GB2119018A - Compressor housing for a turbo-compressor - Google Patents

Abstract

The compressor housing if formed by die casting using an inner mould 292 and an outer mould 290 to produce the housing 24. Moulds 294 and 296 are used to form the compressor outlet. A wall insert 126 of the housing is also formed by die casting using moulds 320, 322 and is provided with protrusions 140 which engage apertures in the housing 24 to secure the insert in the housing. A throat portion 128 of the insert is provided with notches 326 which receive vanes for mounting a bearing support cylinder. <IMAGE>

Description

SPECIFICATION Turbocharger and adaptationsthereof The present invention relates to tu rbochargers for use on internal combustion engines and to turbojet and turbofan engines.

This application is a continuation-in-part of application Serial No. 759,773, filed January 14,1977, by the present applicant and entitled "Turbocharger." Turbocharging is a mens of greatly extending the powerrange andflexibility of internal combustion engines and has cometo be an accepted practice, and in many cases a necessity, for heavy duty diesel engines of 200 or more horsepower. Turbocharging is also used to maintain power at increasing altitude, for instance in aircraft engines.

The effective application of a turbochargerto an internal combustion engine will normally increase power output 50-100 percent and reduce full load specific fuel consumption (sfc) by 5-10 percent. The reduced specific fuel consumption is attributed to two items. First, at a given speed the engine internal friction remains relatively constant even though the power output is increased considerabiy. This results in an effective improvement in mechanical efficiency.

Second, if the efficiences of the turbocharger compo nines are high enough and the exhaust temperature of the engine on which the turbocharger is used is sufficiently high, there will result a positive pumping loop that addsto the netcycleoutput.

Turbocharging is normally thought of only as a means for increasing horsepower or decreasing full load specific fuel consumption. In the evolution of engines presently utilizing turbochargers, such as large heavy duty diesel engines, aircraft engines, racing engines and the like, this understanding has been adequate in that in the end application these engines are operated at or nearfull load for a large portion of the duty cycle. However, most applications do not require that the engine operate ator nearfull load for extended periods oftime. In fact, in most applications, the engine generally is operated below 50% power and in many applications the engine operates well below 20% power during most of its operation.Examples ofthese applications are engines used in automobiles, light and medium trucks, gener ator sets, compressors, tractors, construction equipment and the like.

Engines operating at these low power settings are very inefficient. In a diesel engine, this inefficiency is a result of thermal efficiency decay as combustion temperature decreases and because the internal friction remains relatively constant regardless of load.

In a gasoline engine, this inefficiency results from pumping loop loss increases with decreasing load and because the internal engine friction remains relatively constant regardless of load.

Thus, by utilizing a smaller "effective engine size", that is, an engine having a smaller displacement rate (the product of 1/2 displacementtimes engine speed for a four-cycle engine) by either a reduction in displacement, a reduction in operating speed or a combination of both, the part load fuel consumption may be improved. In gasoline engines, this improve ment results from reduced engine friction andre- duced pumping loop losses. In diesel engines, this improvement is a result of reduced engine friction and a higherthermal efficiency due to higher combustion temperatures.

In many applications, turbocharging may be used to permit the use of engines having smaller, "effective engine size." By turbocharging, it is relatively easy to obtain twice the naturally aspirated power per cubic inch ofdisplacementandinsomecasesthreetimes the power. However, attempts to turbocharge smaller engines have generally been unsuccessful.Thisfailure can generally be attributed to the present design ofturbochargerswhich are built around a journal bearing and a flat disctype thrust bearing. This type bearing system requires from one to three horsepower (depending upon the particularturbocharger and the speed required in the application) justto overcome friction.While this loss may be insignificant in applications whereturbines are required to develop in excess ofthirty horsepowerforthe compression process (typically engines of 200 or more horsepower), it becomes very significant when turbocharging engines having less than 100 horsepower. For exam pIe, where theturbochargerturbine power is 60-80 horsepower, a bearing friction loss of 2-3 horsepower is insignificant. However, in a smallerturbocharger where the turbine power is only 15 horsepower, a bearing friction loss of 2-3 horsepower, or more likely 4-5 horsepower due to the higher rpm at which the smallerturbochargers are operated, represents almost a third of the total turbine horsepower produced and is completely unacceptable.

The presently used bearing systems also require considerable radial and axial clearances to provide for oil flow and rotor stability. These clearances are translated into a relatively large clearance over the bladingofthecompressorandturbine rotorsthereby affecting the efficiency of both the compressor and turbine. For example, the journal and dise th rust bearings, commonly used in present dayturbochargers, may require a clearance of 0.015 inches between the turbine and compressor blades and the surround structure. Where the blade height is 1 inch, the clearance to blade height ratio is only 1-1/2 percent.

However, where a smaller turbocharger is desired, having blade heights of, for example, 0.2 inches, a clearance of 0.015 inches between the blades and surround structure amounts two 7-1/2 percent ofthe blade height. Therefore, where a 0.015 inch clearance is acceptable in largerturbocharger applications, it is completely unacceptable when smallerturbochargers are being designed. Therefore, in smallerturbochargers, this clearance becomes morn and more critical to the overall performance of the tu rbocharger and ultimately to the performance ofthe engine.

The bearing systems now being used in turbochargers are lubricated with engine oil, although most bearing failures are the result of contaminated engine oil or lack of engine oil pressure during starts. Where high speed journal bearings are used in a conventional turbocharger, continuous oil flow is inherently required to provide shaft stability as well as to carry away heat generated by viscous friction. Oil flow is also required to carry away the heat transferred into the bearing system from the adjacent turbine (which operates attemperatures as high as 1600 degrees F).

Even if antifriction ball bearings were substituted for the journal bearings in conventional turbochargers, a continuous oil flow would be required to carry away heat transferred from the tu rbine*Thus, while lubrication its a necessity forthe proper operation of present day turbochargers, lubrication also accounts for many of the failures. Further, continuous oil flow lubrication requires substantial plumbing and associated structurefor providing the lubricant to the bearings.

The present dayturbochargers have failed to efficiently control the flow of motive gases through the turbine. Presently, there are basicallytwo methods used for controlling power output ofthe turbine. The first of these methods is by careful sizing ofthe turbine and turbine nozzle area so that at maximum engine operating speed and load the desired boost pressure will not be exceeded. The disadvantage of this method isthat at low engine speedsthe available boost pressure is limited and the responseto demand is slow. The second method used for controlling the pressure th rough the turbine is the use of a "wastegate" in conjunction with a turbine nozzle sized to produce excessive turbine power at maximum engine speed and load.In this method, when the predetermined boost pressure is reached, the "wastegate" opens and bypasses a portion ofthe exhaust gases.

While this method increases the available boost at the lower engine speeds and provides improvements in response, it is quite inefficient in that the bypassed, high pressure exhaust gas is simplywasted atthe expense of increased engine back pressure. Additionally, at part load, when the turbocharger is essentially inoperative, the small nozzle area acts as a restriction to the exhaust and causes an increase in the pumping loop loss.

Therefore, a need has arisen for a turbocharger which can be efficiently operated to turbocharge both small and large internal combustion engines. The need isfor a turbocharger having a bearing system which eliminates the problems heretofore experienced by continuous engine oil lubricated bearings and makes the most efficient use of the motive gases for driving the turbocharger turbine. Further, the bearing assemblies supporting the rotation of the compressor and turbine muetfacilitatethe reduction ofthe required compressor and turbine rotor clearances.

The present design and method of production of prior art compressor housings fortu rbochargers have also presented a considerable problem.

Centrifugal flow compressors are one of the most widely used dynamic compressors in turbochargers.

In this compressortype, air or an air-fuel mixture enters the compressor inlet, is channeled to the compressor rotor and is accelerated to near sonic speeds at a right angle to the inlet flow path. Increase in air pressure is accomplished by reducing the velocity of the accelerated gases as discharged from the tip of the compressor rotor blades. This process, known as diffusion, is more efficiently achieved by slowing down the gases without turbulence so that a large percentage of the velocity energy is converted into pressure energy, raising the static pressure.

To facilitate this diffusion process, turbochargers employing centrifugal compressors have normally included a compressor rotor wall closelyfollowing the contour of the compressor rotor bladesfrom the blade leading edgetoits outertip. This compressorwall then extends pastthe outertipofthe blade to form one of two wa I Is of the diffuser, then terminates to provide a circumferential gap through which the compressed gases are channeled into a circumferential chamber leading to the intake manifoldoftheengine.Thiswall, facing the compressor rotor blades ariddosely contoured to the rotor blades then extending outward, uniformly decreases the velocity ofthe gases aftertlte gases leave the rotor blades and prior to their entry into the chamber leading to the engine. Thus, this wall structure greatly increases the static pressure generated by the compressor.

To form this structure, mostturbochargercom- pressor housings have been sand-castwith the compressor wall cast in one piece with the compressor outer surround housing. This has normally been accomplished by using a sand coretoform the circumferential chamber leading to the intake manifold ofthe engine. After casting, this sand core is dislodged to produce the chamber on the opposite side ofthewall from the compressor in which gases are channeled off of the tips of the compressor rotor.

Although die-casting ofthe compressor housing would be substantially less expensive and more accurate than sand-casting, die-casting of an optimum design has not been possible because ofthe inability to use die-cast molds to form the circumferential chamberwhichchannelsthecompressed gases to the intake manifold ofthe engine and atthe same time form the diffuserwall. Becausethe variable area chamber is necessarily largerthan the inlet gap through which gases are injected from the compressor rotor blades, die-casting an optimum design compressor housing has not been possible because of the inability to design moldsttlEt- ;;!ould form this passageway behind thewallfacing the bladesofthe compressor rotor.

Where die-cast compressor housings are used, the wall normally formed in sand-cast compressor hous- ings is merely deleted so thatthe molds may be broughttogetherand parted to form the casting.

However, without th is wall, gases accelerated by the compressor rotor are prematurely dumped from the diffuser into the circumferential chamber leading into the engine intake manifold. As a result, this arrangement realizes a substantially lower compressor efficiency and thus lower performance.

The present invention provides an improvedturbochargerfor use with internal combustion engines which overcomes many of the disadvantages heretofore experienced by prior artturbochargers. In accordance with one embodiment ofthe invention,the turbocharger includes a compressorfor supplying air to an engine and a turbinefor receiving driving exhaust gas from the engine. The turbine has a shaft integrally attached therewith with whichtheturbine rotates, and the compressor is fitted on the turbine shaftfor rotation with the turbine. The shaft is rotatively supported by bearing assemblies such that the compressor and turbine are overhung to one side ofthe bearing assemblies with the turbine separated from the bearing assemblies by the compressor.As a result of this arrangement, the bearing assemblies are removed from theturbine and the intense heat to which the turbine is subjected.

In accordance with another embodiment of the invention, antifriction ball bearing assemblies are used for supporting the compressor and turbine shaft.

The use of ball bearings is particularly critical in providing better control over the radial and axial movement ofthe turbine and compressor. This additional control over the movement of the rotating assembly allows reduced blade tip clearances, thereby appreciably improving compressor and turbine efficiency. The use of antifriction ball bearings further reduces the turbine work required, thereby improving the specific fuel consumption of the engine.

In one embodiment ofthe present invention, the bearing assemblies include a first and second inner raceway formed in the compressor and turbine shaft.

Afixed outer raceway corresponding to thefirst inner raceway is attached to the turbocharger housing and a plurality of bal Is are received between the fixed outer raceway andthefirst inner raceway. Asecond outer raceway ring is provided and is slidable relative to to the first outer raceway. This second outer raceway ring isslidablerelativetothehousing and a compression spring acts between the turbocharger housing and the ring to bias the ring away from the first outer raceway and in engagement with balls positioned between the ring and the second inner raceway.

Simultaneously therewith, the first outer raceway, attached to the tu rbocharger housing, is engaged against the balls between such raceway and thefirst inner raceway on the turbine and compressor shaft.

In accordance with another embodiment ofthe invention, a turbocharger includes a variable area turbine nozzle having a plurality of nozzle vanes spaced circumferentially about the turbine and in the inletthrough which exhaust gas from the engine is received into the turbine. A control linkage system is positioned intermediate ofthe compressor and tu rbine and controls the position of the nozzle vanes to varytheflowvelocity ofexhaustgases into the turbine. In one embodiment of the invention, the variable area turbine nozzle control linkage system includes a control ring having an outer ring movable angularly relative to an inner ring. An actuation lever is rotatable with each vane and has one end engaged in the outer ring of the control ring.A control rod is attached to one of the actuation levers and structure is provided for'selectively moving the control rod to pivot the actuation lever attached thereto. This in turn rotates the outer ring ofthe control ring thereby pivoting each ofthe actuation levers and the vanes attached thereto.

The use ofvariable area turbine nozzles provides more efficient control of peak boost pressure while allowing greater boost at lower engine speeds than now attainable. Variable area nozzles also improve transient response and provide a better part load specific fuel consumption in that the nozzle area can be opened to minimize engine back pressure during part load operation. Further, nozzles can be positioned to optimize emissions. Moreover, the present invention provides a very straightforward arrangement for conveniently controlling a plurality of nozzle vanes circumferentially spaced in the turbine inlet.The control linkage is conveniently positioned in the unused space intermediate of the compressor and turbine and in addition to providing for a very compact turbocharger "package", the control linkage space further provides an insulation shield between the heat of the turbine area and the bearing assembly supporting the compressor and turbine shaft.

In one embodimentofthe invention, the compressor rotor housing unit includes a compressor housing and a compressor wall insertfor mating with the compressor housing. The compressor housing includes a tubular shaped inlet port having a transverse wall attached to one end of the tubular port and extending outwardly therefrom. A circumferential chamber is formed in the wall opposite the inlet port and has a circumferential opening substantially in the plane of the wall. The circumferential chamber also has an opening which communicates with the intake manifold ofthe engine to which the turbocharger is attached.

The wall insert is formed having a tubular throat with a circular disc attached atone end ofthetubular throat transverse thereto. The disc has an aperture corresponding to the opening formed bythe tubular throat. Rivet-like protrusions are integrally formed and extend from the disc corresponding to and facing apertures formed in the wall of the compressor housing. These protrusions are circumferentially spaced around the tubular throat and correspond in position to the aperturesformed around the inlet port in the transverse wall of the compressor housing. The protrusions engage the apertures to join the insertto the compressor housing such that the circular disc forms a forward diffuser wall ofthe turbocharger.

Thetubularthroatoftheinsertfitswithintheinlet port allowing the circular disc to abut the transverse wall ofthe compressor housing for engagement thereto. The disc extends beyond thetransversewall and partially encloses the opening in the plane of the wall ofthe circumferential chamber.

In accordance with one embodiment of the invention, the inlet port is formed with an inside surface having a first portion remote from the transverse wall with a varying diameter converging toward the wall and a second portion adjacent the compressor housing transverse wall diverging in diametertoward the wall.Thetubularthroatofthewall insert has an inside surface converging in diametertoward the circular disband an outside surface diverging in diameter toward the disc two mate with the diverging inner surface of the inlet port ofthe compressor housing.

The inner surface of the throat portion of the wall insert converges at substantially the same rate as the inlet port innersurface to form a continuous converg ingsurfacefromtheopening into the inlet port to the area adjacentthe insert disc where the compressor rotor is positioned.

As is now appreciated, the compressor housing unit consists of two components which are cast separately and joined to form the completed unit. By so doing, the components are designed so thatthey may be die-castwhilst still providing a forward diffuser wall for closing the circumferential chamber exceptfor an annular gap through which compressed airischan- neled from the blade tips of the compressor rotor.

In accordance with another embodiment ofthe invention,the housing unit further includes a bearing support cylinder positioned from the inner wall of the inlet port by a plurality of vanes extending from the inner wall. The vanes are aligned to be substantially radial from the center line ofthe inlet port with a leading edgefacingthe opening to the inlet port and a trailing edge substantially opposite the leading edge.

The leading and trailing edges are separated by thicker intermediate vane sections.

In accordance with still another embodiment ofthe invention, one end ofthetubularthroat of the compressorwall insert is notched to correspond with the vanes such that the throat straddles the vanes when the tubularthroat is inserted within the inlet port.

In accordance with still another embodiment of the invention, a turbocharger having a compressor rotor and turbine rotor mounted for rotation on a shaft supported in a housing forming an inlettothe compressor rotor includes a concentrically positioned cylindrical module mounted within the inlet upstream ofthe compressor rotor and supportedfrom the inlet wall by struts extending from the inlet wall to the cylidrical module. The struts dividethe inlet into more than one inlet passage. The compressor inlet has a converging inner wall from the mouth of the inletto the compressor rotor. The struts are formed with a leading and trailing edge thinnerthan the portion of the strut intermediate thereof.The inlet diameter is shorterthan the length of the inlet. In amorespecific embodiment of the invention, three struts extend substantially radiallyfrom the cylindrical moduleto the side wall ofthe inlet and divide the inlet into three inlet passages.

It has been found thatthis inlet structure substan tially eliminates the surge area normally encountered by conventional turbochargers at low compressor flows and allows the turbocharger to continue to function at low compressor flows and corresponding low engine speeds without producing unstable flow at low operating speeds.

In accordance with still another embodiment ofthe invention, the turbochargerfurther includes control structureforvarying the velocity of exhaust airto the turbine rotor in response to the speed ofthe rotor shaft to maintain the compressor rotor speed at any desired level at various engine conditions. A plurality of nozzle vanes are spaced around theturbine rotor and in the inletthroughwhich exhaust airfrom the engine is received to the turbine rotor. Control structureis provided for rotating the nozzle vanes to vary the flow ofexhaustairtotheturbine rotor.A shaft speed monitor operates to control the position ofthe nozzle vanes in accordance with the speed ofthe shaft such thatthe nozzle vanes are moved to reduce the inlet area tothe turbine rotorwhen the shaft speed drops belowa predetermined level and toincreasethe inlet area to the turbine rotorwhen the speed of said shaft increases above a predeterminedvalue. In this way, the speed of the turbine and the compressor driven by the turbine is controlled as desired without regard to the engine speed.

In accordance with still another embodiment of the invention, an axial compressor is mounted on the compressor rotor and turbine rotor shaft in the turbine compressor air inlet upstream ofthe radial compressor. The axial compressor operates to compress air passing into the air inlet priorto introduction ofthe air to the radial compressor. The radial compressor and turbine rotors are overhung to one side of the shaft bearing support structure and the axial compressor rotor is mounted on the shaft on the side ofthe bearing support structure oppositethe radial compressor and turbine.

In accordance with still another embodiment ofthe invention, an armature is mounted for rotation with the shaft and afield winding is mounted in the turbocharger housing adjacentthe armature such that an electriccurrentis produced bythe rotation ofthe armature with the shaft within the field winding. In accordance with a more specific embodiment of the invention, wherein the shaft is supported for rotation in a bearing housing supportedfromthe compressor and turbine rotor housing,thefield winding is mounted in the bearing housing.

In accordance with oneembodimentoftheinven- tion,theturbochargerdescribed above is mountedto an internal combustion engine with theturbine and compressor axis of rotation in a vertical orientation.

When mounted to a "V" engine, theturbochargeris nested between the cylinder bankswith the engine exhaust ported directlyfrom the cylinderstothe turbocharger to drive the turbine rotor.

In accordance with another embodiment ofthe invention,the basicstructure oftheturbocharger may be modified to operate astheturbojet. In this embodiment, the turbojet includes a shaft, having a turbine rotor and first compressor rotor mounted on the shaft. Bearing structure is provided for rotatively supporting the shaft within a housing such that the compressorandturbine rotors are overhung to one side ofthe bearing structure with the bearing structure on the side of the compressor rotor opposite the turbine rotor. An airflow channel is formed within the housing for directing airto the compressor rotor and thereafter to the turbine rotor. Structure is provided for introducing fuel within the airflow channel to form an airfuel mixture within the airflow channel.An ignitor is provided within the airflow channel intermediate ofthe compressor rotorandturbine rotorfor igniting the airfuel mixture to drive the turbine and provide thrust from the turbojet.

In accordance with a more specific embodiment of the invention,theturbine rotor and compressor rotor shaft has an axial bore formed therein. Ports communicate between the axial bore and the shaft and through the compressor rotortothe airflow channel providing forfuel from the axial bore in the shaftto move from the bore to the airflow channel. Fuel is supplied to the bore ofthe shaft and is directed by centrifugal force during the rotation of the compressor rotor from the bore in the shaft th rough the compressor rotortothe airflow channel thereby eliminating the needforafuel pump.

In accordance with still a further embodiment ofthe invention, an axial compressor is mounted on the shaft upstream ofthefirst compressor rotorfor compressing air moving into the airflow channel. The axial compressor rotor is mounted on the side ofthe bearing structure opposite the first compressor rotor.

This structure provides a higher cycle pressure ratio while maintaining lower centrifugal stresses. Alternatively, where compressor rotor stresses are not critical,a much higher cycle pressure ratio maybe achieved.

In accordance with still a further embodiment of the invention, an armature is mounted on the shaft within the airflow channel and on the side ofthe bearing structure opposite the first compressor rotor. A field winding is mounted in the housing adjacent the armature, whereby electric current is produced when the armature rotates relative to the field winding. Use of rare earth metals such as selenium-cobalt would permit current generation at the high rotative speeds.

In accordance with still a further embodiment of the invention, a turbofan is provided including a turbine rotorandfirstcompressor rotormounted on a shaft for rotation thereon. The shaft is supported by bearing structure mounted within a housing such that the compressor andturbine rotors are overhung to one side ofthe bearing structure. An axial compressor having an axial compressor rotor is mounted on the shaft on the side of the bearing structure opposite the first compressor rotor.A primaryairflowchannel is formed within the housing for directing airto the axial compressor, then to the firstcompressor rotor and thereafter to the turbine rotor A fuel system introduces fuel within the airflow channel to form a fuel air mixture within the primary airflow channel and an ignition souce is positioned within the airflow channel intermediate of the compressor and turbine rotors to ignite the airfuel mixture. A bypass channel is also formed within the housing to channel a portion ofthe inlet airsubsequentto compression by the axial compressorthrough a second nozzle. The axial compressor rotor is adapted with an armature which cooperates with a stationary field winding to produce electric current.

For a more complete understanding of the present invention, and forfurther details and advantages thereof, reference is now made to thefollowing description taken in conjunction with the accompanying drawings, in which: FIGURE lisa perspective view oftheturbocharger of the present invention, FIGURE 2 is a vertical section taken along lines 2-2 of the turbocharger illustrated in FIGURE 1, FIGURE 3 is an enlarged view of the bearing assemblies supporting the compressor and turbine shaft of the present invention, FIGURE 4 is an enlarged section viewtaken along line 4-4 of FIGURE 2 with the compressor backwall removed for clarity, FIGURE 5 is a partially broken away end view as viewed from line 5-5 of FIGURE 2 looking in the direction of the arrows, FIGURE 5a is a section view taken along line 5a-5a of FIGURE 5 looking in the direction of the arrows, FIGURE 6 is a partially broken awayendviewofthe turbocharger ofthe present invention with an alterna tive method of controlling the turbine inlet vanes, FIGURE 6a is a block diagram further illustrating the control unit illustration in FIGURE 6, FIGURE 7 is a vertical section taken along line 7-7 of FIGURE 6, FIGURE 8 is an exploded perspective view showing the turbine rotor, compressor-turbine shaft, the compressor and turbine backwalls, compressor bushing, compressor rotor and retaining sleeve, FIGURE 9 is a section view taken along line 9-9 of FIGURE 2, FIGURE loins an alternative embodiment of the present invention showing the use of a spring member between the retaining sleeve and the compressor rotor, FIGURE 11 is an exploded perspective view showing the two components of the compressor housing unit separated from molds used to makethecomponents, FIGURE 12 is a frontal view looking into the inlet port oftheturbochargerillustrated in FIGURE 1, FIGURE i3 is a section viewtaken along line 13-13 of FIGURE 12, FIGURE 14 is a section viewtaken along line 14-14 of FIGURE 12, FIGURE 15 is a section viewtaken along line 15-15 of FIGURE 12, FIGURE 16 illustrates an alternative embodiment of the present invention wherein the forward diffuser wall is modified to modify the turbocharger characteristics, FIGURE 17 illustrates a compressor performance map for a conventional turbocharger, FIGURE 18 illustrates a compressor performance mapfortheturbocharger illustrated in FIGURES 1-16, FIGURE 19 illustrates a vertical section of an alternative embodiment ofthe turbocharger of FI GURES 1-16, FIGURE 20 illustrates a vertical section of the turbocharger illustrated in FIGURES 1-16 mounted on a "V" engine, FIGURE 21 is a top plan view of the assembly illustrated in FIGURE 20, FIGURE 22 illustrates a vertical section ofthe turbocharger of FIGURES 1-16 mounted on an in-line engine, FIGURE 23 is a top plan view of the assembly illustrated in FIGURE 22, FIGURE 24 illustrates a vertical section of the upper half of a turbojet in accordance with the present invention, FIGURE 25 illustrates a vertical section of a turbofan in accordance with the present invention, FIGURE 26 illustrates a vertical section ofthe upper half of an alternative embodiment of the turbojet illustrated in FIGURE 24, FIGURE 27 is a section view taken on line 27-27 of FIGURE 26, FIGURE 28 illustrates a vertical section showing the upper half of a turbofan according to the present invention, and FIGURE 29 is a vertical section of an electric generator adapted to the turbocharger of FIGURES 1-16 and for adaption to the turbojets and turbofan of FIGURES 24-29.

FIGURE lisa perspective view of a turbocharger 20 embodying the present invention. The turbocharger includes an outer structure 22 consisting of a com pressor housing unit 24 coupled to a turbine housing unit 26 by a V-clamp band 28.

Referring to FIGURES 1 and2, compressor housing unit 24 includes a tubular inlet port 40 with a transverse wall 42 attached to one end of port40 and extending outwardly therefrom. A circumferential chamber44 is attached from wall 42. Inlet port 40 defines a compressor air inlet 50 and circumferential chamber44 defines a compressor exhaust 52. Turbine housing 26 defines a turbine air inlet 54 and a turbine exhaust 56.

In operation of the turbocharger, air is drawn into inlet 50 and compressed air is discharged from exhaust 52 to an internal combustion engine to which theturbocharger is mounted. Exhaust air from the engine is channeled into turbine air inlet 54 to drive the turbocharger turbine and is exhausted through turbine exhaust 56.

Referring still to FIGURES 1 and 2, a bearing support cylinder 60 is mounted within inlet port40 buy a plurality of vanes 62 extending from the inside wall surface 64 of inlet port 40. A cap 66 is mounted over the end of support cylinder 60. A piston type actuator 80 is mounted by bracket 82 (FIGURE 1) to turbine housing 26. Actuator 80 includes a controller 84 operated to extend and retract control rod 86 as will be discussed hereinafter in greater detail. Air lines 88 and 89 provide airto controlier 84 as necessaryto operate rod 86. An oil reservoir cover plate 90 is attached to compressor housing unit 24 by a plurality of screws 92.

Referring specifically to FIGURE 2, a compressor backwall 100 and a turbine backwall 102 are posi tioned intermediate ofcompressor housing unit 24 and turbine housing unit 26 when these two units are assembled. These fourcomponentsare piloted one to anotherand held in assembly byV-clamp28. As can be appreciated from a close review of FIGURE 2, only a single V-clamp is required to hold the entire assembly together. As previously described, compressor hous ing unit 24 includes a tubular inlet port 40 with a transversewall 42 attached to one endthereofto one ofthe inlet ports and extending outwardly therefrom.

Acircumferential chamber44 is attached to the end of wall 42 remote from inlet port 40 and has a varying area around its circumference increasing to the discharge provided by compressor exhaust 52 (FI GURE 1).

Inletport40 hasafirst inside wall portion 110 having a converging diametertoward wall 42 and a second inside wall portion 112 joined to first inside wall portion 110 by a step 114. The second inside wall portion 112 has a diverging diametertoward wall 42.

Wall 42 has a plurality of circumferentially spaced apertures 116 therethrough. Chamber 44 has an opening 118 substantially in the plane of wall 42 in addition to compressor exhaust 52.

Aforward compressor wall insert 126 includes a tubularthroat 128 and a circular disc 130 attached transversely from one end ofth roat 128. Throat 128 has an inside wall surface 131 having a diameter converging toward disc 130 and an outer surface 132 having a diameter diverging toward disc 130.The diverging diametersurface 132 corresponds to the diverging surface of inside wall portion 112 of inlet port 40 such that throat 128 may be inserted within and mated with inletport40.Theconverging diameter inside wall surface 131 ofthroat 128 corresponds to the extension of converging diameter offirst inside wall portion 110 of inlet port 40. When insert 126 is mated into inlet port 40, a continuous converging diameter is providedfromthe inletof port40 inwardly intotheturbocharger.

A plurality of rivet-like protrusions 140 extendfrom disc 130 and correspond to apertures 116 in wall 42.

With the insertengagedtocompressorhousing 24 with the end of throat 128 engaging step 114 of inlet port 40, protrusions 140 are engaged through apertures 116 with disc 130 abutting the corresponding surface of wall 42. As is shown in FIGURE 2, protrusions 140 have been inserted into apertures 116 and the heads thereof deformed to attach insert 126to housing 24. Disc 130 extends beyond wall 42 to partially cover opening 118 of chamber44. A circumferential gap 146 is formed between the outertip of disc 130 and the wall of chamber 44, and a diffuser area 148 is formed between disc 130 and compressor backwall 100 between centrifugal flow compressor rotor 172 and gap 146 leading to chamber44.

Referring still to FIGURE 2, bearing support cylinder 60 is supported concentrically within inlet port 40 by a plurality of vanes 62 extending inwardlyfrom wall surface 64 of port 40. Turbocharger 20 further includes a shaft 160 supported for rotation in bearing support cylinder 60 bytwo ball bearing assemblies 162 and 164. A radial flowturbine rotor 170 is mounted at one end of shaft 160 and a centrifugal flow compressor rotor 172 is mounted intermediate ofturbine rotor 170 and bearing assemblies 162 and 164. Shaft 160 passes through aperture 176 in compressor backwall 100 and labyrinth seal 178 in turbine backwall 102.

Turbine rotor 170 is fixedly attached to shaft 160, such as byweldingrand compressorrotor 172 is retained in position on shaft 160 by retainer nut 180.

Compressor rotor 172 is drilled to receive shaft 160 and counterbored to form a bore 182. Bore 182 has a diameter largerthan the outer diameter of retainer nut 180 such that retainer nut 180 may be pressed onto shaft 160 into engagement with the bottom wall 184 of bore 182 to retain the compressor rotor in position on shaft 160. Acompressor rotor shim 186 is positioned between compressor rotor 172 and a step 188 in shaft 160 to accurately position the compressor rotor in the axial direction.

Referring to FIGURES 2 and 3, a ring 200 is fitted within the end of cylinder 60 adjacent compressor rotor172 and is preventedfrom moving into cylinder 60 bya retaining ring 202 attached to cylinder 60.

Outer raceway 204 of bearing assembly 164 is formed in ring 200, the inner raceway 206 being integrally formed in shaft 160. Balls 208 are engaged between the inner and outer racewaysto form bearing assembly 164.

Bearing assembly 162 includes inner raceway 210 formed integrally in shaft 160 and an outer ring 212 slidable within cylinder 60 with an outer raceway 214 formed therein for receiving balls 216. A compression spring 218 is engaged between ring 212 and a retaining ring 220fixed within cylinder60 and biases ring 212 outwardly to fix the position of balls 216 and 208 in bearing assemblies 162 and 164, respectively, thereby fixing the position of shaft 160.

As is shown in FIGURES 2 and 3, outer raceway 204 is formed in ring 200 with the ball radius on only one side. Thus, the assembly of bearing assembly 164 is made by positioning a full complement of balls 208 in raceway 206, and engaging ring 200 therearound.

Similarly, outer raceway 214 is formed in ring 212 with the ball radius on only one side. Balls 216 of bearing assembly 762 are assembled by moving outer ring 212 to compressspring 218 and inserting a full comple- ment of balls 216 in raceway 214 of shaft 160. By releasing ring 212, spring 218 automatically forces the ring into engagement with balls 21 6to form bearing assembly 162 while simultaneously engaging ring 200 against balls 208 of bearing assembly 164.

Alternatively, less than a full complement of balls 208 and 216 may be used in bearing assemblies 162 and 164 by the use of an appropriate retainer.

Depending upon the application, an oil impregnated retainerora sacrificial retainerwhich replenishes a self-lubricating coating to the balls may be used. The mounting of shaft 160 within cylinder 60 is completed by the engagment of cap 66 on the end of cylinder 60 to close the opening in cylinder 60 remote from compressor rotor 172.

The use in the present bearing system of the ball bearing arrangement employing integral inner raceways permits a larger diameter shaft and thus provides a very "stiff" shaft. Additionally, this bearing arrangement provides a very tight bearing system permitting very little radial or axial movement. As a result, the present bearing system substantially reduces the clearance required between the compressor and turbine and surrounding housing, and concentricity problems are minimized.

In a preferred embodiment ofthe invention, bearing assemblies 162 and 164are "starved" of oil.The only lubrication provided to the bearing assemblies is through wicks 222 and 224 which transfer oil from a reservoir R by capillary action to ramps or slingers 226. Oil supplied to slingers 226 is projected by centrifual force to bearing assemblies 162 and 164 during rotation of shaft 160.

In this way, the use of engine oil as a lubricantforthe turbocharger bearings, and the associated plumbing and seals are eliminated. Moreover, the failu res resulting from the use of contaminated engine oil as a lubricantorthe lackofengineoil during starts is avoided. Moreover, no oil seals are required and bearing failure resulting from thefailure of seals is also eliminated.

Alternatively, the bearing assemblies may be permanently lubricated with a heavy consistency oil or grease which is packed into the raceways and around the balls ofthe bearing assemblies. As another alternative, an oil impregnated phenolic retainer may be used to provide lubricant to the ballsfora considerable period. In either of these cases, the need for wicks 222 and 224 and slingers 226 would be eliminated although with a reduced bearing life.

In conventional turbochargers where journal bearings and disctype thrust bearings are employed, continuous lubrication of the bearings is required.

Additionally, becausetheturbine is subjected to temperatures upto 1600"F, where the bearings are adjacent the turbocharger turbine, continuous lubrication is required to sufficiently cool the bearingsto prevent failure through overheating. Even ball bearings would require a continuous flow of oil for cooling.

The ability ofthe present system to successfully function without conventional flood lubrication used to both lubricate and cool the bearings is the result of the particular arrangement of the type of bearings used in the present invention and the relative location ofthe bearings to the compressor and turbine.

As will be described hereinafter in greater detail, because the present invention eliminates the need for a continuous flow of oil for bearing lubrication and cooling, the present turbocharger may be mounted at any desired orientation. In contrast, conventional turbochargers have been severely limited asto their possible orientation.

In the present invention, journal and disc bearings of conventional turbochargers are replaced by the more precise ball type bearing assemblies, thereby eliminating the need for a continuous flow of lubrication for the bearings. Additionally, both the compressor and the turbine ofthe presentturbochar gerare overhung to one side of the bearing assemb- lies, and the turbine is maximally removed from the bearing assemblies with the compressor positioned between the bearing assemblies and the turbine. As this arrangement provides sufficient thermal insulation between the turbine and the bearing assemblies, the bearings will not be sufficiently heated as to require conventional lubrication methods.

The backto back compressor/turbine arrangement of the present invention not only reduces substantially the heat conducted to the bearing assemblies but also minimized the thermal expansion effects, thereby lowering the bladetip clearances required. This arrangementfurther eliminates the conventional bearing housing and provides a more compact packagethan present straddle mounted rotors having the support bearings intermediate of the compressor and turbine.

Additionally, the use of antifriction ball bearing assemblies 162 and 164 provides much better control overthe radial and axial movement ofthe compressor and turbine, thereby allowing reduced blade tip clearance. This in turn appreciably improves the compressor and turbine efficiency. Compared to journal bea rings, the use ofthe antifriction ball bearing assemblies 162 and 164 also reduces the turbine work required to drive the bearings. This in turn reduces the engine back pressure resulting in the improvementof the specific fuel consumption as well as enhancing the ability to accelerate the rotor.

The present invention is also directed to a nozzle area control structure 228 for selectively varying the turbine nozzle area to control the speed or pressure output oftheturbocharger. Referring to FIGURES 1 and 2, exhaust gas from the internal combustion engine on which the turbocharger is mounted is injected into the turbochargerthrough turbine air inlet 54 and channeled against the blades of turbine rotor 170through a nozzle area 230formed byturbine backwall 102, a wall 232 parallel thereto andthe nozzle vanes 234.This nozzle area is controlled by structure 228 including a plurality of movable nozzle vanes 234 positioned circumferentially about the nozzle area and rotatable to vary flow velocity and angle of exhaust gas to turbine rotor 170. Referring to FIGURES 2 and 4, vanes 234 includetrunnions 236 and 238 extending from opposite sides thereof. Trunnion 236 extends through turbine backwall 102 and isattachedto actuation lever 240. Trunnion 238 extends into wall 232.

A nipple 242 is formed on one end of each actuation lever. These nipples extend into radial slots 244 formed in a control ring 246. Control ring 246 and actuation levers 240 are situated in air space gap 247 intermediate of compressor rotor 172 and turbine rotor 170. Control ring 246 is concentrically positioned aboutthe axis of shaft 160 and is recieved on a cylindrical surface248 extending from compressor backwall 100.

In a preferred embodiment ofthe invention, control ring 246 includes an inner ring 250 and an outer ring 252 formed with an inner and outer raceway, respectively, for receiving a plurality of balls 254 therebetween. Inner ring 250 is fixedly attached to the cylindrical surface 248 extending from compressor backwall 100, and outer ring 252 rotates angularly relative to the inner ring. Referring to FIGURE 4, it is seen that by the rotation of outer ring 252, each ofthe actuation levers 240 is rotated aboutthe axes of trunnions 236 and 238, resulting in the simultaneous rotation of each nozzle vane 234. As is shown in FIGURES 4 and 5, one ofthe actuation levers 240 is provided with an extension 262. Control rod 86 is attached to the end of extension 262 remote from nipple 242.Control rod 86 includes a threaded eye bolt 36a attached to the extension 262 by an axis pin. The opposite end of eye bolt86a is threadedly received within control rod 86 and is adjustable therein to allow forthe readjustment of nozzlevanes 234 abouttheir rotational axes. Bythe movement of control rod 86, actuation lever 240 is pivoted to angularly rotate outer ring 252 of control ring 246 thereby rotating each of the other actuation levers 240 and nozzle vanes 234 attached thereto.

Referring to FIGURES, an actuatorfor use on an otto cycle engine is shown. In the system illustrated, control rod 86 is controlled by a piston type actuator 80. Alternatively, the pistons could be replaced by diaphragms. Actuator 80 is controlled by compressor discharge pressure fed into controller84through line 88. Increased pressure into actuator 80 causes the extension of control rod 86 and the corresponding opening ofthe compressor nozzle area.

Referring to FIGURE 5a, control rod 86 is controlled bycontroller84tovarythe position of nozzle vanes 234. Controller 84 includes a housing 264 defining a large cylinder 265 and a small cylinder 266 therein. A large piston 268 having a forward face 268a and a rearward face 268b is received within cylinder 265. A smaller piston 270 having a forward face 270a and a rearward face 270b is received within small cylinder 266. Afront plate 272 is attached to housing 264 by suitable bolts and has an aperture 273 therein for receiving control rod 86 into the housing. Piston 270 is fixedly attachedto the end of control rod 86while piston 268 is free to slide relative to control rod 86 being arrested only by ring 274 fixedly attached to control rod 86 adjacent forward face 268a of piston 268.

A spring 276 is entrained around control rod 86 and captured between piston 268 and front plate 272. A larger spring 278 encircles control rod 86 and is captured between piston 270 and piston 268. Air lines 88a and 88b communicate to cylinder 266 to the rearward face 270b of piston 270 and within cylinder 265 to the forward face 268a of piston 268. Air lines 88a and88b have their ends opposite controller 84 attached to the engine intake manifold to convey manifold pressure to the forward face 268a and rearward face 270b of pistons 268 and 270, respectively. Line 89 provides atmosphericpressuretothe rearward face 268b and forward face 270a of pistons 268 and 270, respectively.

Annular grooves 280 and holes 282 th rough guide cylinder 284 protruding from the rearward face 268b of piston 268 assures the transmission of atmospheric pressure along the full back surface of piston 268.

Likewise, radial grooves 286 on the rearward face of piston 268 facilitate the communication of pressure to the full surface of the rearward face of the piston.

Controller84 is designed to extend control rod 86 to open nozzle vanes 234 to turbine rotor 170 at low and high manifold pressures and to retract control rod 86 to close these nozzle vanes at intermediate pressures.

More specifically, while operating at a closed throttle position where the intake manifold is under a vacuum condition, for example with manifold pressures of 5 to 8 pounds per square inch absolute,thecontroller operates to extend rod 86to open nozzle vanes 234 permitting the engine to operate with as little back pressure as possible. This results from the larger atmospheric pressure (approximately 14.7 psi absolute) exerted on rearward face 268b of piston 268 from line 89 compared to pressurefrom line 88 which is substantially less than atmospheric pressure.

Whilethe net pressure differential across piston 270 provides a force tending to retract control rod 86, the surface areas of pistons 268 and 270 are designed so that the outward force on piston 268 is sufficient to overcome both the retractive pressure on piston 270 as well as the spring force of spring 276.

The piston size and spring rate of spring 276 are designed such that as the manifold pressure nears atmospheric pressure,such as 12 pounds persquare inch absolute, control rod 86 is drawn within controller 84to close nozzle vanes 234thereby increasing the speed of turbocharger 20 and the pressure ratio provided from the turbocharger compressor. Control ler84isdesignedto begin closing vanes 234 immediately before atmospheric pressure is reached because this intake manifold pressure corresponds to an open carburetorthrottle and thus indicates a requirementforturbocharger boost.

Referringto FIGURE 5a, asthe manifold pressure in line 88a exerted againstforwardface268a of piston 268 nears atmospheric pressure. this pressure in conjunction with the action of spring 276 overcomes the atmospheric pressure exerted from line 89 on the rearwardface 268b of piston 286 to withdraw control rod 86 within controller 84. The pressure differential across piston 270 still provides a net force retracting piston 270 within its cylinder thus assisting in the retraction of control rod 86 within controller 84.

As the manifold pressure increases to a value in excess of atmospheric pressure, for instance 20 pounds per square inch absolute, the manifold pressure exerted through line 88b on the rearward face 270b of piston 270 overcomes the spring force of spring 278 causing the movement of control rod 86 outwardly from controller 84to again open vanes 234 thereby preventing overpressurization. Control rod 86 is extended from controller 84 in this pressure situation even though piston 268 is retained in a fully retracted position within cylinder 265 because ofthe capability of control rod 86 to move freely relative to piston 268 when piston 270 overcomes the spring force of spring 278 to move toward piston 268.

Therefore, the present controller provides a system for opening the nozzle vanes 234 at manifold pressures having an absolute reading below a predetermined value less than atmospheric as provided for by the design of the controller and components therein.

Likewise, the controller provides forthe closing of vanes 234 at manifold pressures ranging from some predetermined value less than atmosphericto some value greaterthan atmospheric pressure and forthe opening of vanes 234 at manifold pressures in excess of a predetermined value above atmospheric pressure. Of course, it will be understood that the particular pressures at which the vanes are opened and closed can be readily altered by variation of the areas of pistons 268 and 270 and the variation in the spring rates and initial deflections of springs 276 and 278.

For diesel cycle engines, a solenoid may be used to control piston 268 such that the solenoid maintains shaft 86 in an extended position during light load operations ofthe engine but releases the shaft when a heavy load is demanded. Upon release ofthe shaft, spring 276 retracts shaft 86 to close the nozzles thereby increasing turbochargerspeed. When sufficient pressure is communicated from the intake manifold to face 270b of piston 270, shaft 86 is extended as previously described to open the nozzles thereby controlling manifold pressure.

In operation of the variable area turbine nozzle, a control signal, such as compressor discharge pressure, is communicated to actuator 80 which appropriately extends or retracts control rod 86 in accordance with the signal to actuator 80. This in turn results in the rotation of actuation lever 240 attached to control rod 86 and the simultaneous angular rotation of outer ring 252Of control ring 246. The rotation of outer ring 252 in turn rotates each actuation lever 240 to correspon dinglysettheangle of nozzle vanes 234.

FIGURE 4 shows in phantom and solid lines nozzle vanes 234 in closed and open positions. It will be appreciated that each of the nozzle vanes is set by the rotation of control ring 246through the movementof a single control rod 86 controlled by a single actuation lever 240.

Conventional uncontrolledturbochargers normally produce a boost pressure directly related to com pressorspeed,with compressor speed generally corresponding to engine speed. Therefore, boost pressure in conventional uncontrolled turbochargers is substantially less at low engine speeds than at higher engine speeds. Therefore, at a low engine speed a conventional turbocharger is ineffective in improving engine performance. Moreover, upon acceleration, there is a substantial "lag time" until the turbocharger reaches a speed producing a sufficient boost pressure to effectively improve engine performance.

Theturbocharger of the present invention may be controlled to provide boost pressure at low engine speeds. This results in improved engine performance at lower engine rpm and eliminates or reduces the "lag time" associated with conventional turbochargers.

The present invention accomplishes this result through the control of the variable turbine nozzles to maintain the turbochargerturbine and compressor speed at a level that produces the desired boost at all engine RPM levels. Referring to FIGURES 6 and 7, illustrating this embodiment of the present invention, a control unit 400 is substituted for actuator 80.

Control unit400 includes a magneticsensor402 (FIGURE 7) connected to a monitor 403 by way of a pulse-to-voltage converter 404. The revolutions of shaft 160 are monitored by sensor 402 in cooperation with a disc 160" mounted on shaft 160. Disc 160" has a plurality ofdiscontinuitiesthereon. The speed of shaft 160 is determined in the well known manner by converting the electrical pulse generated by the movement ofthe disc relative to magnetic sensor 402 and by inputting these pulsesto pulse-to-voltage converter404. lnturn,thisvoltage, proportional to the rotational speed of disc 160" and shaft 160, is converted to shaft speed by monitor 403.Monitor 403 is connected by appropriate leads to a proportional controller4l 4. Controller 414 is connected to a power source 416 and to nozzle area controller unit 415. Unit 415 controlsthe movement of rod 86to rotate lever extension 262 and vary the orientation of vanes 234 in theinletareatoturbinerotorl70.

In one embodimentofthe invention, controller 414 is a microprocessor including a conversion system for receiving analog input from sensors 417 and speed monitor 403 for conversion to digital format usable by the microprocessor. The microprocessor may be of the general type now used in automobile ignition systems. Alternatively, a microprocessor which can be incorporated in the present system is microprocessor model 8080 manufactured by Intel. Proportional controller 414 also includes an amplification system for amplifying the signal provided from the microprocessor for controlling controller unit 415.

The microprocessor may be programmed to control theturbine speed, and thus the boost pressure provided by the turbocharger compressor, as a function of several parameters, for example, engine speed, throttle position, intake manifold pressure, engine torque, ambient temperature, ambient pressure or transmission gearing. As is shown in FIGURE 6, sensors 417 are provided to meter each ofthese engine parameters and provide the corresponding analog reading to the proportional controller414 where they are converted to the digital form for use by the microprocessor. Of course, the microprocessor may be programmed to position the turbine inlet nozzle to control compressor speed according to any directorvariable relationship ofthe parameters.

Referring nowto FIGURE 6a,the proportional controller414shown in FIGURE 6 is illustrated in greater detail. A plurality of sensors 417a-g measure a parameter relevant to the operation of the automobile engine. Each ofthe sensors 417 plus the turbine speed monitor 403 produces an analog signal which is transmitted to respective analog to digital converters 424a-h. The signals produced by each of these sensors is transformed within the corresponding analog to digital converter into a series of digital words which representthe numerical value ofthe parameter being measured.

The digitized signals produced by the analog to digital converters 424 are conveyed to a microp rocessor426which is included within proportional controller 414. Microprocessor 426 receives the digitized signals which represent the parameters being measured by sensors 417. The microprocessor evaluatesthe combination of signals received and produces therefrom a signal which corresponds to the optimum control position forthe nozzle vanes 234. The output signal produced by microprocessor 426 is transferred to a digital to analog converter428which produces a corresponding nozzle control signal. The output ofthe digital to analog converter 428 is connected to a driver circuit 429 which amplifies the nozzle control signal.

The amplified signal istransmitted tothe noale area controller unit 415 for selectively positioning the nozzle vanes 234. The driver circuit 429 not only provides amplification forthe nozzle control signal, but also isolates the proportional controller 41 4from the controller unit 415 as well as from any spurious signals generated in the region of proportional controller414.

The proportional controller414 and sensors 417 of the present invention utilize the technologywhich has previously been developed for automotive engine control directedto sparktiming and fuel metering.

Existing systems ofthis nature are described in an article by Gerald M. Walker, "Automototive Electronics Gets The Green Light," ELECTRONICS, Volume 50, No.20, (September 29,1977), pages 83.92. Thus, with the present invention, the engine designer may provide a desired boost pressure corresponding to one or more of these or other parameters. Of course, the parameters listed above are only indicative of some ofthe parameters which an engine designer maywish to use in choosing the optimum turbocharger compressor speed throughoutthe operation range ofthe engine and turbocharger.

The simplest arrangement is represented bymerely maintaining theturbine speed, and thus the compressor speed, at a constant level throughoutthe operation ofthe engine. In this situation, a set point speed is selected for controller 414. Controller 414 provides a signal to actuator 418 in response to a comparison of the set point speed to the speed indicated from monitor403. When the shaft speed monitored by monitor403 isbelowthesetpointvalue of controller 414,controller 414 provides a signal to actuator 41 8to retract rod 86 resulting in the rotation of vanes 234 and closing down ofthe turbine nozzles.

As the turbine nozzles are closed down, the turbine speed is increased, increasing the rotation speed of shaft 160. When the shaft speed indicated by monitor 403 is greaterthan the set point value, the signal from controller 41 4to actuator 41 causes the extension of rod 86 and the corresponding opening ofthe turbine nozzles resulting in a slowdown in the speed ofthe turbine and shaft 160 attached thereto. By continuing this operation, the speed oftheturbine and compressor may be controlled in accordance with a predeter minedsetpointvalue programmed into controller 414. Therefore, regardless ofthe engine speed and throttle position, a constant compressor and turbine speed may be maintained by varying the position of vanes 234 to control the flow velocity of exhaust gases againstturbine rotor 170.

Using the arrangement ofthe present invention, a relatively high compressor speed may be maintained regardless of engine speed. This is accomplished by the capability ofthe present invention to varythe turbine nozzle area. Thus, upon engine acceleration, no "lag time" in boost pressure is experienced becausethere is no requirementthatthe compessor speed be increased from a relatively low speed to the higher values constantly maintained bythe present turbocharger. Moreover, by maintaining high boost pressures at relatively low engine speeds, the present turbocharger provides improved engine performance at low engine rpm as well as at higher engine speeds.

The operation ofthe turbocharger as just described assumes that the turbocharger compressor is capable of operating through broad flow ranges as in the present invention. As will be described hereinafter, while some advantage would be provided to conven tional turbochargers,the advantage provided bythe variable area turbine nozzles of the present invention is greatly extended by its application to the present compressorwhich is capable of operating over a broad flow range. Thus, conventional turbochargers, plagued with a stall line prohibiting the operation of the compressor over a broad flow range, would not be benefited to the same degree as the present invention bythe application ofthe variable area turbine nozzles.

Therefore, the present invention provides a turbochargerwiththeturbine nozzle area controlled by the movementofa single control ring rotating a plurality of actuatorarms circumferentiallyspaced aboutthe control ring. Moreover, the control ring and actuation levers are positioned in the unused airgap space existing between the turbochargerturbine and compressorthereby making a very compact unit.

Moreover, the particular location ofthe linkage further insulates the bearing assembliesfrom the heat to which the turbine and the surrounding area is subjected. In this way, the bearing assemblies may be oil starved or limitedly lubricated. This in turn eliminates the need for conventional lubrication by use of engine oil and the associated plumbing requirements required by such a lubrication method.

This, of course, makesthe turbocharger ofthe present invention less expensive to construct and operate as well as more reliable.

Moreover,theuse inthe presentinvention of ball bearing supports, and the corresponding elimination ofthrust and disc type bearings, provides for a more controlled or "stiffer" turbine and compressor rotat ing assembly thereby allowing smaller clearances between both the turbine and the compressor and their surrounding structures. For example, in the present invention,the compressor blade height for a turbochargerfor use with relatively small internal combustion engines could be on the order of 0.2 inches. The use ofthe ball bearing assemblies described above permits the design ofthe compress orwith only a .005 inch clearance amounting to only 2.5 percentofthe overall heightofthe blade.By contrast, were a journal and disotype bearing to be employed, a substantially greater clearance would be required (normally about .015 inch). As a result, the present invention is particularly adaptable to the construction of a turbochargerforsmaller internal combustion engines where the smaller blade heights require closer clearances between the blades and the surrounding structure.

The use of ball bearings and their resultant lower frictional losses when compared to the journal and disctype bearings also makes possibleforthefirst time the efficient use of a turbocharger which produces less turbine horsepower without losing a substantial portion oftheturbine horsepower to bearing losses and friction. Where a journal and disc type bearing could account for a 4-5 horsepower loss in friction, the loss in friction by the present arrangement would be on the order of 0.1-0.4 horsepower.

FIGURE 8 illustrates compressor rotor 172 and retainer nut 180 separated from turbine rotor 170 and compressor-turbine shaft 160. As is shown in FIGURE 8, turbine rotor 170 is attached to one end of shaft 160, such as by welding or other suitable permanent attachment means. The shaft, priorto its attachment to turbine rotor 170, is formed with an enlarged bearing surface 160a, and a step 188 to a narrower diameter shaft portion 160b. As has been previously discussed, raceways 206 and 210 are formed directly in shaft 160.

In assembly, shaft 160 is inserted through apertures in compressor backwall 100 and turbine backwall 102. Shim 186 is positioned over shaft 160 into engagement with step 188 on shaft 160. Compressor rotor 172 is engaged overshaft 160 and rests on shaft portion 1 60b. Retainer nut 180 is then pressed onto portion 160b into bore 182 of compressor rotor 172 and into engagement with bottom wall 184 of bore 182. Nut 180 consists of a sleeve having a smooth inner bore 180a therethrough. The borethrough nut 180 forms in interference fit with portion 1 60b of shaft 160. The interference, in one embodiment ofthe invention, is on the order of .001 of an inch.

Because the inner raceways are formed directly in thecompressor-turbine shaft, the shaft must be heat treated to extremely high deg rees of hardness. As a result, the present retaining sleeve secures compressor rotor 172 to shaft 160 without necessitating the grinding or cutting ofthreads into the hardened shaft. Thus, the cost and problems associated with forming threads on a heattreatedshaftareelimin- ated. Further, because the shaft is substantially hardened, the retaining nut 180 may be pressed onto the shaft and withdrawn without damage to the shaft surface.

As also shown in FIGURE 8, nut 180 is formed with threads around the outer facing surface thereof. Bore 182 is of a sufficient diameter as to provide a gap 420 (FIGURE 9) between bore 182 and the threads on the outersurfacoofnut 180. This gap permits the insertion of an appropriate internallythreadod tool for withdrawing the nut from the shaft for removal of the compressor rotor.

In some applications of the present invention, a spring device may be required between the retainer nutandthecompressorrotorto maintain an axial force on the compressor rotor during expansion and contraction of components of the turbocharger.

Unlike an internally threaded nut engaged on a threaded shaft, the retainer nut of the present invention does not have the capability of developing high compressive loading in the compressor rotor or tension in the compressor-turbine shaft when mounted in position. Therefore, in the alternative embodiment illustrated in FIGURE 10, a cone shape or belville spring 422 is inserted between the retainer nut 180 and the compressor rotor 172.

Because the components in the embodiment illustrated in FIGURE 10 are identical or correspond to elements in the embodiment illustrated in FIGURES 2 and 8-9, the same numeral, with the designation prime (')will be used to identify like or corresponding parts to those in the embodiment illustrated in FIGURES 2 and 8-9. Referring now to FIGURE 10, shaft 160' is engaged through compressor rotor 172'.

Retainer nut 180' is engaged over the end of shaft 160' with belville spring 422 mounted on shaft 160' between nut180' and wall 184' of bore 182' in compressor rotor 172'. Belville spring 422 is initially compressed as retainer nut 180' is mounted on shaft 160'. The engagement of nut 180' against shaft 160' is sufficient to overcome any expansive forces developed between nut 180' and compressor rotor 172' by belville spring 422. Instead, the compression of spring 422 between rotor 172' and nut 180' induces an axial load in rotor 172'. Thus, with spring 422 in place, contractions or expansions in shaft 160' or compressor rotor 172' will not result in a zero engagementforce between nut 180' and rotor 172'.

FIGURE 11 illustrates compressor housing unit 24 andforwardcompressorwall insert 126 and their respective molds used in die-casting these two pieces. Heretofore, the compressor housing portion of many turbochargers have been sa nd-cast so that the circumferential chamber in which compressed air is directed for channeling to the intake manifold ofthe engine with which the turbocharger is used, could be formed with a substantially enclosed configuration having a narrow circumferential gap for receiving compressed airtherein,thereby allowing the formation of a diffuser. Forming the compressor housing unit in one piece by sand-casting is substantially more expensive and produces a less accurate structurn with much rougherwall surfaces in the circumferential chamber than a comparable die-cast unit.

However, where the turbocharger compressor housing has been die-cast in the past, it has not been possible to produce the forward diffuser wall as is possible in a sand-casting because of the necessity of providing proper ingress and eg ress for th e die-cast molds.

The present invention provides a two component die-cast compressor housing unit which upon assem bly provides the advantages heretofore only afforded bya sand-cast unit. As is shown in FIGURE 11, compressor housing unit 24 is formed by using an outer mold 290, an inner mold 292, a core mold 294 and a cap mold 296. Referring still to FIGURE 11, mold 292 is formed with raised contour 298 to form circumferential chamber 44. Mold 292 also has a protrusion 300 for mating with a corresponding protrusion from mold 290forforming inlet port 40, vanes 62 and bearing support cylinder 60. Core mold 294 and cap mold 296 are used to form compressor exhaust52.

As can be seen in FIGURE 11, molds 290,292,294 and 296 cooperateto make possiblethe die-casting of compressor housing unit 24. Molds 290 and 292 are constructed with abutting surfaces to form parting line310 on housing unit 24.

Molds 320 and 322 cooperate to produce forward compressorwall insert 126. Molds 320 and 322 have abutting surfaces which engage one anotherto produce insert 126 with a parting line 324 on the outer edgeofdisc 130. As can best be seen in FIGURE 11, throat 128 is formed with notches 326 which receive vanes 62 when insert 126 is mounted into housing unit 24.

FIGURE 12, and section views 13, 14 and 15, illustrate the positioning and configuration of vanes 62. Referring to FIGURES 13, 14and 15, each of the vanes 62 has a leading edge 330 and a trailing edge 332 with a thicker intermediate midsection 334. In each case, the thickest midsection is that indicated by a line 336 defining the parting line between molds 290 and 292 used intheformation of housing unit24.

Thus, vanes 62 may be formed by die-casting using molds 290 and 292 to produce the desired airfoil configuration of a leading and trailing edge separated by a thicker midsection therebetween. This con figuration, shown in FIGURES 13, 14and 15, greatly facilitates the ingress of air into the compressor inlet area, and in the configuration ofthe present invention, may be cast using well known die-casting techniques.

To accomplish an airfoil configuration for struts 62 such thatthe vanes taperfrom a thicker midsection to thinner leading and trailing edges, the die-cast molds mustbe inserted into inlet port40from both the forward and rearward ends. Thus, referring to FI GURE 1 1,thesecond inside wall portion 112 (FIGURE 2) is formed by protrusion 300 of mold 292 while first inside wall portion 110 (FIGURE 2) is formed by mold 290. Both the first and second inside wall portions 110 and 112, respectively, diverge outwardly to permit removal ofthe die-cast mold afterformation ofthe piece.

Thus, the formation of the desired geometry of struts 62 requires a diverging diameter wall portion 112 in inlet port 40 to permit a removal ofthe die-cast molds. However, it is critical to have a continuously converging diameterfrom the inlet of inlet port 40 to the compressor rotor 172. This is accomplished by use ofwall insert 126. Wall insert 126 is formed, also by die-casting, with a diverging diameter outer surface 132 (FIGURE 2) corresponding to the diverging surface of inside wall portion 112 of inlet port40 for mating therewith. Inside wall surface 131 ofthroat 128 isformed with adiametorconvergingtowarddisc 130 corresponding to the extension ofthe converging diameter offirst inside wall portion 110 of inlet port 40.Thus, when insert 126 is mated with inlet port 40, a continuous converging diameter is provided from the inlet of port 40 inwardlyto compressor rotor 172 (FIGURE 2).

Thus, the present invention provides a two-piece structure forforming the compressor housing for a turbocharger, both of which may be die-cast. Upon assembly, those two components produce a com pressorhousing having a pluralityofstruts62 supporting a bearing support cylinder 60 with the vanes having a thicker intermediate section converging to a thinner leading edge and to a thinner trailing edge. The housing further provides a continuously converging inlet nozzlefrom the inletofthe inlet port to the compressor rotor. Further, the compressor housing providosaforwarddiffuserwallforcomplet- ing the circumferential chamber in which compressed gases are channeled to the compressor exhaust.

FIGURE 16 shows an adaptation of the invention illustrated in FlGURE2wherein differing geometries of wall inserts 126 may be substituted one forthe otherfor use with corresponding compressor rotors.

Referring to FIGURE 16, wall insert 126a provides a more restrictive airflow into theturbochargerwhile wall insert 126b providesfora larger compressor rotor and greater airflow into the turbocharger. It will be now appreciated that modification of the design of the present invention may be accomplished by merelyfitting differing wall inserts 1 26to a standard compressor housing unit 24. Thus, several different turbochargers, having different flow capabilities, may be produced from the present invention using a standard compressor housing unit 24 by selecting one of any numberof possible geometriesforwall insert 126 with corresponding compressorwheels.

Thisfeature is significant in thatwall insert 126 is the simpliercomponent ofthe compressor housing unit.

In prior art turbocharger units, either of the die-cast orsand-castvariety, any modifications would involve producing a totally new casting for each new design.

In the present invention, a different airflow/pressure ratio compressor design is achieved by simply substituting a wall insert of a different design in conjunction with anew or modified compressor wheel.

Therefore, the present invention provides a tur bochargercompressor housing unit which may be die-cast. The compressor rotor housing unit includes a compressor housing and forward wall insertfor engagement with the housing. This two component construction permits each of the two components to be die-cast using standard die-casting techniques. In i their combined configuration, a compressor housing is produced having a circumferential chamber closed on all sidesexceptforacircumferential gapfor receiving compressed gases from the compressor rotor. A circumferential passageway is also provided leading to this gap to the chamber wherein acceler ated gases are diffused to increase static pressure.

Moreover, a compressor bearing support cylinder is cast concentric with the inlet port and supported therein by a plurality ofvanes extending from the innerwall ofthe inlet port. The vanes are formed with a leading and trailing edge separated by a thicker midsection. This is accomplished through the use of molds having a parting line substantially at the thicker cross sectional area to permit die-casting of the compressor housing.

While the present invention of die-casting a compressor housing unit in two components has been applied a a to aturbochargerwherein the compressor rotor and turbine rotor are overhung to one side of and supported from a shaft rotatable on ball bearing assemblies, it will be apparentto one skilled in the art thatthepresentinvention may readily be adaptedto turbochargers ofthe conventional design having the compressor rotor and turbine rotor supported on opposite sides of shaft bearing support assemblies.

In this particular application, the heretofore described compressor housing unitwould be directly usable with only modification ofthe bearing support cylinder.

A critical feature of the present turbocharger is its capability of performing over an extremelywide range of compressor speeds and flows without experiencing a "surge line" commonly encountered in conventional turbochargers. This highly significant feature is demonstrated by reference to the performance graphs, known as "compressor maps", illustrated in FIGURES 17 and 18. FIGURE 17 illustrates a typical compressormapfora conventionaltur- bocharger,while FIGURE 18 illustrates a compressor map forthe turbocharger ofthe present invention.

These maps plot airflow in cubicfeet per minute versus the ratio of discharge pressure to inlet pressurefortheturbochargorcompressor.

Referring to FIGURE 17, constant compressor speed lines 430 and isentropic efficiency lines 432 are shown. A surge line 434 identifies the performance points to the left of which the turbocharger compressor cannot operate to produce a uniform output of air.

While the components oftheturbocharger may be modified such thatthe surge line is shifted toward the vertical axis to provideforstable operation at lower turbocharger speeds, and thus lower engine speeds, such a shifting of the operating characteristics ofthe turbocharger reduces the high speed capabilities of the unit. Therefore, turbochargers have heretofore been relegated to providing increased performance at only higher engine speeds.

The present invention provides a turbocharger capableof producing a uniform output at very low turbocharger airflows and effectively eliminates the surge line limitation encountered in conventional turbochargers. Referring to FIGURE 18, compressor performance map forthe turbocharger of the present invention is illustrated including constant compressor speed lines 440 and efficiency lines 442.

As can be appreciated by viewing the compressor map illustrated in FIGURE 18, the compressor ofthe presentturbochargercontinuesto operate with no irregularity in compressor output to very low airflow values. Effectively, the turbocharger of the present invention does not experience a "surge line" corres ponding to that illustrated in the compressor map of FIGURE 17.

Referring to FIGURES 17 and 18, the breadth ofthe operating range of the conventional turbocharger may be compared with that of the present invention.

Referring specifically to FIGURE 17, at a compressor pressure ratio of 1.9, the airflow range varies from a high of 350 at an efficiency of 60% to a low of 200 cubic feet per minute at the surge line. Therefore, this ratio of high to low flow rate (350 divided by 200) is 1.75. Referring to FIGURE 18, at a pressure ratio of 1.9, the present turbocharger operates from a flow rate of 275 at a 60% efficiency to a low of 60 cubicfeet per minute without encountering a surge line. The ratio of the high to low flow rate (275 divided by 60) is 4.58, an increase of overtwo and one-halfthatofthe conventional turbocharger.

This achievement is highly significant in that the turbocharger can be effectively used to improve performance both atlowerengine speeds swell as at higher engine speeds. This development becomes even more significant when coupled with the feature ofthe present invention for providing a constant compressor speed forvarying engine speeds or constant boost at varying engine speeds in that a higher desired pressure ratio may be attained at low engine rpm's withoutsurge orturbulence problems normally experienced atthe lower compressor flows in conventionalturbochargers.

This significant advancement provided by the turbocharger of the present invention is primarily attributabletothe compressor air inlet configuration in conjunction with advanced radial flow compressor technology-such as backward swept blading. As has been described, bearing support cylinder 60 is supported within tubular inlet port 40 by a plurality of struts 62 extending from the inside wall of inlet port 40 to the bearing support cylinder 60. In one embodiment, three struts extend from the inside wall ofthe inlet port to the bearing support cylinder to position the cylinder concentrically within the compressor air inlet. As a result, the inlet is divided into morethan one inlet channel.Additionally, these channels are proportionately long in the flow direc tion compared to any width dimension. Further, the struts have a contour including a leading and trailing edge narrower than an intermediate section of the strut. As is also shown in FIGURE 2, the inside wall portion of inlet port 40 has a converging diameter from the inlet end to the area adjacent compressor rotor 172.

This structure stabilizes the backflow conditions normally encountered at low compressorflows. By stabilizing the backflow conditions, the compressor is able to continue to function to provide a steady state flow of air at much lower compressor flows than heretofore possible in conventional turbochargers.

FIGURE 19 illustrates a vertical section of an alternative embodiment oftheturbocharger illustrated in FIGURES 1-16. The turbocharger of FIGURE 19 is identical to the turbocharger illustrated in FIGURES 1-16 with the exception that an axial stage compressor rotor 450 is mounted in the compressor inlet port. Because of the substantial identity between many components in the alternative embodiment illustrated in FIGURE 19 andtheturbocharger illustrated in FIGURES 1-16, like or correspond1ng parts are identified by the same numeral.

In the embodiment of FIGURE 1 9,the compressor inlet port is modified to include a larger diameter inlet port 452 having an inner wall 454 defining the inlet channels 456 to radial compressor rotor 172. A plurality of stators 458 extend from inner wall 454 of inlet port 452 to support bearing support460. An axial stage compressor rotor 450 is mounted on the end of shaft 160 on the side of bearing assemblies 162 and 164 opposite compressor rotor 172 and is retained in position byan appropriate nut462. The compressor rotor includes a hub 464 and a plurality of struts 466 attached between hub 464 and struts 468. Hub 464 engages a step 470 on shaft 160 and is positioned between step 470 and nut 462.

As can be appreciated from viewing FIGURE 19, air entering inlet port 452 is compressed by the axial stage compressor and is carried through channels 456 to radial compressor rotor 172. The use of a two-stage compressor is significant in that a broader operating range may be accomplished with the use of a two-stage compressor as compared to the range possible from a single stage compressor. For example, in use of a two-stage compressor, the first stage may be used to achieve a pressure ratio of inlet pressure to discharge pressure of 1 .5to 1.6 while the second stage may be used to achieve a pressure ratio of 3.0 to 3.5. These combined pressure ratios effectively produce a pressure ratio on the order of 4.5 to 5.6.While such pressure ratios may be achieved in a singlostage,the higher compressor speed required to produce such pressure ratios would dictate the use of higher quality and more expensive compressor and turbine components as well as producing signifi cantflow range restraints. In the present arrange ment, the two-stage unit accomplishesthe higher pressure ratios without this additional expense or disadvantage. Moreover, in the present invention, the axial stage compressor is mounted on the opposite side ofthe bearing support structure from the radial compressor and turbine rotors thereby providing balance to the system and no particular ill effects.

FIGURES 20-23 illustrate a novel way of mounting the turbocharger illustrated in FIGURES 1-16 on an internal combustion engine. FIGURES 20 and 21 illustrate a vertical section view and top plan view, respectively, ofturbocharger 20 mounted on a V-6 engine 444. Of course, itwill be understoodthatthis same application can be made on a "V" engine of any number of cylinders. Engine 444 represents a conventional V-arrangement engine having a right cylinder bank445 and a left cylinder bank 446. The section shown in FIGURE 20 illustrates a section through an exhaustvalve 448 in the right cylinder bank445 and through an intakevalve 449 on the left cylinder bank 446. Turbocharger 20 is mounted with the turbine and compressor rotational axis in a vertical position, an orientation not heretofore possible for conventional turbochargers.Because of the provisions in the present turbocharger for providing oil to the bearing assemblies by way of wicks andthe corresponding construction which eliminatestho need for a con tinuous flow of lubricantto the bearings, the present invention may be oriented with the compressor and turbine rotational axis in a true vertical orientation or any angle between vertical and horizontal. By contrast, conventional turbochargers, normally requiring a flow of lubricantto and from the bearing structures, have not been successful in operating in a true vertical orientation as illustrated in the arrangement of FIGURES 20 and 21.

Referring still to FIGURES 20 and 21, engine exhaust is channeled to turbocharger 20 byway of manifold 451 wherein exhaust gases are communicated directly to turbine rotor 170 of tu rbocharger20 from each individual exhaust valve. Turbocharger compressor exhaust isfed into the cylinders of engine444through manifolds 447 bywayofcross over tubes 453 to intake valves 449.

In operation, air is drawn into turbocharger 20 at inlet 50, is compressed by compressor rotor 172, and discharged through tubes 453 to manifolds 447 into the cylinders on both the left and right banks of the engine. Exhaustgasfrom the engine is delivered by way of manifold 451 to turbine rotor 170 at spaced points therearound and thereafter discharged through exhaust duct 455. This spacing allows the turbine to utilizethe high energy pulses created when the exhaust valve first opens.

As can be appreciated from viewing FIGURES 20 and 21,theturbine rotor is mounted in a turbine housing having a circumferential chamber with the variable nozzles therein aligned substantially in a plane perpendicularto the axis of rotation of the turbine rotor. Further,theengine exhaustports are aligned in a common plane oriented horizontally. As can be seen the normal turbine housing commonly used is replaced by manifold 451 which serves as both exhaust manifold and turbine housing.The abilitytousethepresentinvention in a vertical orientation makes possible the alignment of the plane ofthe exhaust ports substantially parallel to the plane ofthe circumferential chamber of the turbine housing by positioning the turbine shaft axis substantially perpendiculartothe plane ofthe exhaust ports.

In this way, the exhaust gases from the exhaust ports may be more directly and effectively injected to the turbine housing for driving the turbine rotor.

In contrast, priorturbocharger units have required the use of manifolds having extremely complicated designs for directing exhaust gases from the exhaust ports to the turbine rotor Further, in the present invention, the exhaust manifold directs exhaust gases from each ofthe exhaust portsfrom the engine to pointssubstantiallyequallyspacedaroundthe turbine rotor,thus communicating exhaust gases to different points about the circumference ofthe turbine rotor. In this way,the cyclic exhaust pulses from the engine exhaust ports are distributed about the circumference ofthe turbine rotor to take advantage of these pulses.

The vertical orientation of the presentturbochargeralso makes possible the nesting oftheturbocharger between the banks of the cylindersin close proximitytothe engine. Further, the exhaust gases may be ported byway of exhaust manifold directly from the cylinders to the tu rbochar gerturbine with very little loss of energy prior to application ofthe exhaust gases to drive the turbine.

Itwill be understood that a carburetor may be mounted directly above the turbochargerto feed a fuel air mixture into inlet 50. Likewise, the turbochar- ger may be used on a fuel injection or diesel engine wherefuel isfed directly into the engine cylinders.

Referring nowto FIGURES 22and23,thetur bocharger ofthe present invention mounted in a vertical position is shown in use on an in-linefour cylinder engine. Ofcourse, it will be understood that the invention also envisions the use ofthe present turbocharger on an in-line engine of more or less than four cylinders. In the embodiment illustrated in FIGURES 22 and 23, turbocharger 20 is connected to engine480 by way of an exhaust manifold 482 communicating between the exhaust valves ofthe engine cylinders to the turbochargerturbine. Air is drawn into the turbocharger at inlet 50, compressed by turbocharger compressor rotor 172 and dis charged through manifold 484 into the intake valves ofthe engine cylinders.As can be seen in FIGURE 23, exhaust manifold 482 is formed such thatthe gases driving theturbochargerturbine are communicated to the turbine at spaced points aboutthe turbine circumference. Further, the manifold is oriented to facilitate rotation oftheturbine in a clockwise direction as viewed from FIGURE 23. Similarly, as can be seen from FIGURE 21, a similar manifold arrangement is provided when the turbocharger is mounted intermediateoftwo banks of cylinders in aVarrangement engine.

Therefore, the present invention provides a turbochargerwhich can be oriented with the axis of rotation oftheturbine and compressor vertical or at any position intermediate of horizontal and vertical.

This is made possible largely because ofthe provisions in the present invention for lubricating the bearing structures on which the compressor and turbine shaft rotate. Because the present invention may be so oriented, it is ideally suited for mounting to and use with conventional engines of all types and arrangements.

The present turbocharger configuration described with respectto FIGURES 1-16 may be readily converted into an effective turbojet ortu rbofan engine as illustrated in FIGURES 24-28. While these embodiments are illustrated by a vertical section of only the upper half of apparatus, itwill be understood thatthe lower half of the unit is substantially identical to the upper half shown.

Aturbojet 500 embodying the present invention is illustrated in FIGURE 24. Turbojet 500 includes an outer housing 502 consisting of a tubular inlet nozzle 504, main housing 506 and combustion chamber backwall plate 508. Backwall plate 508 includes an integral stiffener 510 and is attached to main housing 506 by bolts 512. Aturbine exhaust nozzle 514 is supported within main housing 506 by attachment to the end of backwall plate 508 opposite main housing 506 by bolts 515.

A bearing support cylinder 516 is mounted within tubular inlet nozzle 504 by a plurality of struts 518 extending from the inside wall surface 520 of inlet nozzle 504. This arrangement issimilarto that illustrated in FIGURES 1with respecttothe bearing support cylinder 60 and its attachment within inlet port 40 of turbocharger 20. A cap 522 is mounted over the end of bearing support cylinder 516.

Referring still to FIGURE 24, a compressor backwall 524 and turbine backwall 526 are positioned within main housing 506 and are attached one to the other by suitable screws 527. In the embodiment illustrated in FIGURE 24, a fixed statorvane 528 is mounted between main housing 506 and compressor backwall 524 by suitable screws 529. A combustion liner 530 is mounted between turbine backwall 526 and a trans verselyoxtending portion of turbine exhaust nozzle 514. Combustion liner 530 may be of a conventional design including a circumferential chamber having a plurality of portsformed therein. The liner includes ports 532 for primary air combustion and ports 534 for dilution air according to known practice.Fuel is supplied to the combustion chamber by fuel spray nozzles 536 extending through combustion chamber backwall plate 508 and upstream wall 538 ofthe combustion liner. An ignitor 540 is received within combustion liner 530 th rough upstream wall 538 and combustion chamber backwall plate 508. Combustion liner 530 has a mouth 542 leading to a turbine inlet area 544formed between turbine backwall 526 and the transversely extending portion ofturbine exhaust nozzle 514.

Aturbine inlet area control structure 546 is used in turbojet 500 illustrated in FIGURE 24 and is identical to turbine inlet area control structure 228 illustrated and described with respect to the turbocharger of FIGURES 1-16. More specifically,turbine inlet area control structure 546 includes movable nozzle vanes 548 positioned circumferentially about the nozzle area and rotatabletovaryvelocityand angleofflow of combustion gasthrough nozzle inlet 544. Vanes 548 includetrunnions 550 and 552 extending from opposite sides thereof. Trunnion 550 extends through the transversely extending portion ofturbine exhaust nozzle 514, and trunnion 552 extends through turbine backwall 526 and is attached to an actuation lever 554.The end of actuation lever 554 is engaged with the outer race of a ball bearing assembly 556, with the inner race ofthe ball bearing assembly being secured to compressor backwall 524 as discussed with respectto turbocharger 20 illustrated in FIGURES 1-16. Likewise, the operation of actuation lever 554 is identical to that discussed with respect to turbocharger 20 of FIGURES 1-1 to control the angular orientation of nozzle vanes 548 as desired.

Bearing support cylinder 516 is supported concen tricallywithin inlet nozzle 504, as previously described, by struts 518. Acompressor rotor 570 and a turbine rotor 572 are supported on a shaft 574 mounted for rotation in bearing support cylinder 516 by two ball bearing assemblies 576 and 578. Turbine rotor 572 is mounted at one end of shaft 574 and radial flow compressor rotor 570 is mounted intermediate ofturbino rotor 572 and bearing assemblies 576 and 578. Shaft 574 passes through turbine backwall 526 and through compressor backwall 524 wherein a labyrinth seal 580 formed around shaft 574 provides a seal between the compressor and turbine rotors.

Turbine rotors 572 is fixedly attached to shaft 574, such as by welding, and compressor rotor 570 is retained in position on shaft 574 by a retainer nut 582.

Nut 582 is ofthe same arrangement as nut 180 described with respect to the embodiments il lus- trated in FIGURES 1-16. Compressor rotor 570 is positioned between nut 582 and a step 586 in shaft 574.

Referring still to FIGURE 24, a bearing ring 590 is fitted within the end of cylinder 516 adjacent compressor rotor 570 and is prevented from moving within cylinder 516 by a retaining ring 592 attached to cylinder 516. Outer raceway 594 of bearing assembly 578 is formed in ring 590, the inner raceway 596 being formed integrally in shaft 574. Balls 598 are received between inner and outer raceways to form bearing assembly 578.

A bearing supportcylinder516 is mounted within inlet nozzle 504 by a plurality of struts 518 extending from the inside wall surface 520 of inlet nozzle 504.

This arrangement is similarto that illustrated in FIGURES 1-16 with respect to bearing support cylinder 60 and its attachment within inlet port 40 of turbocharger 20. A cap 522 is mounted over the end of bearing support cylinder 516.

Referring still to FIGURE 24, a compressor backwall 524 and turbine backwall 526 are positioned within compressor-burner housing 506 and are attached onetothe other by screws 527.

Bearing assembly 576 includes inner raceway 600 formed integrally in shaft 574 and an outer ring 602 slidablewithin cylinder 516 with an outer raceway 604formed therein for receiving balls 606. A com pression spring 610 is engaged between ring 602 and a retaining ring 61 2fixed within cylinder 516 and biases ring 602 outwardly to fix the position of balls 606 and 598 in bearing assemblies 576 and 578, respectively, thereby fixing the position of shaft 574.

The structure ofthe bearing assembliesforthe turbojet 500 is identical to that arrangement used for turbocharger 20 illustrated in FIGURES 1-16. There- fore, the assembly and operation of the bearing assemblies is identical to that described earlierwith respect to turbocharger 20. Likewise, the means of lubricating the bearing assemblies, wicks in contact with inclined ramps 614, is identical to that illustrated with respect to turbocharger 20 illustrated and described in FIGURES 1-16.

In operation oftheturbojet illustrated in FIGURE 24, airenterstubular air inlet port at its inlet 620 following the air path indicated by arrow 622. This air is compressed by compressor rotor 570 and is discharged radially therefrom along the air path indicated by arrow 624. This compressed air is delivered into a combustion chamber area 626 and into combustion liner 530 through ports 532 and 534. Air is mixed with fuel, injected into combustion liner 530 through nozzles 536. A stochiometric mixture is obtained and the mixture is ignited within combustion liner 530 by ignitors 540. Cooling air is delivered into combustion liner 530 through ports 534to cool the exhaust gases from combustion priortotheir entry into turbine inlet area 544.These gases engage turbine rotor 572 to drive the rotor and the compressor rotor attached thereto by way of shaft 574. The combustion gases are thereafter discharged through nozzle 514tothe outer atmosphere, producing the thrust provided by the turbojet.

As can be appreciated by review of the turbojet illustrated in FIGURE 24, the system is compact in size and has a minimum of components. Further,the bearing assemblies on which the turbine and compressor rotor shaft rotate are positioned away from turbine rotor 572 and the extremely hot combustion gases delivered thereto from combustion liner 530.

Moreover, the serious problems encountered as a result ofthe extremely high temperatures such bearing structures are normally subjected to are completely eliminated by the unique bearing arrangement disclosed in the present invention.

Because the bearings are removed from the extremely high temperatures associated with the combustion gases used to drive theturbine and indeed even insulated from these high temperatures bythe compressor rotor and compressor backwall and turbine backwall, limited lubrication is sufficient for the successful operation ofthe unit. This eliminates the need for a constantflow of lubrication to the bearings as well as the numerous components necessaryto provide such lubricationthatwould have to be driven by the rotating shaft.

FIGURE 25 illustrates a turbofan 650. The structure ofturbofan 650 is substantially identical to the structure of tu rbojet 500 illustrated in FIGURE 24 with the exception that a bypass air channel is provided in addition to the primary air channel and an axial stage compressor is provided upstream ofthe radial compressor.

Referring to FIGURE 25, turbofan 650 includes an inlet nozzle 652, a main housing 654 and a combustion chamber backwall 656. Combustion chamber backwall 656 is attached to main housing 654 by appropriate bolts 658. Aturbine exhaust nozzle 660 is attached by bolts 662 to combustion backwall 656. A compressorbackwall 664andaturbinebackwall666 are positioned within main housing 654, and a combustion liner 668 is mounted within a combustion chamber 670 with a mouth 672 opening between turbine backwall 666 and a laterally extending portion 674 of turbine exhaust nozzle 660. Combustion liner 668 is formed with a plurality of ports 676. Afuel injection unit 678 is received through combustion backwall 656 into liner 668. An ignitor 680 is also received within liner 668 for igniting fuel air mixtures in operation ofthoturbofan.

Astatorvane 682 is mounted between main housing 654 and compressor backwall 664 by suitable screws. A bearing support cylinder 684 is mounted within inlet nozzle 652 by a plurality of struts 686. A plurality of stator vanes 688 are also mounted in the inlet nozzle 652 for cooperation with the axial stage compressor as will be discussed hereinafter in greater detail.

Shaft 690 is supported within bearing support cylinder 684 by bearing assemblies 692 and 694.

Bearing assemblies 692 and 694 are identical to those described with respect to turbojet of FIGURE 24 and theturbocharger illustrated in FIGURES 1-16, and reference is hereby made to the earlier description thereof. An oil chamber 696 is formed between inner wall 698 of inlet nozzle 652 and bearing support cylinder 684. An appropriate ring 700 with O-rings 702 and 704 are positioned in the mouth of chamber 696 and retained in position by ring 706. Wicks 708 and 710 communicate through oil chamber 696 to provide oil to inclined ramps 712 and 714, respectively,thereby providing lubrication to bearing assemblies 692 and 694 as described earlierwith respect to the embodiments of FIGURES 1-16 and FIGURE 24.

A radial compressor rotor 716 and turbine rotor 718 are mounted on shaft 690 and a turbine inlet control structure 720, identical to that discussed with respect to the embodiments of FIGURES 1-16 and FIGURE 24, is incorporated to control the velocity and angle of combustion gasestoturbine rotor718.

A bypass flow channel 722isformed between an inner bypass wall 724 and outer bypass wall 726.

Outer bypass wall 726 is attached to inner bypass wall 724 by a plurality of stator vanes 728 positioned therebetween. Outer bypass wall 726 is positioned from main housing 654 by a plurality of struts 730 spaced circumferentially around housing 654.

An axial stage compressor rotor 732 is attached to shaft 690 by an appropriate nut 734 engaging rotor 732 against a retaining ring 736. Rotor 732 includes a hub 738 and a plurality of vanes 740 extending radially therefrom. A plurality of cylindrical arma tures742 are embedded within hub 738 of rotor732 and cooperate with a stationary field winding 744 mounted adjacenttheretofrom outer bypass wall 726 by struts 746 to produce electric current.

The operation of the turbofan illustrated in FIGURE 25 is substantially similarto that of the turbojet illustrated in FIGURE 24 with the exception that air entering the inlet oftheturbofan at axial stage compressor rotor 732 is divided into a bypass flow indicated by arrow 747 and a primary flow indicated by arrow 748. Bypass air is com pressed and dis- charged through a bypass nozzle 749 producing thrustfrom the turbofan. Primary air is directed along the path indicated by arrow 748 and is compressed by axial stage compressor rotor732 and radial compressor rotor 716 prior to being mixed with fuel in combustion liner 668 and ignited therein.The ignited gases move pastthe turbine inlet area control structure and againstturbine rotor 718 to drivethe turbine rotor, as well as the axial stage compressor rotor732and radial compressorrotor716. Exhaust gases from turbine rotor 718 are discharged through turbine exhaust nozzle 660 producing additional thrust to theturbofan.

The turbine inlet control structure 720 is identical to that illustrated and described with respectto tur bocharger20 illustrated in FIGURES 1-16 and turbojet 500 of FIGURE 24. This control structure operates to control both the velocity and angle of combustion gases directed againstturbine rotor 718 from com bustion liner 668. It will be understood that the monitoring structure illustrated and disclosed with respectto FIGURES 6 and 7 relating to turbocharger 20 may be directly applied to theturbojet and turbofan of FIGURES 24 and 25, respectively. Thus, the speed ofthe turbine and associated compressors may be monitored and the turbine nozzle inlet area varied in accordance with this speed or in accordance with other engine parameters as desired.

During rotation of axial stage compressor rotor 732, armatures 742 are rotated relative to windings 744, causing the generation of electricity therefrom.

This electricity may be used to operate various engine or aircraft systems as desired and provides a very economical and compact energy package forthe system.

Referring now to FIGURES 26 and 27, an alternative embodiment of the turbojet illustrated in FIGURE 24 is shown and identified generally by the numeral 750.

Turbojet 750 includes an inlet cylinder 752 joined to a main housing 754 with a back plate 756 attached to main housing 754 by a plurality of bolts 758. Aturbine exhaust nozzle 760 is attached to the end of back plate 756 opposite main housing 754 by appropriate bolts 762. Turbine exhaust nozzle includes an exhaust cylinder 764 and a laterally extending portion 766 formed to join main housing 754.

A bearing support cylinder 768 is supported from inlet cylinder 752 by a plurality of stators 770. A shaft 772 is supported within cylinder 768 by bearing assemblies 774 and 776. Bearing assemblies 774 and 776 are identical in structure and operation to bearing assemblies 576 and 578 illustrated with respect to the embodiment of FIGURE 24. Aturbine rotor 778 is formed integrally with shaft 772 and a compressor rotor780 is mounted on shaft 772 intermediate of turbine rotor 778 and bearing assemblies 774 and 776.

A compressor backwall 782 and a turbine backwall 784 are mounted within main housing 754. A plurality of statorvanes 786 are mounted between main housing 754 and compressor backwall 782 by screws 788. A suitable ignitor789 is mounted within lateral extending portion 766 ofturbine nozzle 760 into the chamber formed between that wall and turbine backwall 784.

Aturbine nozzle inlet area control structure 790 including a movable vane 792 supported for rotation on trunnions 794 and 796 within transverse wall 766 and turbine backwall 784, respectively, is provided.

This turbine nozzle inlet area control structure 790 is identical to that structure disclosed and illustrated in the embodiment ofturbocharger 20 of FIGURES 1-16 and turbojet 500 of FIGURE 24. A labyrinth seal 800 is formed within turbine backwall 784 to provide an effective seal between the turbine and compressor areas. Compressor rotor 780 is mounted on shaft 772 by nut 802 engaging rotor 780 against a step 804 in shaft 772.

An axial compressor 806 rotates with shaft 772 to cooperate with stators 770 to compress air directed into compressor inlet 808 formed by inlet cylinder 752. Compressed air is directed along a path indi cated by arrow 810. In the embodiment illustrated in FIGURES 26 and 27, shaft 772 is formed with an axial bore 812therein. Afuel supply coduit814 is connected with bore 812 to provide a supply of fuel within the axial bore. An appropriate seal 816 is positioned between fuel supply conduit 814 and the surface of bore 812 to prevent loss of fuel atthe point of connection of the conduit and the shaft.

A plurality of radial ports818areformedthrough shaft 772 to communicate between bore 812 and an annular groove 820 formed in compressor rotor 780.

A plurality of radial ports 822 communicate from groove 820 through compressor rotor 780. Suitable seal means such as O-rings 826 and 828 are mounted in annulargrooveswithin rotor780to prevent leakage of fuel atthe point of connection between the rotor and shaft 772.

In operation of the turbojet illustrated in FIGURES 26 and 27,fuel is supplied bore 812 to 2 of shaft 772 through fuel supply conduit 814. Afuel supply path is provided between bore 812 and the airflow path indicated by arrows 810 (FIGURE 26)through ports 818, groove 820 and ports 822. This fuel flow path is indicated by arrows 838. As compressor rotor 780 is rotated, fuel is carried by centrifugal force from bore 812 through this fuel path into the air path indicated by arrows 810.Simultaneously therewith, air is drawn into inlet port 808 by axial stage compressor 806 and is delivered in a compressed state along the airpath.Thisairisfurthercompressed by radial compressor rotor 780 and mixed with the fuel downstream ofthe compressor rotor. The fuel is highly atomized as a result ofthe action ofthe compressor rotor blades on the fuel as it is ejected from the ports 822 through the compressor rotor.

This atomized fuel when mixed with the compressed air delivered into the turbojet is directed along the air path and is ignited by ignitor 836 upstream ofturbine rotor778.The exhaust gases resulting from combustion move past the vanes ofturbine rotor 778 to drive the rotor in a conventional manner. These exhaust gases are then discharged through nonle 760 to provide thrust from theturbojet. It will be understood thatthe combustion temperature is considerably greaterwith this system of combustion and would require materials capable ofwithstan ding the greater temperatures.

The system of FIGURES 25 and 26 eliminates many ofthe components heretofore required in conventional turbojets. The unique fuel supplysystem illustrated in FIGURES 25 and 26 eliminates the need for a fuel pump and all associated hardware.The conventional fuel pump is replaced by a fuel system which automatically discharges fuel in an atomized state by projecting the fuel through a bore in the compressor rotor shaft and through ports in the compressor rotor itself into the air stream compressed by the rotor. The system of FIGURES 25 and 26 further provides a two-stage compression arrangement, wherein an axial stage compressor is applied upstream of the radial or centrifugal compressor.

This arrangement provides additional balance to the system by providing the axial stage compressor on the side ofthe bearing structures opposite that of the radial compressor.

Further, this arrangement positions the bearing structure remote from the turbine rotor and combus tionchamberimmediatelyupstream oftheturbino rotor. This arrangement insulates the bearing structure from the extreme temperatures experienced in this area of the system. Therefore, no oil system is required. Moreover, the units are adaptable to very small turbojets orturbofans.

FIGURE 28 shows an alternative embodiment of the turbojet illustrated in FIGURE 27to produce a turbofan 900. The structure of turbofan 900 is substantially identical toturbojet750 illustrated in FIGURE 27 with the exception that a bypass air channel is provided in addition to the primary air channel. The turbofan includes an inlet cylinder 902 joined to a main body904 and a backwall 906 attached to main housing904bybolts908.Aturbine exhaust nozzle 910 is supported from backwall 906 by bolts 912 with a laterally extending portion 913 of nozzle 910 in engagementwith main housing 904. A bearing support cylinder 914 is supported from inlet cylinder 902 by a plurality of struts 916. A compressor backwall 918 and turbine backwall 920 are supported within main housing 904.Aturbine inlet area control structure 921 identical to that disclosed with respect to structure 790 of the turbojet of FIGURE 26 is provided. A shaft 922 is supported for rotation within bearing support cylinder 914 by bearing assemblies 923 and 924. These bearing assemblies are identical to those disclosed with respecttotheturbojet illustrated and described in FIGURE 26. Aturbine rotor 926 is mounted to one end of shaft 922 and a compressor rotor 928 is mounted intermediate of turbine rotor 926 and bearing assemblies 923 and 924.

In the embodiment illustrated in FIGURE 28, an axial compressorrotor930 is mountedtothe end of shaft 922 and directs air along a primary flow path within inlet cylinder 902 and a seconaryflowpath formed between outersecondaryflowwall 932 and innersecondaryflowwall 934. Outersecondaryflow wall 932 is supported from innerwall 934 by a plurality of stators 936. Air directed through the secondaryflow path, indicated by arrow 938, is compressed bythe action of compressor rotor 930 and stators 936 and is exhaustedthrough exhaust nozzle 940 to producethrustfrom theturbojet.Air directed along the primaryflow path, indicated by arrow 942, is compressed by rotor 928 and mixed with fuel supplied into the airstream in an identical way as described with respectto the embodimentof FIGURE 26 and ignited by an igniter 946. The combustion gases are directedthroughturbine inlet area control structure 921 to drive turbine rotor 926 and compressor rotors 930 and 928. Air exhausted pastturbine rotor 926 through turbine exhaust nozzle 910 providesfurtherthrustfrom the turbofan.

FIGURE 29 illustrates a further embodiment of the systems illustrated in the priorfigures wherein an electrical generator 970 is mounted within the compressor inlet nonle. Referring now to FIGURE 29, electrical generator 970 includes an armature 972 mounted on the end ofcompressorandturbine shaft 974. As is seen in FIGURE 27, shaft 974 is supported by bearing assembly 976 and a second bearing assembly (not shown) identical to the bearing assemblies shown with respect to the turbocharger of FIGURES 1-16 and the turbojets of FIGURES 24-28. A field winding 978 is mounted within bearing support housing 980, which is in turn concentrically mounted within compressor inlet nozzle 982 by vanes 984 in the same manner discussed earlier with respect to both the turbocharger and turbojet and turbofan embodiments. Acap 986 is mounted on the end of bearing support housing 980.

In operation of the generator, the armature 972 rotates with shaft 974 within field winding 978 to produce an electrical current in the well known manner. Electric current so produced is directed by way of leads (not shown) from field winding 978 to any point in the system requiring electrical power.

Thus, with respectto the present invention as used on a turbocharger, the generator may be used to supply electriccurrentfor any components in an automobile or other equipment requiring electrical power. With respect to application of the invention disclosed in FIGURE 29 to the turbojets of FIGURES 24-28, electrical energy may be used for guidance systems or other components requiring electrical power.

Thus,the invention disclosed in FIGURE29pro- vides a very simple and straightforward electric generator driven directly by the rotation ofthe compressorandturbineshaft. Moreover, the location ofthe generator provides a very compact system, light in weight and not interfering with the operation ofthe system from which it derives its power.

Therefore, the present invention provides a turbojet and aturbofan arrangement wherein the combustion chamber is removed from the bearing assemblies usedto supportthe compressor and turbine shaft. In another embodiment ofthe invention, the present invention incorporates an axial stage compressor on the opposite side ofthe bearing assemblies from a radial compressor. The invention also envisionsthe introduction offuel into the turbojet system by supplying fueld directly through the radial compressor rotorfrom the compressor rotor shaft. Inthisembodiment,thefuel is ejected by centrifugal force and a fuel pump and accompanying hardware is eliminated.

In the embodiment of the present invention providing for turbofan, a bypass flow and a primaryflow are provided. In accordance with one embodiment of this invention, an axial stage compressor is incorporated upstream and on opposite sides of the bearing assemblies from a radial compressor.

The present invention also envisions the incorporation of a generator including an armature mounted on the end ofthe compressor and turbine shaft opposite the bearing assembliesfrom the compressor and turbinowithafioldwinding mounted in the bearing support structure. Thisarrnngement provides a very simple and straightforward generatorfor providing electrical powertothe overall system.

Although preferred embodiments of the invention have been described in the foregoing detailed description and illustrated in the accompanying draw ings, it will be understoodthatthe invention is not limited the embodiments disclosed, but is capable of numerous rearrangements, modifications, and substitutions of parts and elemontswithoutdepart- ing from the spirit ofthe invention. The present invention is therefore intended to encompass such rearrangements, modifications and substitutions of parts and elements as fall within the scope ofthe appended claims.

Claims (1)

1. A compressor housing, for use in a turbo compressor arrangement, having: an inlet port, and a circumferential chamber surrounding said inlet port and leading to an exhaust port, said housing comprising: an outer housing defining the inlet port and the circumferential chamber around the inlet port, a wall insert, and means for attaching said wall insert within said outer housing with said wall insertforming one wall ofthe circumferential chamber, wherein, in use, with a compressor rotor mounted for rotation in the housing, said wall insert forms a circumferential passageway between the compressor rotor and the circumferential chamber, such that, upon rotation ofthe compressor rotor, air entering the inlet port is compressed, channeled through the passageway, discharged into the chamber, and exhausted through the exhaust port.

2. A compressor housing according to claim 1, wherein said wall insert comprises a tubularthroat with a disc attached at one end ofthetubularthroat transverse thereto, said disc having an aperture corresponding to an opening in the throat, and said throat is engageablewithin the inlet port when said insertis mounted to said outer housing.

3. A compressor housing according to claim 2, wherein the said attachment means comprises: rivet protrusions integrally formed and extending from the disc and corresponding to and facing apertures in the wall ofthe outer housing circum ferentially spaced around the inlet port, said rivet protrusions being adaptedforinsertion into said apertures to join said insertto said outer housing.

4. Acompressor housing according to claim 2 or claim 3, wherein said inlet porthasan inside surface with a first portion remotefrom said wall having a varying diameter converging toward the compressor rotor, and a second portion adjacentthe compressor rotor diverging in diametertowardthe compressor rotor, and wherein thetubularthroat of said insert has an insidosurfaceconvergingin diamotortowardthedisc to matewiththediverginginsidesurfacooftheinlet port of said outer housing.

5. A compressor housing according to claim 4 whereintheinsidesurface ofthethroatportion of said insert converges at substantiallythe same rate as the inlet port inside surface convergesto form a continuous converging surface from the end of the inlet port remote from the insert disc two the disc aperture.

6. A compressor housing according to any one of claims 1 to 5, wherein both said outer housing and said insertare die cast.

7. A compressor housing according to any one of claims1 to6,furthercomprising: a bearing support cylinder, and a plurality of vanes attached to and extending inwardlyfrom the inside wall ofthe housing inlet port, supporting said cylinder.

8. A compressor housing according to claim 7, wherein the vanes have a leading edge opposite, in use, said compressor rotor and a trailing edge facing, in use, said compressor rotor, said leading and trailing edges being separated byathickerintormedi- atevanesection.

9. Acompressorhousing according to claim 7 or daim 8when dependent on claim 2, wherein one end ofsaid tubularthroat is notched to correspond with saidvanessuch thatsaidthroat straddles said vanes when said insert is attached to said outer housing.

10. Acompressorhousing accordingtoanyone ofclaims 7 to 9, wherein three struts extend substantially radially from said cylindrical module to the inside wall ofthe housing, dividing the inlet port into three inlet passages 11, Acompressor housing according to any one of claims 1 to 10, wherein the inier diameteris shorter than the length ofthe inlet port

12. Acompressor housing, for use in a@urbo- compressor arrangement, substantially as hereinbefore described with reference to Figures 1,2,7, 11 10 16,19 and 25 to 28 of the accompanying drawings.

13. A method of producing a compressor housing according to any one of claims 1 to 12, comprises (i) die casting an outerhousing using an outer mold having on ita protrusion, an inner mold formed with a raised contcur to form the circumferential chamberand having a protrusion for mating with the protrusion on the outer maid for forming the inlet port, a core mold and a cap mold for cooperating with the core mold to form the exhaust port, (ii) die casting a compressor wall insert using a pair of cooperating molds having abutting surfaces which engage one another to produce the wall insert with a parting line on the outeredgeofthe wail, and (iii) attaching said wall insertwithin said outer housing to form the compressor housing

14. Amethod of producing acompressor hous ing, for use in a turbo-compressor arrangement, substantially as hereinbefere described with refer- enceto Azure 11 ofthe accompanying drawings New claims or amendments to claims filed on 28/6/83 Superseded claim 1-14 New or amended claims: 1.Acompressor housing with aninlet port and a circumferential chamber surrounding said in@ et port and leading to a compressor exhaustport, in which housing, in use, a compressor rotor is rotatably mounted for compressing air antering the inlet port and discharging the compressed air into the chamber for exhausithrough the exhaust port, the compressor housing comprising:: an outer housing defining the inlet port and the circumferential chamber around the inlet port, a wall insert in cluding a @ubular throat, for engage- ment within the inlet port, and a disc attached transversely at one end of the tubuler threat the disc having an aperture corresponding to the opening of the throat; and means for attaching said insert within said outer housing with said insert forming one wail ofthe circumferential chamber and forming a circumferen- tial passageway between the aperture and the circumferential chamber for channeling compressed air from the aperture to the chamber; wherein: The inlet port has an inside surface with a first portion remote from the disc having a varying diameterconverging toward the disc, and a second portion adjacentthe disc diverging in diameter toward the disc; and the tubular throathas an inside surface converging in diametertoward the disc and an outside surface diverging in diametertoward the disc, the outside surface being arranged to mate with the second portion ofthe inside surface ofthe iniet port of the outer housing.

2. Acompressorhousing according to claim 1, wherein said attachment means comprises: rivet protrusions integrally formed and extending from the disc and corresponding to and facing apertures in the wall of the outer housing circumferentially spaced around the port, said rivet protrusions being adapted for insertion into said apertures to join said inserted said outer housing.

3. Acompressor housing according to claim 1 or claim 2, wherein the inside surface of the throat pertion of the wall insertconverges along a straight line at substantially the same rate as the first portion Of the inlet port inside surface converges to form a continuous converging surface from the end ofthe inlet port remote from the insert disct to the disc aperture.

4. Acompressor housing accordig to any one of claims 1 to 3, further comprising: a bearing supportcylinder: and aplurality of vanes attached to and extending inwardly from the inside surface of the inlet port, said vanes having a leading edge remotefrom the disc.

and a trailing edge facing the disc, the leading and trailing edges being separated bya thicker intermedi- ate vane section, the surfaces of said vanes converging from the thickest intermediate vane section to the leading and trailing edges.

5. Acompressor housing according to claim 4, wherein one end of the tubular throat is notched to correspond with the vanes such that the throat straddles the vanes when the insertis attached to the outer housing

6. Acompressor housing according to claim 4 or 5, wherein there are three vanes which extend substantially radially from said cylinder to the insIde wall of the housing, dividing the in let port into three inlet passages.

7. Acompressor having according to any one of claims 1 to 6, wherein the in let port is connected to the circumferential chamber by a transverse wall.

8. Acompressor hcusing according to any one of claims X to 7, wherein the inlet diameter is shorter than the length of the inlet port.

9. Acompressor housing according to any one of claims 1 to 8, wherein both the outer housing and the wail insertare die cast lt. Acompressor housing, substantially as he- reinbefore described with reference to Figures, 1,2,7, 11 to 16 and 19 of the accompanying drawings.

11. Amethod of producing a compressor housing according to anyone of claims 1 to ID, comprising --Ii) die casting an outer housing using an outer mold having on ita protrusion. an inner mold formed with a raised contourtoform the circumferential chambre and having a protrusion for mating with the protrusion on the outer mold forfonning the inlet port, a core mold and a cap mold for cooperating with the core mold to form the exhaust port, (ii) die casting a compressorwall insert using a pairofcooperating molds having abutting surfaces which engage one anotherto produce the wall insert with a parting line on the outer edge of the wall, and (iii) attaching said wall insert within said outer housing to form the compressor housing.

12. A method of producing a compressor hous ing, for use in a turbo-compressor arrangement, substantially as hereinbefore described with refer ence to Figure 11 of the accompanying drawings.