US3128822A - tyler - Google Patents

Description

April 14, 1964 s. R. TYLER v 3,128,822

HYDRAULIC SUPPLY SYSTEMS Filed Jan. 4, 1961 4.sheetssheet 1 ,s fg J6 z'i l: la Y /90 l SCHEDULING THROTTLE BOOST A mil coNTRol. VALVE PUMP T ,8./9 23 2/ |2/a 122 o F/G /lG s lill J5 F/G. 4

SOHEDULING CONTROL NOZZLE INVENTO TTORNEY.

April 14, 1964 s. R. TYLER 3,128,822

HYDRAULIC SUPPLY SYSTEMS Filed Jan. 4, 1961 4 Sheets-Sheet 2 INveNToR @MIMI wrm A-r-roRNEY S April 14, 1964 S. R. TYLER HYDRAULIC SUPPLY SYSTEMS Filed Jan. 4, 1961 4 Sheets-Sheet 5 SIMPLEX NOZZLE ATTORNEY April 14, 1964 s, R. TYLER HYDRAULIC SUPPLY SYSTEMS Filed Jan. 4, 1961 4 Sheets-Sheet 4 INVENTOR. A/VLEY I?. 7745/? A fr0/@V6 V5' Sha BY (n, l f "'1 Y l United States Patent O 3,l28,822 HYDRAULEC SUPPLY SYSTEMS Stanley R. Tyler, 4d Jitnxrnside Road, Cheltenham, England Filed Jan. li, 1961, Ser. No. 80,967 14 Claims. (Cl. 15S-36.3)

This invention relates to variable pressure liquid supply systems and while it applies generally to pressure liquid systems it has particular application in connection with fuel supply systems for burners of the spray or vaporizing types as used in gas turbines, in afterburners, or in furnaces or the like for supplying atomized fuel at an accurately controlled rate. Reference will be made hereinafter to the fuel supply to the spill burners of a gas turbine engine, but it is to be understood that this is merely by way of example, and no limitation to such a use nor to this specific type of burner is to be implied. This application is a continuation-in-part of my earlier but copending application Serial No. 676,637, filed August 6, 1957, and now abandoned.

The fuels employed for gas turbines need to be pumped at comparatively high pressure to the spray burners and for this purpose positive displacement pumps are usually used. Since these fuels have very poor lubricating properties difficulties can arise in these pumps such as mechanical seizure or wear if the fuel is dirt contaminated. The centrifugal pump is the most desirable form of pump for supplying fuel under pressure to spray burners, since the parts of the pump to which the fuel has access do not include any rubbing surfaces. However, under restricted output conditions fuel in a centrifugal pump becomes heated unduly due to the mechanical energy dissipated in it and one object of the present invention is to provide a fuel supply system using a centrifugal pump arranged to operate with a vapor core, and in which fuel in the pump is not unnecessarily heated under restricted output conditions. This arrangement is particularly useful for supplying reheat burners or afterburners in a gas turbine engine where the pressure falls substantially with reduced fuel flow. The improved eciency of a vapor core pump at low pressures is then better realized. Also the pump can be run entirely empty of fuel thus eliminating heat rise in the fuel during non-reheat operation.

Spill spray nozzles for burners such as those mentioned above are in common use which operate by supplying liquid fuel tangentially into a swirl chamber and from which that part of the fuel which is to be burned escapes as a spray and the remainder returns to a lower pressure region through spill orifices. These nozzles have the advantage over other spray nozzles of giving adequate atomization over a very large range of spray flow rates. However, when it is necessary to control accurately the rate of spray flow over a very Wide flow range with spill nozzles, it has heretofore been considered necessary to employ two pumps, one of which (the supply pump) delivers metered fuel to a closed circuit operating at high pressure, which includes the burner and the other pump which is normally known as the circulating pump. The rst or supply pump supplies fuel in the quantity delivered at the burner nozzle. The circulating pump is difficult to construct for successful operation since both inlet and outlet operate at high pressures and the mechanical drive to the pump must have a rotary seal operating at either inlet or outlet pressure, being in either instance a high pressure rotary seal. Alternatively, leakage must be tolerated in order to reduce the pressure to a level acceptable to the seal. A further object of the present invention is to provide a centrifugal pump for use with spill spray nozzles which combines both supply and circulating pump features into one pump and in which the high 3,128,822 Patented Apr. ld, 1964 ICC pressure rotary seal problem mentioned above does not exist, notwithstanding that a simple low pressure seal is required.

A still further object of the invention is to provide a simple centrifugal Variable pressure liquid supply pump operable from any prime mover or other power source. In accordance with the broad aspect of the present invention a variable pressure liquid supply system comprises a centrifugal pump suitably driven by a prime mover or the like and liquid flow control means acting to control the rate of flow of liquid entering the pump whereby there is insufficient liquid in the pump to iill it, and the contained liquid forms an annulus around the pump rotor, leaving a central hollow core, so that the output pressure is dependent on the radial depth of the annulus and the output rate of flow always tends to be the same as the controlled input ow.

Another object of the invention is to provide a fuel supply system of the character described (vapor core) which will enable a wide range of flow rates to the burner to be attained accurately, without instability.

In accordance with a further aspect of the present invention a fuel supply system for burner nozzles comprises a centrifugal pump to supply liquid fuel to one or more spray nozzles combined with liquid fuel flow control means controlling entry of liquid fuel into the centrifugal pump so that a hollow core in the liquid fuel exists around the centre of the pump rotor whereby the generated pressure in the circulating system including the spray nozzles is dependent on the radial depth which the liquid assumes around the rotor, this radial depth varying to generate pressure at the pump outlet to cause liquid fuel which enters the pump to be pressurized to the extent that it will tend to leave the pump at the rate at which it enters. It will be seen that if, for example, a higher resistance to flow occurs in the spray nozzle then temporarily the flow rate to the burner will drop allowing the radial depth of fuel in the pump rotor to increase thus increasing the pressure to overcome the nozzle resistance so that the rate of flow to the nozzle is that of liquid entering the pump. The converse also applies. In arranging a centrifugal pump to operate in this manner it will either be driven more quickly than a normal centrifugal pump to supply such a spray nozzle or, alternatively, it will be of a larger diameter. Where a centrifugal pump is designed to operate full of liquid at a maximum flow rate and pressure, it is possible without alteration of the size of the pump or its driving speed to control the pump in accordance with the invention to obtain reduced flow rates. Further, in accordance with the invention, where the spray nozzle is of the spill type (although the invention is also applicable to simplex type burner nozzles and other fixed orifice devices) the high pressure supply to the nozzle is taken from the normal pump output at the circumference of the rotor via a diffuser while the spill return flow from the nozzle is fed through an annulus in the pump casing to a point upstream of the pump outlet, that is to say, radially inwardly of the outlet, at which point the pressure is less than the pressure at said outlet. Since the pressure in the liquid inthe pump increases with increase in radial depth it will be seen that the spill flow is thereby allowed to enter the pump rotor at a lower pressure zone from where it will move outwardly of the rotor and be pressurized to the highest output pressure, without disturbing the hollow core. Thus, metered fuel fed to the inlet of such a centrifugal pump will all ow from the spill spray nozzle as spray at the rate at which it enters the pump. Where the pump and nozzle are used in conjunction with a gas turbine engine and the pump is driven directly by the engine, it will be seen that the pressure difference in the centrifugal pump between the position of entry of the spill flow and the pump outlet is proportional to the square of the rotational speed. Also it is well-known that the difference between inlet and spill pressures of a spill spray nozzle can be made approximately proportional to the square of inlet ilow to the nozzle and from this it is clear that the nozzle inlet flow can be made to vary substantially proportionally to the rotational speed of the engine. In this way the centrifugal pump in accordance with the invention operates in a manner similar to that of a fixed displacement supply pump feeding into a separate circulating system as normally used with spill spray nozzles.

In order that the invention may be clearly understood various embodiments thereof will be described with reference to the accompanying drawings.

FIGURE 1 is a diagrammatic cross-section of a centrifugal pump and associated control elements for feeding spill burners for use with a gas turbine engine.

FIGURES 2 and 3 are respectively cross-section and front elevational views of the centrifugal pump shown in FIGURE 1.

FIGURE 4 is a digrammatic view of an alternative centrifugal pump arrangement for use with spill burners.

FIGURES 5 and 6 are two further embodiments of centrifugal pump arrangements for use with ordinary burners of the simplex type (wherein all fuel permitted to reach the burner is discharged therefrom and burned) in a gas turbine engine.

FIGURE 7 is an axial sectional View of a practical form of centrifugal pump such as might be used, according to FIGURE 5, in supplying fuel to a simplex type burner.

Referring now to FIGURE 1, the centrifugal pump rotor is shown diagrammatically at 10 and is rotatably driven by means of a shaft 11. The rotor is mounted in the chamber 8 of a casing 12 and, being arranged to supply a spill spray nozzle, the pump has adjacent to the rotor a pair of circular channels 13 and 14 disposed in one wall adjacent to the open sides of the rotor blades. Channel 13 is disposed adjacent to the periphery of the rotor and is connected to an output pipe 15 leading to the input connection 16 of a spill spray nozzle 17. An adjustable throttle valve 15a, may be included in pipe 15. The spill flow returning from the nozzle 17 passes through ,pipe 18 and enters the channel 14 of the pump at a position slightly radially inwards of the channel 13. Thereby the channel 14 is of lower pressure than the channel 13. The inlet 19 of the centrifugal pump is disposed at the centre thereof and fuel which enters the inlet 19 passes through a scheduling control 21 and a boost pump 22 from a supply tank. A pressurizing valve 19a in inlet 19, which in effect is a spring-loaded non-return valve, is biased to close against inflow of fuel from boost pump 22, hence insures that no fuel will enter except so much as is required to make up the quantity discharged at the nozzle 17, such quantity increasing the pressure drop across the valve 19a. That valve also prevents vapor from passing back into control 21. The boost pump delivers fuel from the tank at a low pressure to the scheduling control and may comprise a small centrifugal pump or a fixed displacement pump having a valved relief passage (not shown, but conventional) extending from its delivery back to its inlet. The term scheduling control is used to indicate generally any of the usual controls, manual or automatic, which regulate fuel iloW in accordance with operating conditions of the gas turbine. It will be in the nature of a throttling device to govern the rate of fuel ow in accordance with a parameter such as altitude, or in accordance with the desired or constant speed of running of the engine, or in accordance with increased or decreased rates of flow for acceleration or deceleration. All such devices are known, hence its exact nature, construction, and mode of operation, and the nature of external parts through which it effects control, will necessarily vary widely in accordance with the overall design of the system, and since the control per se is not a part of this invention, it is shown only diagrammatically. In any event it senses the rate of fuel ilow and effects control thereof or thereby, in accordance with operating conditions. The scheduling control as shown in FIGURE 1 controls a liquid llow adjusting means 21a responsive to the scheduling control to adjust fuel flow by throttling. The scheduling control and liquid flow adjusting means deliver the fuel at the determined rate through pipe 23 and past the pressurizing valve 19a to the inlet 19 of the pump. This valve 19a insures a buildup of pressure in pipe 23 superior to pressure at inlet 19 before the valve will open to admit fuel to the pump. The pump rotor is so designed and operated that if it were maintained full of fuel it would be capable of delivering a slightly greater quantity than the scheduling control is capable of delivering. As a result the pump is never completely filled with fuel during operation, since the liquid which enters the pump casing is centrifuged outwardly and forms an annulus around the casing, leaving a circular empty space or core at the centre which is normally lled with fuel vapor at very low pressure. This hollow core is not vented nor connected to any other region. The pressure in the channel 13 depends on the radial depth of the annulus and the speed of rotation of the pump. Delivered liquid is drawn off through pipe 15 and connection 16 to the spill burner 17. Spill liquid from the spill burner is delivered through pipe 18 to channel 14 which is disposed radially inwards of the channel 13 in a zone of lower pressure. It will be seen that spill liquid Ventering channel 14 is immediately centrifuged outwardly and eventually again passes through pipe 15 to the spill burner inlet. In fact a circulating system operating at high pressure is formed by spill burner 17, pipes 15 and 18 and the outer part of the pump between channels 14 and 13. Liquid fed by scheduling control 21 into the pump, at the rate determined by the scheduling control, feeds into this circulating system by centrifugal action of the rotor, but at relatively low pressure, with the result that the rate of entry of fuel into the pump is equal to the rate of discharge from the spill burner. If, during operation of the engine, the burner should suddenly increase resistance to ow of fuel due, for example, to a partial blockage then the radial depth of fuel in the pump will build up until the pressure in pipe 15 is increased so that the flow is restored to the rate at which fuel is delivered to the pump. Thus fuel ow to the burners is substantially independent of conditions at the burners. Leakage of fuel along the drive shaft 11 will not occur during normal operation of the system, for while low pressure prevails at the center of the pump, the centrifugal action urges all contained fuel outwardly, and strongly resists its inward movement. A low pressure seal 11a is provided on shaft 11 to prevent leakage either of air along the shaft into the pump, or of low presure liquid from the pump along the shaft. If blockage should increase the radial depth of the fuel annulus temporarily, even to the point where it reaches the shaft, the pressure here still remains low, hence any suitable and simple low pressure seal 11a about the shaft, and not a high pressure seal, will suffice. The throttle valve 15a may be used to restrict flow in pipe 15 at high output flow rates from the pump in order to cause fuel to leave the pump at channel 14 rather than to enter it so that the nozzle 17 will eject as spray the fuel received from both pipes 15 and 18. The use of a throttle such as 15a to induce ow to a spill nozzle through its spill pipe is disclosed in my copending application Serial No. 842,762, filed September 28, 1959, and now abandoned in favor of a continuation thereof led September 7, 1962, and assigned Serial No. 222,107.

Referring now to FIGURES 2 and 3, details are shown of the pump appearing diagrammatically in FIGURE 1. It will be seen that the channel 14 is quite close to the peripheral channel 13 of the pump. Pressure in each is high, due to the centrifugal effect, although pressure in channel 13 is the higher. A diffuser is shown in dotted lines by the divergent fuel flow passage in delivery connection 24. Pressure is gained in the diffuser due to the speed reduction of fuel flowing through it. Connections to the channels 13 and 14 are shown in FIGURE 3 at 24 and 25 respectively. The rotor herein shown is an entirely conventional centrifugal rotor in which blades extend on one side of a backplate.

Referring now to FIGURE 4, an alternative arrangement for feeding a spill burner 26 is shown. Two centrifugal pumps having rotors 27 and 28 are provided mounted in an extended chamber 6 of a single case 29 and driven by a common shaft 31. The entry into the chamber 27 is by means of a port 32 in the casing feeding into the centre of the rotor. A pressurizing valve 30 is located in port 32. Delivery from rotor 27 is taken from peripheral channel 33 around the rotor. Delivery from this channel passes to passage 34 and into a scheduling control 35 of any known type. Fuel leaving the scheduling control passes through passage 36 to the entry port 37 at the centre of rotor 28. Peripheral passage 38 around rotor 28 receives liquid from rotor 28 and delivers it to pipe 39 which leads to the inlet connection of spill burner 26. Spill flow from spill burner 26 passes along the pipe 41 to the port 37. Fuel supply into the entry port 32 of rotor 27 is fed from a boost pump 42 whose output is controlled by a servo control 43 which forms the liquid llow adjusting means. It may operate either on a by-pass valve where the pump 42 is of fixed displacement or, alternatively, it may vary the displacement of pump 42 if it is a variable displacement pump. Pump 42 draws liquid directly from the storage tank. A pipe 44 interconnects the scheduling control 35 with the servo control 43, the arrangement being that ow of fuel through the scheduling control 35 acts through pipe 44 on the servo control to cause the required rate of ilow of fuel, as called for by the scheduling control, to be delivered by the pump 42. The pump rotor 28 is the circulating pump and forms a closed circulating system with pipes 39 and 41 and the spill burner 26. Fuel under pressure developed at the supply pump 27 and delivered by passage 34 is fed from the scheduling control 35 to this circulating system and eventually emerges from the spill nozzle as spray. The pump rotor 27 is made comparatively large in diameter such that the rate of flow of fuel of which boost pump 42 is capable does not fully lill the chamber of rotor 27, the fuel forming an annulus around the rotor to leave a central vapor core. Any leakage from the high pressure circulating pump rotor 28 will only enter the chamber of rotor 27, at lower pressure. Any tendency to leak along the shaft 31, outwardly of rotor 27, is readily countered hy a low pressure seal 31a about this shaft. It will be seen that in the same Way as described with reference to FIGURE l the pressure in channel 33 around the rotor 27 will vary in accordance with the radial depth of fuel and this pressure will vary automatically in order that the flow from channel 33 to pipe 34 is exactly in accordance with the scheduled fuel delivery into the rotor by boost pump 42. One essential difference from the arrangement of FIGURE l is that the scheduling control is now placed in the delivery line from the main centrifugal pump and acts to control the delivery of the boost pump by liquid flow adjusting means 43. The centrifugal pump rotor 28 acts in an entirely normal way and is completely primed with fuel.

It will be seen that in the arrangements of FIGURE l and FIGURE 4 the fact that the supply and spill lines to the spill burner, corresponding to the usual high pressure circulating system, are at high pressure does not necessitate the provision of rotary seals in the pump arrangements. A seal about the rotative shaft 11 or 31 normally is employed, but this is a low pressure seal, and it is still correct to say that the high pressures prevalent in the supply and spill lines do not require the presence of a rotary seal of the type necessary to contain such high pressures. There is no question of leakage of the 6 scheduled fuel flow by way of 15, 18 and burner 17 in the arrangement of FIGURE l, although it might take place to a limited extent in the arrangement of FIGURE 4.

In FIGURE 5 another system is diagrammatically shown for feeding into a spray nozzle 45 of the simplex type. In such a spray nozzle oil fuel delivered through the delivery pipe 57 passes immediately into a swirl charnber and out of the spray orice. While reference has been made to the nozzle 45 as being of the simplex type it may of course represent other kinds of spray nozzle or fixed orifice device in which only one pipe connection extends to the nozzle carrying fuel under pressure. The centrifugal pump rotor 46 is driven by a shaft 47 from the engine and is located in the chamber 4 of a casing 48 having a peripheral channel 49 and a central inlet 5l. The inlet 51 is controlled by a valve 52 from which a stem 53 extends to a servo piston and cylinder unit 54. A compression spring 55 tends to hold the valve on its seat. Liquid fuel is supplied to the inlet 51 through a pipe 56 from a low pressure source (not shown). The pipe 57 extends from the channel 49 to the burner 45 and in this pipe a scheduling control 58 is located. This scheduling control comprises a cylinder '75 within which a waisted piston valve member 76 is slidable under the inuence of a manually adjustable cam 77 and a return spring 78. A port 79 within the cylinder wall is in unrestricted communication with the waisted portion of valve member 76 and pipe 57 is connected to this port. A further port 81 is formed within the cylinder which is partially closed by the edge of land 82 of the valve member '75, adjustment of the valve member causing a throttle to be formed by port S1 which is a function of the setting of cam 77. The pipe 57 then extends from port 81 to the nozzle 45. Pipes 83 and S4 extend from positions downstream and upstream of the throttle port S1 and connect to either end of a cylinder 85 to cause movement of the included piston 86 against the load of a spring 87. A small diameter plunger 88 extends from piston 36 into bore 89 into which a pair of ports 91 and 92 open. A waisted portion 93 of plunger 8S controls fuel ow between ports 91 and 92. Port 91 is connected through restrictor 94 from pipe 84 Whilst port 92 is connected to low pressure through pipe 95. This control operates on the servo piston and cylinder 54 and control is exerted by means of a pipe 59 which extends from port 91. Fuel flowing through port 81 produces a pressure drop in accordance with ow rate, this pressure drop acting through pipes 83 and 84 on piston 86 in opposition to the load of spring S7. A small movement of piston 86 will move plunger 88 from a completely closed to a completely open position in its control of flow through ports 91 and 92 over which movement of the spring 87 is arranged to exert a substantially constant force. Thus at a particular pressure drop at port 81 the plunger 38 will adjust flow through ports 91 and 92 and will thus determine a pressure drop due to ow through restrictor 94. Pipe 59 connects the Variable pressure downstream of restrictor 94 to the servo piston and cylinder unit 54 and the opening given to valve 52 will result from the compression of spring 55 by pressure from pipe 59. If the ow rate through pipe 57 is too large the excess pressure drop at port 81 causes movement of piston 86 and plunger 88 to the left to permit greater ow through ports 91 and 92 and thus to cause a greater pressure drop at restrictor 94. This in turn will reduce pressure in pipe 59 causing spring 55 to move valve 52 to a more closed position to reduce entry of fuel into the pump to compensate the excess output flow. A reduced output flow would cause opening of valve 52 to increase input flow. Since the cam 77 determinues the throttling effect of port 81 it will be clearly apparent that the setting of this cam will accurately determine the entry of fuel ino the pump. Valve 52 will also perform the function of a pressurizing valve included in the previous embodiments. As in the previous embodiments, because of the excess delivery capacity of pump rotor 46 over the delivery capacity of pump 56, the liquid fuel in the pump forms an annulus leaving a central hollow core 4 and the adjustment of radial depth of the annular liquid will cause pressure to be developed at the pump outlet 57 appropriate to the operating conditions so as to cause fuel to leave the pump at the rate at which it enters.

A furher alternative arrangement is shown in FIGURE 6, the object again being to supply fuel to a burner 63 of the simplex type. The centrifugal pump is arranged substantially as described with respect to FIGURE and includes a rotor 64, and drive shaft 65, a casing 66, an output channel 67 disposed peripherally around the casing and an inlet 68. In this instance a pipe 69 leading from channel 67 is connected directly to the burner 63. In the inlet opening 68 of the centrifugal pump a springloaded valve 71 is provided which is spring-loaded to the closed position, the loading being in a direction such that the fuel supplied to the inlet through the pipe 72 must be at a higher pressure than fuel immediately inside the pump inlet 68. Fuel ow is controlled by means of a scheduling control 73 which exerts a servo control 102 on the delivery of boost pump 74, this latter either having a servo control by-pass or a servo-controlled variable displacement mechanism. Fuel is delivered by means of the control 73 and pump 74 to pipe 72 at the rate determined by engine operating conditions to which the control 73 is sensitive.

In all of the described embodiments reference has only been made to one spray nozzle, but it will be appreciated that in most instances a plurality of spray nozzles are fed from one pump, the nozzles all being parallel connected. Further, it will be seen that other types of burner may equally well be supplied by apparatus in accordance with the invention such for example as burners of the vaporizing type.

In the description above it has been assumed that the centrifugal pump casing is stationary and that the rotor therein rotates. Pumps are known wherein the casing rotates and the internal member is non-rotative, but scoops up liquid rotated by the casing. These still function as centrifugal pumps, and no restriction is intended herein in the broad aspects of the invention to one or the other style of centrifugal pump.

I claim as my invention:

1. In combination with a hydraulic load, a variable liquid supply system comprising drive means, relatively fixed and rotative elements arranged cooperatively about an axis of rotation to form a centrifugal pump having a chamber and an inlet and an outlet therefor disposed adjacent the axis and the periphery of the chamber, respectively, said pump being connected with the drive means to operate at a predetermined speed, means for supplying pressurized liquid to the pump, means defining inlet and delivery passages for the pump interconnecting the liquid supply means and the hydraulic load with the chamber inlet and outlet, respectively, valve means disposed adjacent the chamber inlet and operative to admit liquid to the chamber from the inlet passage in response to a pressure differential between such passage and the chamber, and control means operable to vary the rate of liquid flow in the inlet passage in accord with a demand which is variable over a predetermined range of flow rates having an upper limit of less than the delivery rate of the pump at said speed so that liquid flowthrough in the pump chamber assumes the form of an annulus in the relatively peripheral portion thereof, the relatively axial portion of the chamber being closed to atmosphere so that a hollow core is formed centrally of the liquid annulus to enable it to compensate for variation in the hydraulic load at a particular demand by adjusting its radial depth.

2. The combination according to claim l wherein the centrifugal pump has a fixed housing defining its chamber and a rotor mounted in the chamber to perform as the rotative element.

3. The combination according to claim 1 further comprising scheduling control means operable to sense the rate of liquid tiow in one of the passages and to so regulate the rate of such flow in the inlet passage, through the medium of the first-mentioned control means, that the latter rate is maintained in accord with the demand.

4. The combination according to claim 3 wherein the scheduling control means is operable to sense the rate of liquid flow in the inlet passage.

5. The combination according to claim 3 wherein the scheduling control means is operable to sense the rate of liquid flow in the delivery passage.

6. The combination according to claim 3 wherein the liquid supply means includes a boost pump disclosed in the inlet passage.

7. The combination according to claim 6 wherein the first-mentioned control means includes a throttle valve disposed in the inlet passage intermediate the boost pump and the centrifugal pump chamber inlet. 8. The combination according to claim 6 wherein the first-mentioned control means includes a servo control interconnecting the scheduling control means with the boost pump and operable to regulate the delivery from the latter pump.

9. The combination according to claim 3 wherein the valve means performs as the first-mentioned control means.

10. The combination according to claim 1 wherein the hydraulic load includes a spill nozzle having a chamber and an inlet, a discharge outlet, and a spill outlet therefor of which the inlet is interconnected with the pump chamber outlet by the delivery passage, said pump chamber also having a second inlet disposed radially intermediate the first-mentioned pump chamber inlet and the pump chamber outlet, and said passage dening means further deiining a spill return passage interconnecting the spill outlet of the spill nozzle chamber with said second pump chamber inlet.

11. The combination according to claim l0 wherein the delivery passage has a throttle valve therein.

l2. The combination according to claim l0 wherein the delivery passage has diffuser means therein to effect a speed reduction in the liquid flow therethrough.

13. The combination according to claim l wherein the hydraulic load includes a simplex nozzle having a chamber and an inlet and a discharge outlet therefor of which the inlet is interconnected With the pump chamber outlet by the delivery passage.

14. The combination according to claim l further comprising a second centrifugal pump in the delivery passage, said second centrifugal pump also being connected with the drive means to operate at said speed.

References Cited in the file of this patent UNITED STATES PATENTS 787,039 Harris Apr. ll, 1905 1,353,915 Kime Sept. 28, 1920 2,547,959 Miller Apr. 10, 1951 2,575,923 McMahan et al Nov. 20, 1951 2,658,330 Carey Nov. 10, 1953 2,673,604 Lawrence Mar. 30, 1954 2,713,244. l Chandler July 19, 1955 2,720,256 Pearson Oct. 1l, 1955 2,916,875 Morley et al Dec. l5, 1959 FOREIGN PATENTS 1,184,654 France Feb. 9, 1959 161,796 Great Britain Apr. 21, 1921

Claims (1)

1. IN COMBINATION WITH A HYDRAULIC LOAD, A VARIABLE LIQUID SUPPLY SYSTEM COMPRISING DRIVE MEANS, RELATIVELY FIXED AND ROTATIVE ELEMENTS ARRANGED COOPERATIVELY ABOUT AN AXIS OF ROTATION TO FORM A CENTRIFUGAL PUMP HAVING A CHAMBER AND AN INLET AND AN OUTLET THEREFOR DISPOSED ADJACENT THE AXIS AND THE PERIPHERY OF THE CHAMBER, RESPECTIVELY, SAID PUMP BEING CONNECTED WITH THE DRIVE MEANS TO OPERATE AT A PREDETERMINED SPEED, MEANS FOR SUPPLYING PRESSURIZED LIQUID TO THE PUMP, MEANS DEFINING INLET AND DELIVERY PASSAGES FOR THE PUMP INTERCONNECTING THE LIQUID SUPPLY MEANS AND THE HYDRAULIC LOAD WITH THE CHAMBER INLET AND OUTLET, RESPECTIVELY, VALVE MEANS DISPOSED ADJACENT THE CHAMBER INLET AND OPERATIVE TO ADMIT LIQUID TO THE CHAMBER FROM THE INLET PASSAGE IN RESPONSE TO A PRESSURE DIFFERENTIAL BETWEEN SUCH PASSAGE AND THE CHAMBER, AND CONTROL MEANS OPERABLE TO VARY THE RATE OF LIQUID FLOW IN THE INLET PASSAGE IN ACCORD WITH A DEMAND WHICH IS VARIABLE OVER A PREDETERMINED RANGE OF FLOW RATES HAVING AN UPPER LIMIT OF LESS THAN THE DELIVERY RATE OF THE PUMP AT SAID SPEED SO THAT LIQUID FLOWTHROUGH IN THE PUMP CHAMBER ASSUMES THE FORM OF AN ANNULUS IN THE RELATIVELY PERIPHERAL PORTION THEREOF, THE RELATIVELY AXIAL PORTION OF THE CHAMBER BEING CLOSED TO ATMOSPHERE SO THAT A HOLLOW CORE IS FORMED CENTRALLY OF THE LIQUID ANNULUS TO ENABLE IT TO COMPENSATE FOR VARIATION IN THE HYDRAULIC LOAD AT A PARTICULAR DEMAND BY ADJUSTING ITS RADIAL DEPTH.

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