02-10 | Gas turbine engine fuel controller |
| CA000501747A CA1215550A (en) | 1982-04-30 | 1986-02-12 | Pressure responsive actuator mechanism |
Related Parent Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA000421309A Division CA1204601A (en) | 1982-04-30 | 1983-02-10 | Gas turbine engine fuel controller |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA1215550A true CA1215550A (en) | 1986-12-23 |
Family
ID=25669936
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA000501747A Expired CA1215550A (en) | 1982-04-30 | 1986-02-12 | Pressure responsive actuator mechanism |
Country Status (1)
| Country | Link |
|---|---|
| CA (1) | CA1215550A (en) |
-
1986
- 1986-02-12 CA CA000501747A patent/CA1215550A/en not_active Expired
Similar Documents
| Publication | Publication Date | Title |
|---|---|---|
| US5996331A (en) | Passive turbine coolant regulator responsive to engine load | |
| US7096658B2 (en) | Centrifugal pump fuel system and method for gas turbine engine | |
| US5339636A (en) | Fuel splitter valve assembly for gas turbine | |
| CA1130888A (en) | Exhaust nozzle control and core engine fuel control for turbofan engine | |
| CA1174062A (en) | Manually operated metering valve for a fuel control | |
| CA1204601A (en) | Gas turbine engine fuel controller | |
| WO1989001092A1 (en) | Overspeed governor for an electronic controlled fuel system | |
| US3939649A (en) | Fuel control | |
| US4078378A (en) | Automotive gas turbine fuel control | |
| CA1177144A (en) | Fuel control apparatus for a gas turbine engine | |
| US3899879A (en) | Turbine engine fuel control | |
| CA1215550A (en) | Pressure responsive actuator mechanism | |
| US3712055A (en) | Fuel control | |
| EP0820558B1 (en) | Overspeed governor control system | |
| US4074521A (en) | Fuel control system for a gas turbine engine | |
| US3808797A (en) | Fuel systems for aircraft gas turbine engines | |
| US4417440A (en) | Fuel control apparatus for a gas turbine engine | |
| CA1217064A (en) | Power lever apparatus for a turbine engine | |
| US3579992A (en) | Power plant variable geometry control | |
| US3095702A (en) | Afterburner fuel control | |
| US3236047A (en) | Turbine engine fuel controls | |
| US3062005A (en) | Coordinated variable area nozzle and reheat fuel control for a gas turbine engine | |
| US3138926A (en) | Normal and acceleration fuel control for a gas turbine engine having a variable dropcontrol | |
| GB2119858A (en) | Means for controlling air scavenge pressure in the bearing compartment of gas turbines | |
| WO1993015322A1 (en) | Linear actuator for a bleed valve |
Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| MKEX | Expiry |