Abstract:
The present invention provides a turbo heater which utilizes a gas turbine engine and a heat exchanger assembly. The gas turbine engine is adapted to efficiently operate over a prolonged period of time and at varying power outputs without adverse or detrimental effects to the components thereof. For example, the gas turbine engine includes bearing assemblies and a fuel delivery systems which are uniquely designed for the demands of repeated cycling (i.e. starting and stopping), as well as operation at various power outputs without damage to the gas turbine engine. In addition, the use of exhaust gas from the gas turbine engine eliminates direct impingement of combustion on the heat exchanger element, thereby significantly increasing the durability and life span of the turbo heater.

Description:
This application is a continuation of prior application Ser. No. 09/457,224, filed Dec. 8, 1999, now U.S. Pat. No. 6,161,768 of application Ser. No. 09/152,425 filed Sep. 14, 1998 which was patented on Jun. 13, 2000, U.S. Pat. No. 6,073,857. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to a gas powered co-generator, and more particularly, to a self-sustaining co-generator which utilizes a micro gas turbine engine and heat exchanger for generating heat and rotary drive power. 
     BACKGROUND AND SUMMARY OF THE INVENTION 
     Micro gas turbine engines are well-known in the art and have found particular utility in powering projectiles such as a missile. In this application, the turbine engine is operated in a substantially constant environment and for a relatively limited duration. In addition, the turbine engine is typically cycled once (i.e. started once and stopped once) and operated at a near maximum output generation. As such the operating conditions are substantially constant and well defined. Furthermore, the internal components of the gas turbine engine, such as the bearings, are not subjected to repeated cycling through a range of operating speeds. 
     However, to date, micro gas turbine engines have not proven useful in applications where the engine is required to operate in a variety of environments over a prolonged period of time at less than maximum output generation. A primary difficulty has been the inability to properly cool and lubricate the bearing assemblies. Thus, micro gas turbine engines have not been used in applications which require repeated cycling and/or operation in many different environments such as co-generation. 
     Instead co-generators have been developed which use separate heat generating and power generating sources. For example, gas heaters or furnaces typically use a fuel which is delivered to a burn chamber where the fuel is ignited and a blower unit powered by an external power source which blows the heated air generated in the combustion chamber out of the heater. Accordingly, systems of this type do not take advantage of the heat by-product generated during power generation. 
     The present invention provides a self-sustaining system wherein a gas turbine engine functions as a power head for a co-generator to generate heat and rotary drive power for driving the fan of the heat exchanger, as well as the auxiliary components of the engine such as an electrical generator for charging a battery which operates the other components of the system. The overall concept of the present invention is to provide a co-generator which utilizes a micro gas turbine engine for both energy generating functions. 
     A primary object of the present invention is to provide a co-generator which is substantially smaller, and thus portable, than current systems for a given heat generating capacity. 
     A further object of the present invention is to provide a gas turbine engine as a power head for the generation of heat, thereby eliminating direct impingement of combustion on a heat exchanger element, and significantly increasing the durability and life span of the heater unit. 
     Another object of the present invention is to provide a quick-starting, self-sustaining co-generator which is rugged in design and has the ability to operate in adverse locations (e.g. on temporary platforms) and in adverse conditions (e.g. sub-zero temperatures). 
     An additional object of the present invention is to provide a self-contained co-generator in which at least a portion of the gas turbine engine is coaxially located with and surrounded by an annular heat exchanger such that the gas turbine engine is protected and muffled. 
     A further object of the present invention is to provide an extremely high efficiency co-generator which converts a high percentage of the energy of the fuel. 
     Still a further object of the present invention is to provide a bearing assembly for a gas turbine engine which is effectively cooled and lubricated by the combustion fuel such that the gas turbine engine can be repeatedly started and stopped, as well as operated at various power outputs without damage to the bearing assembly. 
     In a preferred embodiment, the present invention generally includes a gas turbine engine having a combustor which is coaxially arranged with the compressor and turbine such that the micro turbo heater housing surrounds the gas turbine engine. The turbo heater further includes a heat exchanger assembly which may take the form of a simple housing, an air-to-air heat exchanger, an air-to-liquid heat exchanger, a catalytic converter, or any combination thereof. In this manner, the turbo heater can be used to generate a heated air supply, a heated water supply or both a heated water supply and a heated air supply, Furthermore, the carbon monoxide content of the heated air can be controlled depending on the particular application. 
     Further scope of applicability of the present invention will become apparent from the detailed description provided hereinafter. It should be understood however that the detailed description and specific examples, while indicating preferred embodiments of the invention, are intended for purposes of illustration only, since various changes and modifications within the spirit and scope of the invention will become apparent to those skilled in the art from this detailed description. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The present invention will become more fully understood from the detailed description and the accompanying drawings wherein: 
     FIG. 1 is a schematic view of a preferred embodiment of the turbo heater of the present invention in which a gas turbine engine is coaxially located within an annular heat exchanger; 
     FIG. 2 is a partial cross-sectional view of the turbo heater of the present invention schematically illustrated in FIG. 1; 
     FIG. 3 illustrates a modification to the turbo heater illustrated in FIG. 2 in which the heat exchanger is replaced by a catalytic converter; 
     FIG. 4 is a partial cross-sectional view of the gas turbine engine of the present invention illustrated in FIG. 2; 
     FIG. 5 is a detailed view illustrating the nozzle hub, rear bearing assembly and fuel slinger of the gas turbine engine illustrated in FIG. 2; 
     FIG. 6A is a cross-sectional view illustrating the passageways formed in the fuel slinger of the gas turbine engine illustrated in FIG. 2; 
     FIG. 6B is a circumferential projection of the fuel slinger illustrating the repeating pattern of passageways formed therein; 
     FIG. 6C is a detailed view of a group of fuel holes in fluid communication with the passageways; 
     FIG. 7 is a detailed view illustrating the interface between the shaft assembly of the gas turbine engine and the gear reduction assembly; and 
     FIG. 8 is a cross-sectional view of the heat exchanger of the present invention taken along line VIII—VIII in FIG.  2 . 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     The co-generator or turbo heater in accordance with the present invention is described in further detail with reference to a preferred embodiment. With reference now to FIGS. 1 and 2, the turbo heater  10  in accordance with the preferred embodiment includes a gas turbine engine  12  and a reverse-flow heat exchanger  14  which are supported within a frame assembly  16 . Gas turbine engine  12  draws ambient air through a compressor  18 , receives fuel from a fuel system  60  to form an air-fuel mixture, combusts the air-fuel mixture in a combustor  20  and discharges and expands the exhaust gases through a turbine  22 . As such, gas turbine engine  12  provides a source of heat as well as a source of rotary power. The rotating components of gas turbine engine  12 , namely compressor  18  and turbine  22 , are mounted on a common high-speed shaft assembly  24 . The shaft assembly  24  is coupled through a reduction gear assembly or gear box  26  to a generator set  28  including a pair of electrical alternators  28   a ,  28   b , an axial fan  32  and a starter motor  34 . The starter motor  34  is coupled to the axial fan  32  through a one-way over-running clutch assembly  36  which permits power transmission in a first rotational direction and free wheeling in a second rotational direction. A presently preferred one-way over-running clutch assembly is Clutch No. RCB-121616 available from The Torrington Company of Torrington, Conn. 
     With particular reference to FIG. 2, ambient fresh air is propelled into the turbo heater  10  between inlet struts  38  adjacent to the axial fan  32 . Diffuser blades  42  are disposed on a downstream side of the axial fan  32  adjacent to plenum  44 . Plenum  44  opens into inner shield  46 , as well as into the heat exchanger  14  and thus, inlet air is directed into heat exchanger  14  and also through ports  48  formed through inner shield  46 . Subsequently, the air which passes through ports  48  is consumed by the gas turbine engine  12 . 
     An exhaust diffuser duct  52  is connected to the gas turbine engine  12  and communicates via the exhaust header pipe  50  with the heat exchanger  14  which includes a plurality of clean air tubes  54  disposed within turbo heater housing  56 . The exhaust gases from gas turbine engine  12  pass between and over clean air tubes  54  and exit through exhaust collector ring  58  disposed on opposite sides of housing  56 . As presently preferred, the exhaust header pipe  50  is dimensioned such that the exhaust gases are discharged at a pressure slightly above atmospheric to facilitate the transfer of heat from the exhaust gases to the fresh air in the heat exchanger  14 . More specifically, the exhaust gas has a tendency to stick to the tubes  54  and increase the heat transfer efficiency. Fresh air is propelled through the interior of the tubes  54  by the axial fan  32  which is disposed on a forward end of the heat exchanger  14  and powered by the gas turbine engine  12 . 
     Turbo heater  10  is a diesel fueled self-contained and self-sustaining heating system for supplying heated air and electrical power in remote locations. Gas turbine engine  12  is designed to supply the majority of its energy as heat in the form of exhaust gases, and a minor amount as shaft power used to drive the axial fan  32  and the electrical generator set  28 . Heat exchanger  14  is used to recover the resulting heat in the exhaust gases. In this regard, turbo heater  10  may be equipped with a combination air-to-air heat exchanger  14   a  and an air-to-liquid heat exchanger or liquid coil  14   b  to supply both heated air and heated liquid as illustrated in FIG. 1, an air-to-air heat exchanger  14   c  to supply heated, breathable air as illustrated in FIG.  2 . If desired, the heat exchanger  14  could also be of a liquid coil type to supply only heated liquid. Alternately, turbo heater  10  may be fitted with a suitable catalytic converter  14   d  which reduces the carbon monoxide in the exhaust gases to supply essentially breathable heated air as illustrated in FIG.  3 . In certain limited circumstances where the air quality is not an issue, the turbo heater  10  may be configured such that heated air is provided directly from the exhaust of gas turbine engine  12  and blended with fresh air supplied by the axial fan  32 . 
     Turbo heater  10  is designed to feature economical construction and is especially designed for reduced manufacturing cost. The internal aerodynamics, such as the turbine and compressor wheels, use well-developed turbocharger technology. For example, the preferred flow and pressure ratios are nearly optimum for automotive turbocharger components, and are thus near-optimum for use in the turbo heater  10 . A peak cycle temperature of 1500° Fahrenheit (° F.) is preferred to allow the use of economical materials for the high temperature components. For instance, combustor  20  of gas turbine engine  12  is fabricated from aluminized steel which is less than a one-third the cost of high nickel sheet alloys typically used in gas turbines operating at higher temperatures. The 1500° F. gases are expanded in turbine  22  and exhausted into the heat exchanger  14  at a maximum of 1300° F. Therefore, the heat exchanger assembly  14  can be made from less expensive materials than those used in connection with direct fired units operating at temperatures up to 2500° F. without sacrificing durability. 
     Gas turbine engine  12  consists of a radial flow (centrifugal) compressor  18 , annular combustor  20  and radial flow turbine  22 . The compressor  18  and turbine  22  are attached to a common high speed shaft assembly  24  and the annular combustor  20  is located therebetween. More specifically, the high speed shaft assembly  24  passes through the center of the combustor  20  with the compressor  18  positioned on one side and the turbine  22  positioned on the other side thereof. With the shaft  24  rotating at approximately 125,000 revolutions per minute (rpm), the compressor  18  takes in ambient air at a rate of about 0.5 pounds per second (lbs/sec) or approximately 371 standard cubic feet per minute (scfm). Thus, at an ambient temperature of 32 degrees Fahrenheit (° F.), the inlet air is compressed to about 15 pounds per square inch gage (psig) above ambient pressure which is standard at 14.7 psia (absolute). As the compressor pressure is approximately twice the pressure of the ambient pressure, the compressor  18  is said to have a pressure ratio of approximately two to one (2:1). 
     The combustor  20  is supplied with pressurized air at a ratio of approximately two to one and diesel fuel is added to form an air-fuel mixture which is ignited therein. The air-fuel mixture burns steadily after the initial ignition and generates an exhaust gas having an elevated temperature preferably not exceeding 1500° F. The exhaust gas is discharged from the combustor  20 , expanded in the radial turbine  22 , and exhausted into the heat exchanger assembly  14 . When the hot gases are expanded at a pressure slightly less than the compressor discharge, enough power is created to drive the compressor  18  as well as to drive the axial fan  32  for delivering the air, and the electrical generator set  28  supplying turbo heater accessory power and surplus power for lights or tools. The rate of fuel flow provided by a fuel system  60  is used to modulate the output of the turbo heater  10 . For example, a fuel flow rate of approximately 32 pounds per hour (lbs/hr) generates approximately 500,000 British thermal units per hour (Btu/hr) of fresh heated air. The combustor  20  on the turbo heater  10  is different from those used in heating applications since the gas turbine engine  12  is designed to operate at a pressure twice that of atmospheric pressure and the velocity of the air entering the combustor  20  is created by a 1 psi pressure drop, yielding combustor mixing jet velocities over 200 feet per second (ft/s) which are significantly greater than a combustor typically used in heating application with an air supplied by a low pressure blower. These conditions increase the combustor burning intensity such that a reduced combustor volume, on the order of twenty-five percent to thirty-three percent (25%-33%) the volume of a conventional, atmospheric combustor/burner. This dramatic reduction in volume is characteristic of the turbo heater design. For example, a 500,000 Btu/hr turbo heater may be designed to fit into a 22″×22″×42″ frame and weigh less than 200 pounds (lbs), making it relatively portable (except for a remote fuel supply  62 ). 
     The high speed shaft assembly  24  which rotates on its mass center by means of elastically mounted bearings is vibration-free for all practical purposes. Therefore, the turbo heater  10  can be located onto temporary platforms and safely operated there. The fuel supply for an 8 hour day (about 50 gallons or 300 pounds) can be provided from a portable tank  62  or a ground level reservoir and pumped through a flexible hose as desired. These features allow the turbo heater  10  to be used in many ways not currently available for conventional heaters. The axial flow fan  32  is also designed to deliver breathable fresh air at a pressure which is sufficiently high enough to inflate moderately-sized portable structures having an approximate internal volume of 20,000 cubic feet. 
     A reduction gear assembly or gear box  26  is interdisposed between the high speed shaft assembly  24  and the other rotating components of the turbo heater  10  such that economical, well developed, efficient components can be used. With reference to FIGS. 4 and 7, gear box  26  includes a pinion gear  64  disposed on an end of the high speed shaft assembly  24 , a pair of counter-rotating gear sets  66 ,  68  and an output shaft  70 . As such, gear box  26  has two reduction stages which yield an final drive ratio of 15 to 1 (15:1), making the output shaft speed a nominal 8,333 rpm. As further described hereinafter, the pinion gear  64  is rotationally supported by an elastically mounted front bearing assembly  146  which provides load sharing between the counter-rotating gear sets  66 ,  68 . The gear box  26  is a self-contained unit having an oil pump  72  disposed therein. The output shaft  70  from the gear box  26  drives the electrical generator set  28 . As presently preferred, a small electric alternator  28   a , such as typically used in automotive applications, supplies 12 volts and 20 amperes (thereby consuming approximately 2000 Btu/hr) and produces the electrical power necessary to perform turbo heater functions such as charging the battery, powering the system control, operating the fuel pump and supplying the ignition energy. A larger electrical generator  28   b  produces up to 3 kilowatts of surplus electrical power (thereby consuming approximately 14,000 Btu/hr). The gear box  26  also drives the axial fan  32  that supplies the necessary fresh air flow at about 2000 cfm and at relatively low pressures about ⅓ psi (thereby consuming approximately 15,000 Btu/hr). Thus, the heat equivalent of the total shaft power amounts to less than 31,000 Btu/hr, which is approximately 6.2% of the total heat output of the turbo heater  10 . Thus, a feature of the turbo heater  10  is that approximately 97% of the energy in the fuel, less the equivalent shaft energy, enters the heat exchanger  14 , as compared to conventional systems which require auxiliary power that is not converted to usable heat. 
     With particular reference to FIGS. 2 and 8, the annular heat exchanger  14  is generally cylindrically shaped and formed by an annular bundle of tubes  54  such that the gas turbine engine  12  and the gear box  26  are contained inside the annulus defined by the heat exchanger  14 . In this manner, the critical dynamic components of the turbo heater  10  are both protected from damaging impacts thereto and contained to control the sound emanating therefrom. Exhaust from the gas turbine engine  12  enters a diffuser duct  74  and is directed into the hot gas side of the heat exchanger  14 . The heat exchanger  14  is a counter-flow type exchanger wherein the hot exhaust gas travels forward between the heat exchanger tubes  54  to an exhaust collector ring  58  located at the front of the heat exchanger  14  where it is released to the atmosphere. When discharged from the turbo heater  10 , the exhaust gases are relatively cool having given up over 80% of the available heat content. Fresh air driven by the axial fan  32  enters the tubes  54  at the front of the heat exchanger  14  and flows straight rearward, acquiring approximately 80% of the available heat. As best seen in FIG. 2, a collector  76  and suitable exhaust pipe  78  may be provided at the rear of the heat exchanger  14  to operably connect the turbo heater  10  into an air handling system. The resulting fresh, breathable heated air is supplied at a rate up to 500,000 Btu/hr and temperatures to over 250° F. and a pressure of approximately ¼ psig. The pressure and flow capabilities of the axial fan  32  generate a positive pressure differential between the fresh air in the tubes  54  and the exhaust gases surrounding the tubes such that the leakage direction is from fresh air into the exhaust flow. When the breathable fresh air exits from the heat exchanger  14 , the axial fan  32  will produce a pressure on the fresh air side of the heat exchanger  14  that is everywhere higher than the exhaust gas side thereof. This produces a heat exchanger that is inherently safe from possible leakage of dangerous exhaust gases into the breathable heated air. 
     With reference again to FIG. 3, a modification to the preferred embodiment, especially suitable for outdoor construction applications, is illustrated wherein heated air is produced using a catalytic element  80  located within the heat exchanger assembly  14   d  in which identical elements are indicated with identical reference numerals and similar elements are indicated with primed reference numerals. Since the combustor (not shown) on the gas turbine engine  12  produces significantly less carbon monoxide (CO) compared to a gasoline spark ignition engine, a properly fitted catalytic element  80  on the gas turbine engine  12  can reduce the emissions to acceptable levels. As such, the use of the catalytic element  80  greatly reduces the weight of the turbo heater  10  as well as the size and cost. The catalytic element  80  is fitted directly to the gas turbine engine exhaust  50 ′ by means of a diffuser duct  52 ′. The exhaust from the catalytic element  80  will be 1250° F. to 1300° F. maximum and the additional air flow from the axial fan  32  will pass around the catalytic element  80  and within the volume defined by housing  56 ′ for mixing and blending with fresh air to produce a relatively even discharge temperature of approximately 350° F. In this manner, the efficiency of the turbo heater  10  can approach 97%, depending upon the amount of electrical power being concurrently generated. 
     Alternately, some applications where human consumption of the heated air is not a requirement, a heat exchanger or catalytic converter may not be required, but the exhaust gas from the gas turbine engine  12  may be subsequently mixed with fresh air from the axial fan  32  to produce a heated mixture of exhaust gases and air. 
     With continued reference to all of the figures, the amount and temperature of the heated air produced by the turbo heater  10  can be controlled over a broad range by means of fuel flow adjustment, and back pressuring either the turbine  22 , the axial fan  32  or any combination thereof. Furthermore, a minor amount of modulation can be accomplished by adjusting the amount of electrical power produced by electrical generator set  28 . More specifically, the amount of fuel basically determines the operating speed of the gas turbine engine  12 , and thus the speed of the axial fan  32 . Once the initial speed is set, the minimum amount of fuel and heat is determined, as well as the maximum fan flow for this speed setting. Back pressuring the gas turbine engine  12  increases the amount of fuel necessary to hold the speed setting constant and increases the gas temperature entering the heat exchanger  14 , thereby raising the delivered air temperature. Back pressuring the heat exchanger  14  increases the fan pressure, thereby reducing the fan air flow and again increasing the delivered air temperature. Similarly, increasing the amount of electrical power generated requires more fuel to hold constant speed, thereby increasing the heat input to the heat exchanger  14  by a minor amount. In the preceding manner, a broad range of heat input, air temperature and air flow can be modulated for the desired operating conditions within the rating of the turbo heater  10 . A more conventional means of heat modulation or temperature control is to simply turn the turbo heater  10  on and off using a conventional thermostat which may be incorporated into a controller  300  or remotely located. 
     The turbo heater  10  is self-contained and nearly instantaneously starting, and will operate at a minimum heat output on a reasonable on-off cycle for lower heat requirements. Operation of the turbo heater  10  in this manner would provide an environment of uniform heat, using the minimum fuel necessary. As such, the turbo heater  10  is an ideal source of heated air as it can supply a large quantity of heat at relatively low ambient temperatures. For example, while nominally rated at 500,000 Btu/hr, the turbo heater  10  can be modulated from less than 250,000 Btu/hr to greater than 750,000 Btu/hr at an ambient temperature of minus 50° F. 
     Referring now to FIGS. 4-7, the details of the gas turbine engine  12  will be further discussed. As with all manufactured items, the components of the turbo heater  10  can only be manufactured to within a given tolerance. To accommodate variations due to these tolerances, nearly all gas turbine engines operating over  30 , 000  rpm must incorporate some form of elastic or damped bearing assembly for high speed shaft assembly  24 , to minimize the resulting bearing loads which would become prohibitive due to the inability to achieve a perfect balance on the turbine shaft. Thus, the shaft assembly  24  is rotatably supported on rear and front bearing assemblies  110 ,  146  which are elastically supported to allow the shaft assembly  24  to rotate substantially on its mass center, as determined by the balance tolerance. 
     Referring now to FIGS. 4 and 5, the high speed shaft assembly  24  is defined by turbine  22 , anti-rotation pin  86 , scavenger blower  88 , rear bearing inner race  90 , timing pin  92 , center shaft assembly  94 , front bearing inner race  96 , pinion gear  64  and tie nut  98 . The center shaft assembly  94  is located forward of the turbine  22  and the scavenger blower  88  and includes fuel slinger  100 , compressor  18  and front seal carrier  102 . The fuel slinger  100  is a press fit onto a pilot spigot  104  which is formed as a part of the compressor  18 . 
     The compressor  18  is piloted and pressed onto the front seal carrier  102 , and the front bearing inner race  96  is mounted on the front seal carrier  102 . The pinion gear  64  is fitted into the front bearing inner race  96 . The front seal carrier  102  and the pinion gear  64  are driven positively by means of mating shaft half lap joints  106  piloted inside the front bearing inner race  96 . All of the previously described components of shaft assembly  24  are bolted together by means of a long tie bolt  108  attached to the turbine  22  and secured by a tie nut  98  tightened against the pinion gear  64 . When assembled, the tie bolt  108  is sufficiently stretched such that the shaft assembly  24  acts as a rigid body. 
     High speed shaft assembly  24  is rotatably supported by rear bearing structure  110  which is disposed within nozzle hub  112  extending from turbine nozzle assembly  114 . The rear bearing assembly  110  includes rear bearing set  116 , fuel deflector  118 , belville spring  120 , rear bearing holder  122 , fuel shield  124 , rear beam spring assembly  126 , scavenger blower cover  128 , pick up plate  130 , turbine seal  132 , compression springs  134 , and retaining ring  136 . Rear bearing set  116  includes bearing cage  138 , rear bearing outer race  140 , and graphite rub ring  142  and ball bearing  144 . 
     An axial preload is applied to the rear bearing assembly  110  primarily by means of six compression springs  134  circumferentially disposed about the rear end (i.e. turbine end) of the shaft assembly to apply a force on the order of approximately  18  to  20  pounds. Additionally, two belville springs  120  locate the rear bearing set  116  in the rear bearing holder  122 . The combination of compression springs  134  and belville springs  120  in series urge the shaft assembly  24  forward such that front bearing assembly  146  engages front bearing shim plate  148 , thereby locating the position of the shaft assembly  24  in the gas turbine engine  12 . Providing a uniform axial load at high frequency is important to control axial vibration, which if induced, would cause an impact load on the rear and front bearing assemblies  110 ,  146 , thereby reducing their life. The amplitude of such an impact load is proportional to the vibrational amplitude which is controlled and greatly reduced by the axial loading provided in the present invention. The belville springs  120  and compression springs  134  define a spring means which is designed such that the natural frequency of the shaft assembly  24  in the axial direction greatly exceeds the rotational operation frequency of the shaft assembly  24 . Thus, if the spring response is not in excess of the operational axial frequency of the shaft assembly  24 , the outer bearing race  140  will separate from the ball bearing  144  and rear bearing inner race  90  and severely damage the rear bearing assembly  110  in a relatively short period of time. 
     As previously mentioned, the rear bearing set  116  is elastically supported by eight radial beam springs  126  disposed circumferentially between rear bearing outer race  140  and rear bearing holder  122 . The rear radial beam springs  126  must also respond at a frequency greater than the operating rotational frequency of the shaft assembly  24 . As presently preferred, the beam springs  126  are designed to have a preload of approximately 6 to 7 pounds and a spring rate between 6,000 and 7,000 pounds per inch resulting in a natural frequency of three to five times the operating speed of shaft assembly  24 . This combination of preload, spring rate and natural frequency is believed to have successfully minimized the radial impact loading on the rear bearing set  116 , while allowing the shaft assembly  24  to rotate on its mass center. 
     High speed shaft assembly  24  is also rotatably supported by front bearing assembly  146  which is disposed within the front bearing support  156  fixedly coupled to engine housing  158 . The front bearing assembly  146  includes front ball bearings  160 , front bearing outer race  162 , and front beam springs  164 . More specifically, the front bearing assembly  146  is supported by eight radial beam springs  164  disposed circumferentially between front bearing outer race  162  and front bearing support  156 . The elastic suspension provided by front beam springs  164 , like the rear beam spring assembly  126  of rear bearing assembly  110 , accommodates minor imbalances in the shaft assembly  24 . 
     Likewise, the pinion gear  64  is elastically supported by the front bearing assembly  146  which also provides an initial centering force. More specifically, the front end of shaft assembly  24  runs in its balance orbit, which is typically less than 2% of the total available radial travel. The remaining radial travel, approximately 98%, provides a means of load sharing between the counter-rotating gear sets  66 ,  68  which amounts to a load correction of about 4% of the maximum tooth load. The elasticity of the system and the resulting load adjustment assist in the smooth tooth action necessary for long life of high speed gears and also accommodate the dimensional tolerances necessary for manufacturing. The inertia of the front end of shaft assembly  24  and the resulting load correction force are such that the movement will take place over a number of cycles. The resulting small but important load sharing and its relatively slow adjustment work harmoniously to greatly reduce the tooth wear on these very high speed gears which provides a significant manufacturing cost reduction for this small high speed turbomachine. As presently preferred, front bearing assembly  146  is disposed within the gear box housing  166  such that cooling and lubrication is provided by oil pump  72  of gear box  26 . 
     The rear bearing set  116  is adapted for high temperature air, fuel, fuel vapor and limited liquid fuel lubrication and cooling, well beyond the limits of other bearing assemblies. The rear bearing set  116  features the use of commercial-grade inner and outer bearing races  90 ,  140  fitted with ceramic silicon nitride balls  144 , which are approximately fifty percent (50%) lighter than conventional tool steel balls. Therefore, the centrifugal load applied on the outer bearing race  140  by the balls  144  is similarly reduced. This lighter load allows the commercial outer bearing race  140  to survive the adverse conditions of high speed, high temperature and reduced lubrication. 
     The rotation of shaft assembly  24  under normal operation causes the rear bearing cage  138 , which rotates at a speed about one-half of the shaft speed, to move in a compound orbital motion (i.e., a series of smaller higher frequency orbits superimposed on the center of rotation of the shaft assembly  24 ) resulting in the potential for impact of the bearing cage  138  on the outer race  140  which would reduce performance and life to an unacceptable level. As will be further described hereinafter, rear bearing assembly  110  is cooled and lubricated by combustion fuel, namely diesel fuel or fuel oil. The bearing cage  138  functions as a centrifugal fluid separator of the air, fuel vapor and liquid fuel mixture delivered to the rear bearing set  116  for dynamically centering purposes. More specifically, liquid fuel provided to the rear bearing set  116  is centrifuged into an annular groove  150  formed in the bearing cage  138  which is in fluid communication with nine radial holes  152  located in bearing cage  138  such that high velocity streams of liquid fuel issue from holes  152  and have a tendency to center the bearing cage  138  in the bearing outer race  140 . When the bearing cage  138  approaches the outer race  140 , the radius from the center of rotation increases. This increased radius creates a higher pressure, thereby increasing the velocity of the fuel flowing through the holes  152 . 
     The resulting thrust exerted on one side of the bearing cage  138  is increased with a reverse effect occurring on the opposite side to reduce the thrust thereon. This thrust modulation stabilizes the movement of the bearing cage  138  by providing a centering force directed toward the center of rotation of shaft assembly  24 . The rear bearing set  116  further includes a graphite rub ring  142  disposed in a relief  154  formed in rear bearing outer race  140 . The graphite rub ring  142  provides a low friction, dry lubricating material and a non-galling surface in the event of contact by the rear bearing cage  138  such as during initial start-up when insufficient amounts of fuel are passed therethrough to generate the centering force. The operational life of rear bearing assembly  110  is significantly increased by providing a dynamically centered bearing cage and graphite rub ring such that the rear bearing assembly  110  performs successfully in the adverse environment of the gas turbine engine  12 . 
     As previously mentioned, the metered fuel used for combustion is directed through the rear bearing set  116 , for cooling and lubrication thereof. Since the fuel consumption of gas turbine engine  12  varies significantly depending on the operational parameters of the turbo heater  10 , the combustion fuel must be conditioned to insure adequate cooling and lubrication particularly during low fuel consumption periods such as idling. To this end, scavenger blower assembly  168  is located near the turbine  22  at the far end of high speed shaft assembly  24  and provides a high speed mixed flow (i.e. air, fuel vapor and liquid fuel) to rear bearing set  116 . While the scavenger blower assembly  168  is typically used to pump hot air, the design of the scavenger blower  88 , and more specifically the blades  170  formed thereon and the size of scavenger blower  88  relative to annulus  172  in which it operates, provides means of preventing liquid fuel and vapor from back flowing (i.e. flowing from the compressor-side to the turbine-side of the rear bearing assembly  110 ) once the shaft assembly  24  is turning, even at relatively low speeds. 
     Referring now to FIG. 5, eight pump impeller blades  170  are circumferentially disposed on the scavenger blower  88  and extend radially inwardly therefrom. The fuel shield  124  also has a plurality of vanes or fins  174  formed therein adjacent the scavenger blower blades  170 . More specifically, fins  174  are formed from the outer lip of the fuel shield and extend radially inwardly therefrom. The fuel shield fins  174  diffuse the rapidly rotating diesel fuel that tends to enter the annulus  172  at initial start up and direct it to the annulus  176  located between the blower hub  178  and the fuel shield  124 . The metered fuel for engine control is introduced into the rear bearing holder  122  by means of a fuel feed tube  180  into the nozzle hub  112  and acts as a sliding rotary joint between the bearing holder  122  and the nozzle hub  112 . An annular ring  182 , which is formed in the bearing holder  122 , extends radially inwardly between the fuel shield  124  and the rear bearing set  116  and serves to locate the rear engine bearing set  116 , thereby providing a spring stop for the belville springs  120  which biases shaft assembly  24  as previously described. 
     Fuel is communicated from the fuel feed tube  180  through a fuel passageway  184  formed in the nozzle hub  112  which feeds a groove  186  formed in the bearing holder  122  which communicates with a fuel passageway  188  formed through bearing holder  122  that terminates between the annular ring  182  and the fuel shield  124 . The annulus  176  is formed by the fuel shield  124  and the annular ring  182  in the center of these parts prior to entering the annulus  172  formed by the hub  178  of scavenger blower  88  and the inner diameter of both the fuel shield  124  and the annular ring  182 . Fuel entering scavenger blower assembly  168  is believed to wet the lip on the forward edge of the fuel shield  124  such that it is mixed and atomized by the air pumped by the scavenger blower  88 , thereby preventing passage of the air-fuel mixture past the scavenger blower cover  128 . 
     The scavenger blower  88  is sealed with a cover  128  and a pick up plate  130 , both of which are held in place with the compression springs  134  and the turbine seal  132  locked in place with a retaining ring  136 . As previously described, the compression springs  134  supply the necessary axial load for the angular contact of ball bearings  144  with inner and outer races  90 ,  140  that support the high speed shaft assembly  24 . The scavenger blower cover  128  seals the scavenger blower  88  and provides the necessary operating clearance. The scavenger blower cover  128  has an entry lip  190  extending radially inward for retaining the fuel within scavenger blower assembly  168  during start up and low speed operation. The pick up plate  130  further prevents the leakage of a minor amount of fuel from the scavenger blower cover  128  and from the clearance space between the rear bearing holder  122  and the nozzle hub  112 . More specifically, the outer diameter of the pick up plate  130  is closely matched to the bore  192  formed in the nozzle hub  112  such that a clearance space  194  is provided between the pick up plate  130  and the scavenger blower cover  128  to draw leakage fuel into the scavenger blower cover  128 . Furthermore, the scavenger blower cover  128  is beveled to create a pressure differential therebetween. This pressure differential draws the leakage fuel back into the scavenger blower assembly  168 , thereby preventing emission of unburned fuel in the turbine exhaust. In the running state, the cavity pressure surrounding the pick up plate  130  is higher than the pressure in the combustor  20  which tends to make the air and fuel flow forward into the combustor  20 , in addition to the effort from the scavenger blower  88 . 
     The scavenger blower  88  draws hot air through primary holes  196  formed in turbine nozzle assembly  114  into an annulus  176  formed by the bearing holder  122  and shaft assembly  24  where it mixes with fuel and is driven past the fuel shield  124  and the annular ring  182 . The fuel, which has been heated through contact with the heated metal surfaces of nozzle hub  112  is partially vaporized and mixes with the hot air to form a mixture of air, fuel vapor and hot fuel that is passed through the bearing set  116  such that the mixture subsequently lubricates and cools the rear bearing set  116 . 
     It should be noted that the cooling and lubrication of rear bearing set  116  occurs in various ways depending upon the operating state of the gas turbine engine  12 . In a first starting state associated with starting of the gas turbine engine  12 , a one-phase combustion fuel (liquid fuel) wants to back flow until the rear bearing set  116  warms up and the shaft assembly  24  is rotating in at least an idling range. During the engine starting state (and particularly during cold starting), the scavenger blower assembly  168  scoops liquid fuel tending to back flow from the scavenger blower  88  and drives it upwardly and through the rear bearing set  116 . In a second running state with the gas turbine engine  12  is running, the fuel flow quickly changes to a two-phase combustion fuel (liquid fuel and air) and then a three-phase combustion fuel (liquid fuel, fuel vapor and hot air) which tends to flow in a forward direction through the bearing set  116 . 
     The three-phase combustion fuel exits from the rear bearing set  116  and impinges upon the fuel deflector  118  which delivers the combustion fuel to the fuel slinger  100 , a rotating component of the shaft assembly  24 , such that fuel is centrifugally driven forwardly and outwardly through the inner cavity  200  of the fuel slinger  100 . A dam ring  202  is located on the fuel slinger  100  and a retainer/rear liner  204  is located on nozzle hub  112  to direct fuel, which might otherwise back flow into the exhaust gas, into the combustor  20  where it is consumed. The dam ring  202  is internally tapered to direct any fuel landing thereon forwardly into the fuel slinger cavity  200 . The nozzle hub  112  has a series of eight ( 8 ) secondary holes  206  drilled at an angle to pressurize and add swirl to the area immediately adjacent the retainer/rear liner  204 , thereby propelling any fuel/air mixture out of the clearance between the fuel slinger  100  and the retainer/rear liner  204 . 
     The fuel slinger  100  provides the final atomization and mixing of the combustion fuel, and more specifically functions as a rotary atomizer which intensely atomizes and cylindrically distributes the fuel across the axial gap between the front combustor liner  208  and the rear combustor liner  210 . Nine (9) passageways defined by axial bores  212  and nine (9) groups of three (3) fuel holes  214  are formed in fuel slinger  100  and extend forwardly (toward the compressor  18 ) and radially outwardly from inner cavity  200  relative to the longitudinal axis a—a of the fuel slinger  100 . The fuel slinger  100  has nine (9) groups of fuel holes  214 , each group having a pattern of three (3) radial holes  216 ,  218 ,  220  (or a total of 27 holes). Each group of fuel holes  214  provide fluid communication from one of the axial bores  212  to combustor  20 . The nine axial bores  212  and the nine (9) groups of fuel holes  214  are generally equally spaced around the circumference of the fuel slinger  100  with the pattern of three holes  216 ,  218 ,  220  in each group of holes  214  being longitudinally and slightly angularly displaced from one another. Moreover, the holes  216 ,  218 ,  220  are oriented radially outwardly to insure a high velocity delivery thereto. 
     Referring now to FIGS.  5  and  6 A- 6 C, each pattern of three holes  216 ,  218 ,  220  are generally centered in the gap between the combustor liners  208 ,  210  and are drilled into each of the axial holes  212 , with one radial hole  216  being disposed on the centerline thereof and the remaining two semi-radial holes  218 ,  220  being disposed on alternate sides of the centerline. When the fuel enters the slinger cavity  200 , centrifugal force slings the fuel to the axial bore  212  where it flows forwardly as a narrow trough therein toward the compressor  18 . The volume of fuel determines which of the three holes  216 ,  218 ,  220  will dispense the fuel into the combustor  20 . More specifically, the radial hole  216  which is aligned with the centerline of axial bore  212  will receive the greatest amount of fuel of the three holes. The semi-radial holes  218 ,  220  will receive a lesser amount of fuel but provide additional fuel atomizing capacity over a slinger having only a radial hole. 
     As presently preferred, the location of the radial hole  216  is alternated between a forwardmost position f, an intermediate position i and a rearwardmost position r for each adjacent axial bores  212 . In this manner, the position of the maximum fuel flow from any given group of fuel holes  214  is alternated axially, thereby yielding an extremely uniform fuel spray having a generally cylindrical distribution about the fuel slinger  100  which provides for the excellent combustion characteristics within the combustor  20 . Moreover, the fuel flow in the combustor  20  is designed to have a characteristic swirl with the primary combustion located forwardly in combustor  20  and the remaining combustion as the gases progress therethrough. 
     With reference again to FIGS. 4 and 5, combustor  20  is defined by the front combustor liner  208  and the rear combustor liner  210  described, as well as a outer combustor liner  222 . More specifically, the outer combustor liner  222  is attached to a radially outward portion of front combustor liner  208 , and a plurality of combustor feed tubes  224  extend radially inward from outer combustor liner  222  and open onto nozzle hub  112 . The rear combustor liner  210  is attached to nozzle hub  112  which along with turbine nozzle assembly  114  defines a chamber  226  therebetween. The configuration of the combustor  20  is such to provide proper flame holding characteristics once the gas turbine engine  12  has been started to make the gas turbine engine  12  self-sustaining. The combusted gases are exhausted between exhaust cover  228  and outer combustor liner  222  and directed through turbine nozzle cover  230  so as to impinge on turbine  22 , thereby driving shaft assembly  24 . The combusted gases are then exhausted through the exhaust port  232  and exhaust header pipe  50  into diffuser duct  52 . 
     With reference now to FIG. 1, the turbo heater  10  is provided with an engine controller  300  which is operably coupled to a rotor speed pick up  302  and an EGT (exhaust gas temperature) sensor  304  for providing engine operating parameters thereto. An ignition system  306  having an ignition coil  308  and a spark plug  310  are also operably coupled to the engine controller  300  for initiating combustion. Controller  300  is also operably coupled to electronic fuel system  60  which includes a fuel supply  62  of the type previously described and an electric fuel pump  312  that pumps fuel to fuel passageway  180  formed in gas turbine engine  12 . Similarly, a starting system  314  having the starter motor  34  which is adequately sized to insure adequate power to start the gas turbine engine  12  during very cold weather is operably coupled to the controller  300 . 
     The starting system  314  preferably utilizes a soft start circuit  316  which limits the electrical current drawn by the starter motor  34  from the battery  318 . The soft start circuit  316  features a first or soft circuit  320 , a second or normal circuit  322  which are selectively actuated by controller  300 . The soft circuit  316  has a resistance element  324  interdisposed between the battery  318  and the starter motor  34 . As presently preferred resistance element is a coil of 0.093 inches in diameter stainless steel wire with straightened length of 10 inches within a resistance of about 0.0084 ohms as compared to the copper conductor of less than 0.0012 ohms. As illustrated in FIG. 1, the coil  324  of soft circuit  320  is enclosed in a housing  326  through which fuel passes so that the heat generated by the coil  324  may be partially recovered by preheating the fuel prior to combustion, thereby further enhancing cold weather starting of the gas turbine engine  12 . 
     The start sequence of turbo heater  10 , as controlled by engine controller  300 , proceeds as follows. The system is turned “ON” and a low current is drawn through soft circuit  320  by the starter motor  34  for initiating rotation of the shaft assembly  24  with a low torque. The lower torque, resulting from the high resistance of coil  324 , gently engages the clutch  36  and relatively slowly accelerates the gas turbine engine  12 . The spark plugs  310  are firing, but the electronic fuel pump  312  has not initiated fuel flow. As presently preferred, a two-spark plug arrangement is used to ensure starting particularly in cold weather conditions. After approximately two seconds when the gas turbine engine  12  is being driven at a steady low rpm speed, the controller  300  selectively actuates second circuit  322  and a high current is drawn therethrough by the starter motor  34  which quickly accelerates the speed of the shaft assembly  24 . The above-described start sequence is beneficial to the overall system in two ways. First, it improves the cold temperature starting capability of the engine by eliminating large current draws on the battery. Secondly, it eliminates high torque loading between the starter  34  and the turbine shaft assembly  24  by allowing the rotating components to slowly achieve operational speed. 
     The invention being thus described, it will be obvious that the same may be varied in many ways. Such variations are not to be regarded as a departure from the spirit and scope of the invention, and all such modifications as would be obvious to one skilled in the art are intended to be included within the scope of the following claims.