Patent Publication Number: US-2005126171-A1

Title: Uncoupled, thermal-compressor, gas-turbine engine

Description:
REFERENCES CITED  
      U.S. PATENT DOCUMENTS  
                               U.S. PATENT DOCUMENTS                                                        2,127,286    8/1938   Bush   62/6           2,175,376   10/1939   Bush et al.   62/3.1           3,991,586   11/1976   Acord   62/6           4,133,173    8/1977   Cataldo   60/39.63           4,984,432    1/1991   Corey   62/87           5,473,899   12/1995   Viteri   60/684           5,590,528    1/1997   Viteri   60/684           5,894,729   10/1997   Proeschel   60/508                      
 
     OTHER PUBLICATIONS  
     
         
          R. B. Aronson, “Stirling-Engine: Can Money Make It Work?” Machine Design, pp. 20, Apr. 24, 1980.  
          D. S. Beck and D. G. Wilson,  Gas - Turbine Regenerators , Chapman and Hall, pp. 52-55, 1996.  
          J. G. Bralla,  Design for Manufacturability Handbook , McGraw Hill, pp. 2.93 &amp; 4.194-4.195, 1999.  
          A. P. M. Glassford, “Adiabatic Cycle Analysis for the Valved Thermal Compressor,” American Institute of Aeronautics and Astronautics, J. Energy, Vol. 3, No. 5, September-October, 1979, p 306.  
          B. J. Hamrock,  Fundamentals of Fluid Film Lubrication , McGraw Hill, pp. 362-366, 1994.  
          B. A. Hands,  Cryogenic Engineering , Academic Press, pp. 449, 1986  
          J. B. Heywood,  Internal Combustion Engine Fundamentals , McGraw-Hill, pp. 575, 1988.  
          M. Lamm, “The Big Engine That Couldn&#39;t,” American Heritage of Invention and Technology, Volume 8/Number 3, pp. 40, Winter 1993.  
          D. G. Shepherd,  Principles of Turbomachinery , Macmillan, pp 320-323, 1954.  
       
    
     BACKGROUND  
      This invention is for a heat engine that: 
          1. uses a regenerator,     2. uses a displacer,     3. uses a turbine or other gas drive,     4. uses an energy source derived from continuous combustion or solar energy,     5. uses a quasi-constant-pressure process,     6. has very-low emissions,     7. has a very-long, continuous-use service life,     8. operates with an efficiency near the Carnot cycle,     9. operates efficiently when operating at a small fraction of rated output, and     10. belongs to the family of Ericsson cycle engines.        

      The regenerative gas turbine belongs to the family of Ericsson cycle engines. It is the preferred small gas turbine configuration. This engine has important limitations: 
          1. back work (work required to drive the compressor) puts a premium on turbine and compressor component efficiency, i.e., a small drop in component efficiency results in a much larger drop in engine efficiency;     2. small engines have low operating efficiencies;     3. turbine blade life is limited by high temperature metal fatigue and creep failure, significantly adding to operating cost and lowering service life;     4. the engine operates best as a constant output engine, i.e., operating efficiency can be poor at, say, 10% of rated output and thus not useful for many applications;     5. the regenerator requires a high pressure and high temperature gas seal; however, this problem can be overcome by accepting a lower efficiency and using a recuperator in place of the regenerator; and     6. it is costly relative to some engine types.        

      Gas turbines such as those used on aircraft have gained wide use because they have a low specific weight and are powerful, reliable and durable. However, to achieve good efficiency they require high combustion temperature that results in considerable emissions, and use turbine blades that require costly materials and typically fail due to creep failure or fatigue failure. In addition, they have very poor efficiency when operating at a small fraction of rated power.  
      Steam power plants operate on the Rankine cycle. There is essentially no back work for this system; however, the efficiency of the Rankine cycle is substantially lower than the Carnot cycle and steam plants consequently are limited to efficiency near 40%. Steam power plants operate efficiently only at a constant output and require a long time to power up.  
      The spark ignition (SI) engine has a moderate specific weight, cost and efficiency. It has gained universal use as a light-duty automotive engine. The SI engine requires an elaborate emission control system. The SI engine&#39;s high wear rates and service requirements preclude its use for applications requiring long continuous operation. Although better than the gas turbine or steam turbine, it has poor efficiency when operating at a small fraction of rated power.  
      Compression ignition (diesel) engines have become the premier heavy truck and industrial engine type. They have high emissions, significant wear and require regular maintenance. This engine is not stable at very low engine speeds.  
      Another regenerative gas cycle engine is the Stirling cycle engine that uses a constant-volume process as opposed to constant-pressure processes. Stirling cycle engine limitations include low volumetric efficiency and high-pressure, pushrod seal wear.  
      Another regenerative gas cycle engine that uses constant-pressure processes is the Ericsson engine. This engine has not gained significant market acceptance except for small engines and has high wear characteristics.  
      U.S. Pat. Nos. 2,127,286; 2,175,376; 3,991,586; 4,133,173; 4,984,432; 5,473,899; 5,590,528 and 5,894,729 have information on several Ericsson cycle engines or related information. However, each of these references suffers from the disadvantages of gas turbines and/or diesel and/or Stirling cycle engines.  
      Cogeneration units that generate electricity and use rejected heat to provide space heating and to heat water have gained limited acceptance for medium-size commercial and industrial facilities. They are essentially nonexistent for home use. Cogeneration reduces energy consumption and can offer considerable economic advantages to the user. Small cogeneration units such as for a single-family house or small business have not been successful because a heat engine with the necessary requirements has not been available. Such an engine ideally should: 
          1. operate continuously for at least ten years without the need for servicing;     2. have very low exhaust emissions over the ten-year interval;     3. have a good efficiency at both a very low and high output;     4. have a low manufacturing cost; and     5. ideally, be compatible with solar-based energy augmentation.        

      Small (5 kW) solar-thermal heat engine driven electric generators have failed to enter the market because the required engine has not been available. Such an engine would be low cost, have a ten-year maintenance-free service life and have a high efficiency.  
      A large (100 kW) space solar thermal power system has not been used because the required engine has not been available. Such an engine would have a fifteen-year, continuous, maintenance-free service life and have a high efficiency.  
      Due to cost, coal is the fuel of choice for electric power generation. Coal plants almost exclusively operate on the Rankine cycle and are typically limited to 40% energy conversion efficiencies. They operate as base power plants with a constant output.  
      Large natural gas electric power plants operate on either (1) the Rankine cycle and are typically limited to 40% efficiencies or (2) a gas turbine cycle and are typically limited to somewhat less than 40% efficiencies or (3) a combined cycle and are typically limited to less than 50% efficiencies.  
      Engines used with ground transportation systems operate at high combustion temperatures; consequently, they require complex and costly emission control systems, and operate at efficiencies that are much lower than are theoretically possible.  
      For the foregoing reasons, there is a need for a gas-cycle heat engine with the following capabilities: 
          1. scales well from 1 kW to 1 GW;     2. has a very long, continuous, maintenance-free service life;     3. has very low emissions without the need for a costly emission control device;     4. operates with an efficiency near the Carnot cycle so that relatively low combustion temperature can be used;     5. has responsive controls;     6. has a version that can be used as part of a home cogeneration unit;     7. has a version that can be used as part of a low-cost, solar-thermal, power system;     8. has a version that can be used as part of a space solar thermal power system;     9. has a version that operates with coal for large power plant use;     10. has a version that can use natural gas very efficiently for large power plants; and     11. has a version that is light and compact so it can operate with ground transportation vehicles, and in addition provide high torque at low speed or ideally a constant power output.        

     SUMMARY  
      The present invention, a heat engine, satisfies the needs stated in the background. The engine uses continuous combustion, a quasi-constant-pressure process, a thermal compressor (TC) that compresses gas and a drive that transforms the energy in the compressed gas into spinning shaft power. The TC compresses gas directly with heat. Internal pressure is typically high and varied as a means of varying torque output. The TC comprises a means of bringing heat into the engine, a cooler that removes heat, and a TC displacer drive. The engine brings heat in with a heater that uses external combustion or a heater that uses continuous internal combustion or continuous combustion directly in the hot chamber of the TC.  
      The engine has embodiments that remove (uncouple) heater and/or cooler interior volumes during gas compression. This improves volumetric efficiency, improves fuel use efficiency and enables the engine to scale well to large sizes. The engine cycle closely approximates the efficiency of the Carnot cycle and yields a high efficiency while limiting combustion temperature. Low combustion temperatures allow the engine to operate with very low emissions.  
      The continuous internal combustion version, which uses air as the working fluid, requires a clean fuel such as natural gas or clean distillates in order to avoid regenerator clogging. It requires a means of compressing air to the internal operating pressure and a means of extracting energy from the products of combustion before discharging them back into the atmosphere. This version can be light, small and powerful.  
      A version of the engine uses an innovative electromagnetic displacer drive. The displacer is spun and levitated with an integral small clearance seal and gas bearing. A linear electromagnetic motor induces oscillatory translational motion. By inducing a gas-dynamic bounce at the end of each stroke, engine speed increases. This version operates without the need for displacer wear surfaces and is a preferred version for applications, which require a long continuous service life. Another version uses a motor powered crank and pushrod as the displacer drive. A unique method obviates the high-pressure, pushrod-seal, wear problem. Another version of displacer drive uses a slender center rod to support the displacer and a linear electromagnetic drive. Very-low push rod or center rod seal wear occurs by means that equalize the pressure across this seal.  
      Encapsulation of high-temperature elements in a pressurized chamber and a partial vacuum in the interior of the displacer minimizes tensile stresses in ceramic components and allows the use of low cost ceramics for high temperature components.  
      The engine uses an innovative staged combustion heater with very low emissions. It does not require a catalytic converter. A compact version of the heater has a monolithic ceramic structure that implements staged combustion.  
      The engine innovation includes a constant power gas turbine drive. In essence, this drive can vary output torque so that power output remains constant over an operating speed range.  
      For a small, natural gas home cogeneration system a version of the engine uses an external combustion heater, electromagnetic displacer drive, decoupled cooler, and an impulse turbine that directly drives a constant-speed, 60 Hz, AC generator (costly frequency conversion power electronics are not required). The invention includes a version for use in outer space, as a solar thermal power system that can meet the stringent long continuous, maintenance-free requirements. For an automobile, a version of the engine uses several continuous internal combustion TCs and a constant power turbine drive. For a variable-output, coal power plant a version of the engine uses an external combustion heater, which is similar to a coal boiler. It also uses multiple crank-powered TCs; decoupled heater and cooler; constant speed turbo generator; and a pump system for varying system pressurization as a means for varying turbine torque. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
      In the specification the numbers 1 to 200 were reserved for figures and numbers larger than 200 are used for figure callouts.  
      The description given in the Summary and other features, aspects, and advantages of the present invention will become better understood with regard to the following description, appended claims, and accompanying drawings where:  
       FIG. 1  is a simplified engine schematic of a thermal compressor powering a turbo-generator. It is a version of the invention with an electromagnetic thermal compressor drive and an integral small clearance seal and gas bearing. This engine also includes valves and connectors required by a decoupled heater and decoupled cooler.  
       FIG. 2  is an engine schematic that shows the present invention as an amplifier that responsively transforms a low-power control signal into a high-powered output torque. The engine shown has a subsystem used to change system pressurization.  
       FIG. 3  is an engine schematic of the present invention and differs from  FIG. 2  by including the embodiments that decouple a heater and cooler.  
       FIGS. 4A, 4B ,  4 C, and  4 D are identical schematics except for valve position. These schematics are for the invention with the embodiments of an electromagnetic, linear displacer motor; bounce valves; heater and cooler decoupling valve; and external combustion heater.  
       FIG. 5A  shows ideal thermal-compressor, temperature-entropy curves of the present invention with the embodiments of a decoupled heater and cooler as shown in  FIGS. 4A, 4B ,  4 C,  4 D and  8 .  
       FIG. 5B  shows ideal thermal-compressor, temperature-entropy curves of the present invention with the embodiments of a decoupled cooler and continuous internal combustion system as shown in  FIGS. 7A and 7B .  
       FIG. 6  is an engine schematic of the invention with the embodiments of a displacer crank. This crank is in an oil-flooded crankcase that maintains an equal pressure between the crankcase and TC cold chamber. It has an oil recovery system that compensates for oil leakage at the pushrod and displacer.  
       FIG. 7A  is an engine schematic of the invention with the embodiments of a continuous internal combustion TC that uses a displacer crank, and maintains a constant pressure between the crankcase and TC cold chamber. This schematic includes a subsystem that compresses air from atmospheric pressure to the engine operating pressure. In addition, it uses an expander to extract compressed-gas energy from the products of combustion before discharging them into the atmosphere.  
       FIG. 7B  is an engine schematic similar to the one in  FIG. 7A  except that the displacer crank drive is replaced with a displacer using a center rod and linear electromagnetic drive.  
       FIG. 8  is an engine schematic of the invention with the embodiments of an continuous internal combustion heater (ICH) that uses a displacer crank, and maintains a constant pressure between the crankcase and TC cold chamber. This schematic includes a subsystem that compresses air from atmospheric pressure to the engine operating pressure and uses an expander to extract compressed-gas energy from the products of combustion before discharging into the atmosphere.  
       FIG. 9  shows a schematic of a home-based, natural gas cogeneration system augmented with solar power.  
       FIG. 10  shows a schematic of the invention with the embodiments of a solar energy receiver used as a heater that is part of the thermal compressor system.  
       FIG. 11  is a schematic that shows how concentric glass ceramic cylinders form an integral gas bearing and small clearance seal.  
       FIG. 12  shows an enlargement of the indicated area shown in  FIG. 11 . The figure shows one concentric cylinder attached to the displacer and the other attached to the closed container.  
       FIG. 13  shows an enlargement of the indicated area shown in  FIG. 11 . The figure shows how gaps insure pressure equalization on both sides of the two concentric cylinders forming an integral gas bearing and small clearance seal.  
       FIG. 14  shows an isometric of a herringbone gas bearing that can be incorporated into the small clearance seal.  
       FIG. 15  shows a descriptive drawing of the invention with the embodiments of a displacer and closed container. For this displacer spin and translation are electro-magnetically induced. It uses a permanent magnet attached to the displacer.  
       FIG. 16  shows a section parallel to the axis of the  FIG. 15  permanent magnet.  
       FIG. 17  shows a section perpendicular to the axis of the  FIG. 15  permanent magnet.  
       FIG. 18  shows an enlargement of a magnet region defined in  FIG. 16 . It shows magnet polarity of the outer portion of the magnet used to induce displacer translation.  
       FIG. 19  shows an enlargement of the inner portion of the magnet region defined in  FIG. 17 . It shows the spin motor magnet polarity.  
       FIG. 20  shows a wiring schematic of the two-phase push-pull drive coils that induce translation for the  FIG. 15  configuration.  
       FIG. 21A  shows a partial wiring schematic of the two-phase push-pull coils that spin the displacer for the  FIG. 15  configuration.  
       FIG. 21B  shows a partial schematic that completes the  FIG. 21A  schematic  
       FIG. 22A  shows a descriptive drawing of the invention with the embodiments of a displacer and closed container where the displacer drives use a magnet and sound-speaker-like linear coil drive and a center rod with balanced pressure.  
       FIG. 22B  shows a motorized bushing for use with a displacer center rod that eliminates both bushing and displacer labyrinth seal wear.  
       FIG. 23  shows a descriptive drawing of the invention with the embodiments of a displacer and closed container where spin and translation are electromagnetically induced, but does not use a permanent magnet. A linear brushless motor and a spin brushless motor in essence drive the displacer.  
       FIG. 24  shows a wiring schematic of the concept described in  FIG. 23 .  
       FIG. 25  is a descriptive elevation of elements of the invention including (1) an electromagnetic displacer drive and closed container assembly, (2) the valves and connectors required by a decoupled heater and decoupled cooler, (3) a set of valves used to induce a gas dynamic bounce at the end of a stroke, (4) an external pressurized ceramic cylinder which contains the displacer, (5) a high performance regenerator and (6) a pressurization chamber.  
       FIG. 26  shows an enlargement and section view of the aft coil region indicated in  FIG. 25 .  
       FIG. 27  shows an enlargement of the aft coil region indicated in  FIG. 26 .  
       FIG. 28  shows an enlargement of the mid displacer region indicated in  FIG. 26 .  
       FIG. 29  shows an enlargement of the aft coil and bounce valve region indicated in  FIG. 26 .  
       FIG. 30  shows an enlargement and section view of the hot chamber end indicated in  FIG. 25 .  
       FIG. 31A  shows a schematic of an active, vibration-mitigation system used with an electromagnetic-drive thermal compressor.  
       FIG. 31B  shows an integrated, gas-compressor system assembly comprising four electromagnetic-drive thermal compressors; a heater; a tilted-disk, vibration-mitigation subsystem; and a pressurization vessel.  
       FIG. 32  is a section view of the invention with the embodiments of a crank-driven displacer that obviates the problem of high-pressure seals by using a bellow that equalizes the pressure between the cold displacer region and the crankcase.  FIG. 6  is a schematic, which includes this embodiment.  
       FIG. 33  shows an enlargement of the region indicated in  FIG. 32 .  
       FIG. 34  shows a section view of a high-performance regenerator that has a large presented area achieved by means of a folded, heat-absorbing-media configuration. It has a short effective length and thus can rapidly absorb a large amount of heat while maintaining a low-pressure drop.  
       FIG. 35  shows the section of the high performance regenerator indicated in  FIG. 34 .  
       FIG. 36  shows the section for a high temperature ceramic heater valve set that uses a single electromagnetic actuator. It diverts gases through or around the heater, incorporates a dual position poppet valve and incorporates a check valve. This valve set appears in  FIGS. 25 and 32 .  
       FIG. 37  shows the section for a valve set that uses a single electromagnetic actuator used to divert gases to the cooler and incorporates two compressor check valves. This valve set appears in  FIGS. 25 and 32 .  
       FIG. 38  shows a schematic for a heater that uses staged combustion in order to limit peak combustion temperatures while consuming most of the oxygen.  
       FIG. 39  shows a section view of a staged combustion heater.  
       FIG. 40  shows a section view as indicated in  FIG. 39 .  
       FIG. 41  shows a section view as indicated in  FIG. 39 .  
       FIG. 42  shows an isometric exploded view of a high performance ceramic cloth to ceramic tube (air to working fluid) heat exchanger.  
       FIG. 43  shows an enlarged view of the ceramic front plate of the ceramic heat exchanger as indicated in  FIG. 42 .  
       FIG. 44  shows a section view defined in  FIG. 43  of the ceramic heat exchanger.  
       FIG. 45  shows a section view defined in  FIG. 44  of the ceramic heat exchanger.  
       FIG. 46  shows an enlarged view of the ceramic interior tube plate of the ceramic heat exchanger as indicated in  FIG. 42 .  
       FIG. 47  describes the concept of velocity compounding, which is an integral element of the constant power turbine, and shows how torque multiplies when the turbine-blade tip speed is much lower than the gas velocity.  
       FIG. 48  shows the trajectory of a particle of gas when the turbine-blade tip velocity is half the nozzle discharge velocity, i.e., turbine optimal speed.  
       FIG. 49  shows a curve of the ratio of turbine torque (T) divided by stall torque (T o ) versus the ratio of turbine speed (ω) divided by turbine optimal speed (ω o ) and a curve of turbine power (P) divided by maximum power (P o ) versus the ratio of turbine speed divided by turbine optimal speed.  
       FIG. 50  shows a front view of a constant power turbine.  
       FIG. 51  shows a section view as indicated in  FIG. 50 .  
       FIG. 52  shows a section view as indicated in  FIG. 51  and shows stator and turbine blade details.  
       FIG. 53  shows an unfolded and flattened view of the stator blades for the constant power turbine as defined in  FIG. 52 .  
       FIG. 54  shows an unfolded and flattened view of the of the turbine blades for the constant power turbine as defined in  FIG. 52 .  
       FIG. 55  shows the trajectory of a gas particle as it moves through the stator and turbine blades of a constant power turbine for the case when the turbine blade is not rotating. In this view, dashed lines indicate the portion of the trajectory in the stator blades and solid lines indicate the portion of the trajectory in the turbine blades.  
       FIG. 56  shows the trajectory of a gas particle as it moves through the stator and turbine blades of a constant power turbine for the case when the turbine blade is rotating. In this view, dashed lines indicate the portion of the trajectory in the stator blades and solid lines indicate the portion of the trajectory in the turbine blades.  
       FIG. 57  shows the trajectory of a gas particle as it moves through the stator and turbine blades of a constant power turbine for the case when the turbine blade is rotating at optimal speed.  
       FIG. 58  shows a schematic of a drive system for automotive applications that uses a thermal compressor as the power source.  
       FIG. 59  shows a wiring schematic for the drive system shown in  FIG. 58 .  
       FIG. 60  shows the geometric arrangement of a coupling coil with a view of the coil unfolded from the surface of a cylinder and placed on a flat plane. This coil is for the system shown in  FIGS. 58 and 59 .  
       FIG. 61  shows a section view of a wheel mounted automotive drive system.  
       FIG. 62  shows an enlargement as indicated in  FIG. 61 .  
       FIG. 63  shows a scaled down section view of the wheel mounted drive system showing the planetary gears as indicated in  FIG. 61 .  
       FIG. 64  gives equations and numerical values for ideal system efficiencies.  
       FIG. 65  gives the equation for gas-flow pressure drop in porous materials used to estimate the pressure drop in the regenerator.  
       FIG. 66  gives equations for the initial NO formation rate as a function of temperature and is used to show that the invention can operate efficiently without producing significant NO compounds.  
       FIG. 67  gives a defined scaling law and deduced relationships.  
       FIG. 68  gives system efficiency equations, related parameters and computed range of efficiencies.  
       FIG. 69  gives the specifications for a small home cogeneration unit that uses the embodiments of this invention to generate electricity and uses rejected heat to provide space heating and heat water.  
       FIG. 70  gives the specifications for a solar receiver TC system that uses the embodiments of this invention and can operate with the home cogeneration unit of  FIG. 69 .  
       FIG. 71  gives the specifications for a space solar thermal power system that can operate continuously and maintenance free for 17 years.  
       FIG. 72  gives the specifications for a cogeneration ICTC turbo generator, which uses the embodiments of this invention.  
       FIG. 73  gives the specifications for a base-load, central-power-plant, ICTC turbo generator, which uses the embodiments of this invention.  
       FIG. 74  gives the specifications for an automotive, ICTC engine, which uses the embodiments of this invention.  
       FIG. 75  gives the specifications for a heavy-truck, ICTC engine, which uses the embodiments of this invention.  
       FIG. 76  gives the specifications for a railroad, ICTC turbo generator, which uses the embodiments of this invention.  
       FIG. 77  gives the specifications for a TC turbo generator system that uses a helium-cooled, coal-fired heater, which uses the embodiments of this invention.  
       FIG. 78  describes the concept of velocity compounding in reverse, which enables the conversion of vehicle kinetic energy to compressed gas energy by taking a jet of gas increasing its velocity and compressing it by having it enter a converging-diverging nozzle.  
       FIG. 79  gives the equation for gas jet and relates pressure to the jet velocity for an inviscid  
       FIG. 80  is an engine schematic of the invention with the embodiments of a continuous internal combustion TC that uses a three stage reciprocal compressor, a three stage reciprocal expander and thermal compressor displacer, all on one shaft. An approximate isothermal compression and expansion results by removing heat in each stage of compression and transferred it to gas in each stage of expansion. 
    
    
     DESCRIPTION  
      1. Definitions  
      Presented below are the definitions of some of the specialized terms used in this specification: 
          A: regenerator area (see Section 3.14)     Adjustable turbine nozzle: a set of converging-diverging turbine nozzles and a means to turn them on and off, or a device that performs the equivalent function     Back iron: iron used in a magnetic circuit     Bounce valve: a valve used with electromagnetic drive TCs to stop gas flow out of either the TC hot chamber or TC cold chamber in order to cause gases to compress in the chamber and induce an adiabatic gas compression bounce of the displacer     C: controller     C P : specific heat of a gas at constant pressure     Ceramic heater valve assembly: a valve assembly constructed from refractory materials that perform three valve functions and effectively cause gases to either pass through the heater or bypass the heater     Closed container: a closed chamber containing a displacer which separates a cold chamber at one end with a hot chamber at the other     Cold chamber: the low temperature volume formed by the displacer and closed container     Constant-power turbine: an impulse turbine that uses velocity compounding to vary torque and achieve a constant power output as the speed changes (see Section 3.19)     Constant-pressure process: a heat engine cycle with pressure-volume curves which include two constant pressure lines     Cooler: a device that removes heat from the engine     Coupling coils: A pair of coils, one stationary and one rotating, used to energize a rotating circuit     CP: constant power turbine     D: displacer diameter     D&amp;C: displacer and closed container     Displacer: a piston-like structure that moves in a closed container and divides it into hot and cold chambers     Displacer drive coil: an electromagnetic coil mounted on the displacer and used to induce longitudinal oscillatory motion in the displacer     Displacer drive: an assembly of components that perform the function of inducing oscillatory motion in the displacer     Displacer exciter coil: a coil attached to the displacer that has AC current induced in it by the stationary exciter coil, the resulting current is rectified and used to power the displacer-mounted drive, spin and position coils     Displacer volume: the volume defined by the displacer cross-section area times the displacer stroke length     Drive coil: a coil used to induce translational motion in a displacer     ECH: external combustion heater     Eddy current clutch: a clutch that does not require contact and transfers torque using a coil and hysteresis loses     Electromagnetic drive: a displacer drive system, which uses electromagnetic forces to induce translational and rotational motion in the displacer     Engine specific weight: the engine weight divided by the rated power     Exciter: a device used to induce a current in a moving circuit     External combustion thermal compressor: a TC that receives heat through a heat exchanger     Gas drive: compressed gas motor and gear system     Gas-dynamic bounce: gas-dynamic spring effect caused by the closed container and displacer     GD: gas drive or gas motor     H: Hall sensor     HE: heat exchanger     Heat exchange module: a monolithic ceramic heat exchanger     Heater: a device that transfers heat into the system     Herringbone-groove journal bearing: a gas bearing with helical groves that improves gas-bearing stability     High performance regenerator: a regenerator with a small volume but a large throughput and configured as a folded heat absorbing media     Hot chamber: the high temperature volume formed by the displacer and closed container     HPT: high-pressure tank     ICH: continuous internal combustion heater     ICTC: continuous internal combustion TC, combustion in the hot chamber of the TC or along the path between the hot chamber and the regenerator     Continuous internal combustion thermal compressor: a TC that uses air as the working fluid and receives heat by means of combustion in the interior of the TC at a hot region     L: regenerator length (see Section 3.14)     LPT: low-pressure tank     M: motor     Magnetic window: a structure with a very low magnetic permeability used to pass magnetic fields from a high-pressure zone to a low-pressure zone     Monolithic ceramic heater: a complete ceramic heater fabricated by sintering a stack of ceramic plates     Motion response sensor: a device, such as an accelerometer, that senses motion on a structure     MTBF: mean time between failures     N: magnetic North Pole     O: oxygen sensor     Oil-gas separator: a liquid-gas separator used to separate oil from the working fluid     P: turbine power     P O : maximum turbine power     Position sensor slit: long thin optically transparent window, part of a sequential set     Position transducer: a coil used with a permanent-magnet displacer-drive-system to establish the position of the displacer     Pressure-equalizing bellow: a closed bellow located in the oil-flooded crankcase of a displacer drive with the interior of the bellow containing gas with a pipe to the cold chamber of the thermal compressor     Q: heat symbol     Quasi-constant-pressure process: a engine cycle process that can output a constant pressure and when pressure does change it is slow when compared to the cycle time     R: resultant velocity vector     Radiation-cone mirror: a conical structure with a mirrored surface in the interior of the displacer used to insulate the cold end from the hot end     Recuperator: a device that transfers heat across a surface     Regenerator: a device that receives and returns heat across the same surface     Regenerator filling factor: regenerator heat-absorbing media solid volume divided by its total volume     Regenerator length: the average distance between the hot and cold surfaces of the regenerator heat-absorbing media     Regenerator presented area: the area of the hot or cold surface where gas can enter the heat absorbing media     Regenerator total volume: the total volume of the regenerator heat-absorbing media     S: magnetic South Pole or entropy symbol     Small-clearance seal: a seal established by forming a small clearance so that leakage is at an acceptable level     Solar collection efficiency: the energy delivered to the turbine and hot water heat exchanger divided by the total solar energy striking the parabolic mirror     Solar receiver: a heater that uses solar energy     Specific weight: engine weight divided by rated power     Spin coil: a TC coil, either stationary or spinning, used to induce displacer spin     ST: storage tank     Stationary drive coil: fixed coils that interact with coils on a displacer and induce longitudinal oscillatory motion     Stationary exciter coil: fixed coils that interact with coils on the displacer and form a transformer that without contact transfers power to the moving displacer     Stationary spin coils: the stationary coils of a spin motor that spin the displacer     Stator blade mask: a structure that when positioned, effectively replaces turbine stator blades with a smooth surface and used to minimize the retarding force of a spinning turbine     System pressurization: the low-pressure side of a TC system     System support plate: the primary structure that support the thermal compressor and other system elements     T: temperature sensor or turbine torque symbol     T O : turbine stall torque or system ambient temperature     TC: thermal compressor     Thermal compressor: a device that compresses gas with the direct action of heat and without mechanical work     Thermal compressor (TC) displacement: the displacer-swept volume     Tr: transistor     Turbine, variable nozzle area: an impulse turbine that can vary the nozzle flow rate while maintaining a constant pressure drop through the nozzle     U: turbine tip velocity     UECTC: uncoupled, external-combustion thermal compressor     Uncoupled thermal compressor: a thermal compressor that uses valves to effectively remove the interior volume of the heater and/or cooler from the thermal compressor and thus dramatically improve volumetric efficiency     UTC: uncoupled thermal compressor     V: gas velocity     Velocity compounding: a method of using the fluid energy discharged by a turbine nozzle and consisting of multiple passes through turbine and stator blades so that each pass absorbs part of the discharged fluid energy and proportionately increases turbine torque (see Section 3.19     WT: water tank     ω: angular velocity     η: cycle efficiency        

      2. Overview  
      The subsection numbers in Section 2 correlate to invention claim numbers, i.e., 2.1 correlates to claim  1 , and so on. The material in these subsections gives an overview of the corresponding claim.  
      2.1 The first claim is for a heat engine with a thermal compressor (TC) that receives low-pressure gas (system pressurization) and delivers high-pressure gas to an output gas drive.  
      2.2 The second claim is for a version of the heat engine as recited in Section 2.1 and further comprising: 
          (1) continuous internal combustion in the hot chamber of the thermal compressor, and further comprising,     (2) a thermal compressor with combustion occurring in the hot chamber or at some point in the gas dynamic circuit in or between the hot chamber and the regenerator,     (3) means of pumping fuel into the hot chamber of the thermal compressor, and     (4) a pushrod-driven integral compressor, expander and displacer which respectively pressurizes air up to the system operating pressure, extracts energy from the products of combustion before discharging them into the atmosphere and provide system pressurization.        

      2.3 This claim is for a version of the heat engine as recited in Section 2.2 with the regenerator integrated into the close container so that none of the closed container structure is subjected to both high temperatures and high tensile stresses, and also comprising a regenerator that conforms to the hot chamber of the thermal compressor.  
      2.4 This claim is for a version of the heat engine as recited in Section 2.3 with elements that improve volumetric efficiency by effectively removing the cooler interior volume during compression, and further comprising a thermal compressor valve set configured so that: 
          (1) during the compression stroke gas follows a path from the cold chamber, then through the regenerator and then into the hot chamber,     (2) during the intake stroke gas follows a path from the hot chamber and regenerator and then discharges from the thermal compressors to an external cooler, and     (3) simultaneously, during the intake stroke, fresh gas directly enters the cold chamber.        

      2.5 This claim is for a version of the heat engine as recited in Section 2.4 with elements that significantly reduce noise, friction and wear and further comprising: 
          (1) a pushrod that interfaces with the crank drive by means of an integral thrust bearing and spin motor, and in so doing the integral pushrod, compressor, expander and displacer assembly can spin continuously;     (2) a noise mitigator to transform the pulsating intake and exhaust gases into a near continuous intake and exhaust flow processes by means of a cylinder divided by a spring loaded piston wherein one side is connected to the intake and compressor and the other side is connected to the exhaust and expander;     (3) a heat exchanger that transfers heat of compression in the compressor to expanding gas in the expander;     (4) an integral lubrication and heat exchanger system that pressurizes oil, sprays it in compressor and expander chambers, and separates it from air and products of combustion; and     (5) an integral cooler and exhaust gas scrubber comprising a gas to atmosphere heat exchanger, a chamber with means to form a dense water aerosol and a liquid-gas separator wherein the gas entering the cooler first through the heat exchanger, then the water aerosol and finally the liquid-gas separator.        

      2.6 This claim is for a version of the heat engine as recited in Section 2.1 with a TC displacer and a closed container that has no contact, and therefore no wear surfaces between the displacer and the closed container. The engine includes: 
          1. a gas bearing that supports the displacer relative to the closed container,     2. a small clearance seal consisting of two concentric cylinders with one attached to the displacer and one attached to the closed container,     3. a spin motor that induces axial rotation and an electromagnetic linear displacer drive that induces reciprocating motion of the displacer, and     4. a means of determining the position of the displacer relative to the closed container.        

      This innovation is ideal for systems that operate continuously and/or do not require maintenance beyond an annual air filter change.  
      2.7 This claim is for a version of the heat engine as recited in Section 2.6 and further comprises at least one TC valve configured so that confined gas cause a gas dynamic displacer bounce near the end of the displacer stroke at both ends of the closed container. This innovation conserves displacer kinetic energy and increases system performance by permitting a higher displacer speed.  
      2.8 This claim is for a version of the heat engine as recited in Section 2.7 and further comprising a set of nested cylinders attached to the displacer and interlaced with a set of nested cylinders that attach to the cold end of the closed container. These nested cylinders form: 
          1. a spin motor that magnetically induces a displacer torque,     2. a linear motor that magnetically induces a longitudinal force in the displacer,     3. a magnetic transducer system from which the position of the displacer can be determined, and     4. an integral air bearing and small clearance seal.        

      This innovation also allows the displacer to float and not contact the closed container walls and eliminates TC surface wear.  
      2.9 This claim is for a version of the heat engine as recited in Section 2.7 and further comprising an optical position sensor in place of a magnetic position sensor.  
      2.10 This claim is for a version of the heat engine as recited in Section 2.7 and further comprising a permanent magnet attached to the displacer in place of an electromagnet.  
      2.11 This is a claim for a version of the heat engine as recited in Section 2.7 and further comprising an optical position sensor in place of a magnetic position sensor.  
      2.12 This claim is for a version of the heat engine as recited in Section 2.1 with a TC displacer drive that uses a permanent magnet and sound-speaker like coil. This is the simplest version of the engine. A lubricated center rod supports the displacer and a dry displacer seal is used.  
      2.13 This claim is for a version of the heat engine as recited in Section 2.12, 2.31, 2.32, 2.33, 2.34 or 2.35 and further comprising a system for varying system pressurization as a means of varying gas-drive torque.  
      2.14 This claim is for a version of the heat engine as recited in Section 2.13 and further comprising a decoupled cooler.  
      2.15 This claim is for a version of the heat engine as recited in Section 2.13 and further comprising a decoupled heater.  
      2.16 This claim is for a version of the heat engine as recited in Section 2.15 and further comprising a decoupled cooler. Thus, both the heater and cooler are decoupled.  
      2.17 This claim is for a version of the heat engine as recited in Section 2.16 and further comprising: 
          1. an engine structure of ceramic manufacture,     2. an engine structure resistant to thermal fatigue and thermal shock failures, and     3. a pressure chamber that pressurizes high temperature ceramic components so that tensile stresses in ceramic components are small.        

      2.18 This claim is for a version of the heat engine as recited in Section 2.17 and further comprising integration into a cogeneration system. This system includes a heater, turbo generator and cooler incorporated into a hot water tank. This cogeneration system is ideal for both small and large systems.  
      2.19 This claim is for a version of the heat engine as recited in Section 2.17 and further comprising a coal-fired heater. This heater is similar to a steam power plant boiler except that gas circulates through it instead of water and steam. Unlike a steam power plant, the output of this plant can efficiently vary from rated output to a small fraction of rated output.  
      2.20 This claim is for a version of the heat engine as recited in Section 2.17 and further comprising a solar receiver used as a heater. The output of this system is compressed gas that drives a turbo generator, and warm gas used to heat water and/or supply space heating.  
      2.21 This claim is for a version of the heat engine as recited in Section 2.17 and further comprising a reaction turbine as the gas drive. The low temperature gases allow the use of durable, low-cost, complex, multistage reaction turbines for high torque, low speed applications. The reaction turbine is very efficient and low speeds allow minimization of overall system costs.  
      2.22 This claim is for a version of the heat engine as recited in Section 2.17 and further comprising elements for a space solar thermal power system that can operate continuously and maintenance free for a very long period.  
      2.23 This claim is for a version of the heat engine as recited in Section 2.1 with a natural gas power system augmented with solar energy, used in small and medium applications, and comprises a: 
          1. solar receiver and TC,     2. sun tracking parabolic mirror,     3. natural gas heater and TC,     4. hot water tank and heat exchanger, and     5. turbo generator.        

      With a clean mirror, at least 70% of solar energy is collected and then overall efficiency is 56%. Of the collected energy, 25% is transferred to hot water and 31% is converted to electricity.  
      2.24 This claim is for a version of the heat engine as recited in Section 2.1 with a TC displacer seal and integral gas bearing which permits a very durable engine by eliminating displacer wear surfaces. Elements include two concentric cylinders having a small clearance, configured from a material with a small coefficient of thermal expansion and high service temperature, and attached so that pressure equalizes on both sides of each cylinder.  
      2.25 This claim is for a version of the heat engine as recited in Section 2.1 with a motorized, center-rod bushing, a device that induces a displacer centering force, as a means of minimizing bushing and displacer seal wear. This device uses (1) a lubricated slender center rod which supports a displacer, (2) an inner bushing that is motor driven and provides a centering force against the center rod, (3) an outer bushing that interfaces with the inner bushing, (4) a support structure for the outer bushing, (5) a motor that drives the inner busing, and (6) a means to enable the gas pressure at the base of the center rod to equalize with the pressure of the closed container cold chamber. A TC version that uses a labyrinth seal and motorized center rod overcomes the TC wear problem. This motorized bushing maintains an oil film between the bushing and center rod and maintains a centering force on the center rod. Therefore, bushing wear and labyrinth seal wear is not significant after an initial wear-in period.  
      2.26 This claim is for a version of the heat engine as recited in Section 2.1 with an active vibration-mitigation system used with an electromagnetic-drive thermal compressor comprising: 
          1. system support plate,     2. a soft shock isolation spring,     3. an active damper drive coil and structure, and     4. an active damper armature.        

      2.27 This claim is for a version of the heat engine as recited in Section 2.1 with a gas compressor for gas cycle engines integrated with a vibration-mitigation sub-system, electromagnetic-drive TCs, a heater and a pressurization vessel.  
      2.28 This claim is for a version of the heat engine as recited in Section 2.1 with a high-throughput regenerator with a heat recovery media configured as a folded plate. Advantages of this regenerator are that it operates with a low-pressure drop, efficiently recovers heat and has a small interior volume.  
      2.29 This claim is for a regenerator as recited in Section 2.28 and further comprising means to both recover heat from the previous cycle and receive heat transferred from a heater.  
      2.30 This claim is for a regenerator as recited in Section 2.29 and further comprising a oxidation catalytic coating on the heat recovery material so that the regenerator serves the additional function as an oxidation catalytic converter.  
      2.31 This claim is for a version of the heat engine as recited in Section 2.1 with a heater with a sequence of combustion chambers and heat exchangers, and which can be used with a gas-cycle heat engine. Combustion occurs in stages with heat extracted and fuel added after every stage. The formation of NOx compounds is minimized by limiting fuel flow rates so that peak combustion temperatures are below some desired value.  
      2.32 This claim is for a gas-cycle heat engine heater as recited in Section 2.31, with a ceramic heat exchanger configured as a monolithic structure and formed by sintering a stack of alternating plates consisting of ceramic cloth and ceramic tubing.  
      2.33 This claim is for a gas-cycle, heat-engine heater as recited in Section 2.32, with a pressurized containment structure and so configured to minimize tensile stresses on components.  
      2.34 This claim is for a version of the heat engine as recited in Section 2.1 with a monolithic ceramic heater formed by sintering and comprised of a front and back plate plus a three-plate repeated sequence characterized as a cloth layer, a working fluid pipe plate layer and a fuel pipe plate layer.  
      2.35 This claim is for a version of the heat engine as recited in Section 2.1 with a TC structure with means to protect high temperature components against thermal fatigue and thermal shock failures. It comprises a pressure chamber that contains a high-pressure gas, encapsulates the TC structural assembly, and contains insulation between the pressure chamber and structural elements. Thus, high temperature elements predominantly experience compressive stresses. This allows the use of low cost ceramics for high temperature components.  
      2.36 This claim is for a TC structure as recited in Section 2.35 and further comprising a structure of ceramic manufacture for high temperature elements.  
      2.37 This claim is for a version of the heat engine as recited in Section 2.1 with a gas-dynamic drive which maintains a near constant power output over a specified speed range. The drive comprises a turbine, a stator and a gas discharge nozzle configured to permit velocity compounding to occur. It can be designed with a stall torque more than 10 times the torque at rated power and thus act as both a turbine and transmission. It has a very high power-to-weight ratio.  
      2.38 This claim is for a gas-dynamic drive as recited in Section 2.37, and with more than one stator-blade set configured to nullify gas-dynamic forces, not inducing turbine drive torque.  
      2.39 This claim is for a gas-dynamic drive as recited in Section 2.38, and with a forward and reverse-retard capability, and further comprising an additional turbine, stator and nozzle that can induce a reverse torque.  
      2.40 This claim is for a gas-dynamic drive as recited in Section 2.39, and with a mask that covers the stator blades associated with the reverse-retard turbine when operating in the forward drive mode. This innovation minimizes unwanted retarding torques without the need for a clutch that disengages the reverse turbine.  
      2.41 This claim is for a gas-dynamic drive as recited in Section 2.40, and with a magnetic force means that transfers the torque of the turbine from a high-pressure chamber to a low-pressure chamber.  
      2.42 This claim is for a gas-dynamic drive as recited in Section 2.41, and with an electric clutch that transfers the torque of the turbine contained in a high-pressure chamber, to a low-pressure chamber.  
      2.43 This claim is for a gas-dynamic drive as recited in Section 2.42, and with: 
          1. a toroidal-shell pressure chamber containing the turbine,     2. the toroidal shell in contact with rotating elements when the toroid is not pressurized, and     3. a clearance gap, between the toroidal shell and interior rotating elements, when the toroid is pressurized.        

      This concept allows fabrication of the toroid as a fiber composite structure using the inner elements, some of which rotate, as the supporting structure used to form the toroid.  
      2.44 This claim is for a gas-dynamic drive as recited in Section 2.43, and with means to insure a smooth turbine torque output. The means comprising a turbine nozzle which can vary the flow rate, a pressure gauge which measures the pressure upstream of the nozzle and a nozzle controller which modulates the nozzle so that TC induced pressure fluctuations do not induce corresponding turbine-torque fluctuations. This innovation compensates for pressure fluctuations inherent in the TC output.  
      2.45 This claim is for a gas-dynamic drive as recited in Section 2.44 and configured for wheel mounting and further comprising a planetary reduction gear and a disk brake system. The power-to-weight ratio of this innovation can exceed 15 kW/kg. The light weight of this wheel-mounted drive insures good motor vehicle suspension-related drivability.  
      2.46. A gas dynamic drive according to claims  37 ,  38 ,  39 ,  40 ,  41 ,  42 ,  43 ,  44  or  45 , with elements that can convert mechanical energy into compressed gas energy and comprising a turbine operating in reverse, a stator, a gas discharge nozzle, a gas intake nozzle, means that enable velocity compounding to occur and means to store the compressed gas.  
      3. Detailed Description of the Elements  
      The invention includes innovations at the system level and at the component level. These sections first give detailed describes of the invention at the system level and then give detailed descriptions of components.  
      3.1 Thermal-Compressor, Gas-Turbine Engine  
       FIG. 2  shows a thermal-compressor based heat engine powering a generator. It is comprised of a cooler  302  that transfers heat from the working fluid to another media  300 ; a regenerator  254 ; a heater  304  that receives fuel  314  and air  316 , and discharges the products of combustion  306 ; a closed container  264 ; displacer  262 ; displacer seal  260 ; a pushrod  257 ; a pushrod seal  258 ; a connecting rod  320 ; a crank  319 ; a drive motor  312 ; a controller  310 ; a low pressure tank  318 ; an intake check valve  268 ; a discharge check valve  272 , a high-pressure tank  322 ; a gas drive  329 ; a generator  324 ; a depressurization pump  298 , motor  296 , check valve  290  and storage tank  294  all used to reduce system pressurization; and variable-flow-rate, pressurization valve  292  which is used to increase system pressurization. Start with the displacer at the far right so that the hot chamber is essentially empty and the cold chamber is full. As the displacer moves to the left, gas is driven through the cooler, regenerator and heater causing the average temperature of the gas to go up and consequently the pressure to go up. At some operating pressure, some gas starts to discharge through the exhaust check valve. Gas continues to discharge until the displacer reaches the far left position. As the displacer moves to the right, the average temperature starts dropping with a corresponding drop in pressure. When the pressure drops below the intake pressure new low-pressure gas enters the cold chamber replacing the gas discharged at high-pressure. Varying system pressurization varies turbine torque. The engine controller  310  receives a command  308  and the engine induces a generator output response  326 . The thermal compressor pressurizes the high-pressure tank that in turn drives the turbine. The displacer drive motor is responsive to the controller and, therefore, the heat engine is responsive. Increasing the size of the high-pressure tank decreases the pressure fluctuations at the turbine; however, in this case system responsiveness also decreases.  
      Closed cycle systems use ambient pressures as high as 200 atmospheres. Using engine system pressurization as a design variable as pressure goes up, specific volume improves at first and then peaks. This is due to a need for an ever-larger heater, cooler and regenerator that reduces volumetric efficiency. Poor volumetric efficiency limits the utility of this engine to small engine sizes. In addition, pressures of 200 atmospheres cause very difficult problems with the pushrod seal wear, making them impractical for many applications.  
      3.2 Uncoupled, External-Combustion, TC, Gas-Turbine Engine  
      The engine schematic in  FIG. 3  resolves the problem of poor volumetric efficiency associated with the schematic of  FIG. 2 . The addition of the three-valve set  340 ,  341  and  342  accomplished this by effectively removing the heater and cooler interior volumes. In this schematic, the valves are in the position for the compression stroke that corresponds to the displacer moving to the left. Both the cooler and the heater are not in the gas dynamic circuit. When the displacer starts moving to the right, valves  340  and  341  change position. In this changed position, gas exiting the displacer hot chamber  276  sequentially goes through the heater valve  341 , heater  304 , heater check valve  342 , regenerator  254 , cooler valve  340 , cooler  302 , and low-pressure tank  318 . This process in essence transfers heat from the heater to the regenerator for use in the next compression stroke.  
      3.3 Uncoupled, ECTC with Electromagnetic Displacer Drive Gas-Turbine Engine  
      The engine defined by the  FIG. 1  schematic obviates the high-pressure, pushrod-seal wear problem by eliminating the need for the seal. The use of an electromagnetic displacer drive does not require a pushrod seal. This engine obviates the adverse effect on volumetric efficiency due to heater and cooler interior volumes with a valve set that effectively removes these volumes during the compression stroke. The valve operation is the same as that for  FIG. 3  described above. The heat engine in  FIG. 1  comprises a turbine  328 ; a generator  324 ; a high-pressure tank  322 ; a low-pressure tank  318 ; a cooler  302  that transfers heat to a cooling media  300 ; a discharge check valve  272 ; an intake check valve  268 ; a heater  304  which receives fuel  314  and air  316 , and exhausts products of combustion  306 ; a three-port, two-position solenoid valve  341  and a check valve  342  that operate to divert gases from the hot chamber  276  through the heater  304  and then through the regenerator  254  during the intake stroke and then operate to cause gases to flow directly from the regenerator into the hot chamber during the compression stroke; a three-port, two-position solenoid valve  340  that operates to divert gases directly to the cooler during the intake stroke and bypass the cooler during the compression stroke; and an integrated closed-container  346  and displacer  262 . The displacer  262  spins about its axis, levitated by a gas bearing and given longitudinal motion by electromagnetic drive coils. A detailed description of the closed container, displacer and displacer drive is in Sections 3.10 and 3.11. Here we only note that two types of electromagnetic drive systems are used, e.g., a low-performance, low-cost, permanent-magnet drive and high-performance, electromagnetic drive. There are no dynamic contacting surfaces for either one of these systems.  
      The engine schematic in  FIG. 4A  is an enhancement of the  FIG. 1  schematic by enabling gas dynamic bounce. It has more TC valves configured so that confined gas cause a gas dynamic displacer bounce near the end of the displacer stroke at both ends of the closed container. It is an uncoupled, external-combustion, thermal-compressor gas drive engine (UECTC) that powers a generator.  FIGS. 4A, 4B ,  4 C and  4 D are schematics for the same heat engine and differ by the position of the four-solenoid valves. These four drawings define the four distinct valve arrangements used in the engine cycle. The four arrangements are (1) compression,  FIG. 4A ; (2) cold end gas dynamic bounce,  FIG. 4B ; (3) gas intake,  FIG. 4C ; and (4) hot end gas dynamic bounce,  FIG. 4D .  
      The heat engine in  FIG. 4A  comprises a turbine  328 ; a variable flow turbine nozzle  327 ; a generator  324 ; a controller  310 ; a high-pressure tank  322 ; a low-pressure tank  318 ; a pressure transducer  323 ; a cooler  302  that transfers heat to a cooling media  300 ; a discharge check valve  272 ; an intake check valve  268 ; a heater  304  which receives fuel  314  and air  316 , and exhausts products of combustion  306 ; a two-port, two-position solenoid valve  344  and a check valve  342  that operate to divert gases from the hot chamber  276  through the heater  304  and then through the regenerator  254  during the intake stroke, and then operates to uncouple the heater during the compression stroke; a two-port, two-position solenoid valve  348  which is open during the compression stroke and closed during the intake stroke; a three-port, two-position solenoid valve  340  that operates to divert gases directly to the cooler during the intake stroke and bypass the cooler during the compression stroke; a two-port, two-position solenoid valve  347  that closes when the displacer comes close to the left end in order to trap gases and induce a gas dynamic bounce on the displacer; an integrated closed container  346  and displacer  262 ; and a pressure transducer  345  that measures the pressure in the cold chamber. The displacer  262  spins about its axis, levitated by a gas bearing and given longitudinal motion by electromagnetic drive coils. A detailed description of the closed container, displacer and displacer drive is in Sections 3.10 and 3.11. Here we only note that two types of electromagnetic drive systems are used, e.g., a low-performance, low-cost, permanent-magnet drive and high-performance, electromagnetic drive. There are no dynamic contacting surfaces for either one of these systems.  
       FIG. 4B , which characterizes the valve arrangement for cold end gas dynamic bounce, is identical to  FIG. 4A  except that the two valves  347  and  348  are in the switched positions.  FIG. 4C , which characterizes the valve arrangement for gas intake, is identical to  FIG. 4A  except the two valves  340  and  344  are in a switched position.  FIG. 4D , which characterizes the valve arrangement for hot-end, gas-dynamic bounce, is identical to  FIG. 4A  except the two valves  340  and  348  are in a switched position.  
      The UECTC cycle description starts with the displacer near the far-right end and moving to the left with the valves as shown in  FIG. 4A . In this position, a path exists from the cold chamber  270 , through the valve  347 , through the valve  340 , through the regenerator  254 , then through the valve  348 , and into the hot chamber; and a path exists from the cold  270  chamber, through the valve  347 , through the discharge check valve  272  and into the high-pressure tank  322 . In addition, in this position the pressure in the thermal compressor is at the ambient value. As the displacer moves to the left, gas starts moving from the cold chamber to the hot chamber, the average temperature starts rising in the volume including the hot and cold chambers and the path between them, and then a pressure rise follows. As soon as the pressure exceeds the pressure in the high-pressure tank  322 , gases start flowing to it. When the displacer reaches a point near the far left position, the cold-chamber, bounce solenoid valve  347  closes. This causes the pressure in the cold chamber to start rising independent of the pressure in the hot chamber. This causes the displacer to decelerate and then reverse direction so that it is now moving to the right, which causes the cold-chamber pressure to start dropping. At this point, the cold-chamber pressure is higher than the pressure in the high-pressure tank  322  and the bounce solenoid valve  347  activates to open, however, the high bounce pressure inhibits it. As the cold-chamber pressure approaches that of the high-pressure tank, the bounce solenoid valve opens causing the pathway between the cold and hot chambers to open. As the displacer moves further to the right the pressure in both the hot and cold chambers drops until the pressure falls below the pressure of the low-pressure tank, then the intake check valve opens allowing gas to enter. In addition, as the intake valve  268  opens the heater solenoid valve  344  activates to open, the regenerator solenoid valve  348  activates to close and the cooler solenoid valve  340  activates to open a path to the cooler. This causes gases discharged from the hot chamber to go through the heater, then the regenerator, then through the cooler and finally into the low-pressure tank. When the displacer reaches a point near the right end, the heater solenoid valve  348  actuates causing the pressure in the hot chamber to raise independent of the cold chamber; that in turn causes the displacer to decelerate and reverse directions. As the pressure starts dropping, the regenerator solenoid valve activates to open; however, the high bounce pressure inhibits opening until the hot chamber pressure approaches the low-pressure tank pressure. After the regenerator valve opens, the cycle is complete.  
      Conceptually, when the displacer is at the far right, the cold chamber consists of two parts. The gases in one part undergo an adiabatic compression and discharge into the high-pressure tank; and the gases in the other part first move through the regenerator and into the hot chamber, then move through the heater and cooler, and discharge into the low-pressure tank.  FIG. 5A  gives conceptual, temperature versus specific-entropy curves for the gas that starts in the direction of the regenerator and implies an ideal regenerator. In  FIG. 5A : 
          1. T designates temperature;     2. S designates entropy;     3. C P  is the specific heat at constant pressure;     4. point F corresponds to the gases in the cold chamber at ambient pressure;     5. point A corresponds to the gases in the cold chamber after they are compressed;     6. point B corresponds to the gases after they move through the regenerator and into the hot chamber;     7. point C corresponds to the gases in the hot chamber after the pressure drops;     8. point D corresponds to the gases after they move through the heater;     9. point E corresponds to the gases after they again move through the regenerator; and     10. point F corresponds to the gases after they move through the cooler.        

      The symbol Q corresponds to heat transferred across a surface or mechanical energy extracted from the gas. In  FIG. 5A  the area: 
          1. Q 1  corresponds to the mechanical work extracted from the gases;     2. Q 2  corresponds to the heat extracted from the gases by the cooler;     3. Q 3  corresponds to the heat added to the gases by the heater; and     4. Q 3 +Q 4  corresponds to the heat extracted from the gases by the regenerator in one cycle and delivered to the gases in the next cycle.        

       FIG. 64  presents cycle efficiency equations for cycles defined in  FIGS. 5A and 5B , and the Carnot and Ericsson cycles and respectively designated by η EC , η IC , η CARNOT  and η ERICSSON . Note that heat from the heater is first stored in the regenerator and then used.  
      This cycle approximates the Ericsson cycle. Another cycle, which approximates the Ericsson cycle, is the regenerative, gas-turbine cycle. In the regenerative, gas-turbine cycle, moving from point B to C corresponds to a perfect adiabatic expansion in the inlet nozzle, Q 3  is the heat of combustion and Q 4  is the heat recovered by the regenerator. Thus, the UECTC has thermodynamic similarities to the regenerative gas turbine, but differs by not requiring back work, i.e., extracting turbine energy to compress gases. The energy losses due to back work are more pronounced for small (100 kW) regenerative gas turbines and typically limit efficiencies to values below 30%, whereas uncoupled thermal compressors can typically achieve values above 50%.  
      3.4 Balanced-Pressure, Crank-Drive, TC, Gas-Turbine Engine  
       FIG. 6  shows another schematic of an external-combustion, uncoupled, thermal-compressor gas turbine driving a generator. This system principally differs from the  FIG. 4A  system by its use of a crank and pushrod driven displacer. This heat engine obviates the high-pressure pushrod seal problem by means of a unique pressure-balancing method. Service life is more than an order of magnitude larger than a reciprocating, spark-ignition engine. Only a very small differential pressure acts on the displacer seal.  
      The heat engine in  FIG. 6  comprises a high-pressure tank  322 ; a variable flow turbine nozzle  327 ; a turbine  328 ; a generator  324 ; a controller  310 ; an oil-gas separator  360 ; a discharge check valve  272 ; an intake check valve  268 ; a low-pressure tank  318 ; a cooler  302 ; a cooling media  300 ; a heater  304  which receives fuel  314 , air  316  and exhausts products of combustion  306 ; an oil sump  362  which is used to replenish crankcase oil  368 ; a three-port, two-position solenoid valve  340  that operates to directly divert gases to the cooler during the intake stroke and bypass the cooler during the compression stroke; a regenerator  254 ; a three-port, two-position solenoid valve  341  and a check valve  342  that operate to divert gases from the hot chamber  276  through the heater  304  and then through the regenerator  254  during the intake stroke and then operate to cause gases to flow directly from the regenerator into the hot chamber during the compression stroke; a crankcase  366 ; a drive motor  312 ; oil which floods the crankcase  368 ; a crank arm  319 ; a connecting rod  320 ; a pressure equalizing bellow  370  and low oil sensor which activates the solenoid valve  364  that restores lost crankcase oil; displacer pushrod  357 ; pushrod seal  258 ; a closed container  264 ; a displacer  262 ; a displacer seal  260 ; a hot chamber  276 ; and a cold chamber  270 .  
      An important innovation of this system is the use of a crankcase  366  that is completely flooded with oil and the use of a bellow  370  that insures that the pressure in the crankcase and cold chamber  270  will be essentially equal during operation. Although there is a pressure drop across the regenerator at low speeds, it is very small and for a well-designed system at rated output, the maximum pressure drop can be below 30 psi. Consequently, the loads on the displacer seal, pushrod and crank bearings are all very low. With modest size crank bearings, an oil film thickness large enough to eliminate bearing wear is possible. Seal wear is generally proportional to the load. With low-pressure seal loads, an oil film is maintaining and very low displacer-seal wear rate and pushrod-seal wear rate occurs.  
      3.5 Continuous Internal-Combustion, Uncoupled, TC, Gas-Turbine Engine  
       FIGS. 7A and 7B  show schematics of continuous-internal-combustion, uncoupled, thermal-compressor, gas-turbine engines. These systems differs from the system in  FIG. 4A  by using air as the working fluid in place of hydrogen, helium or nitrogen, using continuous internal combustion in place of a heater, using a compressor to bring high-pressure air into the system and using an expander to extract energy from high-pressure exhaust gases before discharging them into the atmosphere. Note that air, once pressurized, circulates numerous times before it depressurizes and discharges back into the atmosphere. For typical operating conditions, combustion for one cycle will consume an amount of oxygen equal to a small proportion of that in atmospheric air.  
      The heat engine schematic in  FIG. 7A  uses a crank and cam activated cooler valve. The heat engine schematic comprises a controller  310 ; a cooler  302  discharging heat into a cooling media  300 ; a liquid-gas separator  391  that separates out water from the products of combustion and discharges the water  393 ; an expander  380 ; a compressor  384 ; a heat exchanger  383  that cools gases between compression stages and heats gases between expansion stages; an intake choke  432 ; an air intake filter  385 ; an air intake  390 ; an exhaust  392 ; an exhaust choke  430 ; a high-pressure tank  322 ; a gas drive  329  that outputs power to an output shaft  428 ; an oil-gas separator  360 ; a discharge check valve  272 ; a low-pressure tank  318 ; an intake check valve  268 ; a three-port, two-position, cam-actuated cooler valve  394 ; a regenerator  254 ; a fuel tank  396 ; a fuel pump motor  398 ; a fuel pump  400 ; a fuel filter  402 ; an adjustable, fuel flow control  404  which is regulated by the controller to maintain a constant temperature in the hot chamber  276 ; a pressurized, gravity, oil sump tank  362 ; an oil heater  363  that drives off water; an oxygen sensor  361 ; a two-port, two-position oil valve  364  which responds to commands to add oil to the crankcase; crankcase  366 ; a crankshaft cam  410  that drives the cooler valve  394 ; crank drive motor  312 ; crankcase oil  368 ; crank arm  319 ; connecting rod  320 ; a pressure-equalizing bellow and low-oil sensor  370  which activates the solenoid valve  364  that restores lost crank-case oil; displacer pushrod  357 ; pushrod seal  258 ; a closed container  264 ; a displacer  262 ; a displacer seal  260 ; a hot chamber  276 ; a cold chamber  270 ; a fuel-injection nozzle  406 ; an igniter  408 ; and a hot-chamber temperature transducer  412  which the controller monitors.  
      The heat engine schematic in  FIG. 7B  has a thermal compressor that uses a displacer with a center rod and a linear electromagnetic drive. This heat engine schematic comprises a controller  310 ; a cooler  302  discharging heat into a cooling media  300 ; a liquid-gas separator  391  that separates out water from the products of combustion and discharges the water  393 ; a three-stage expander  380 ; a three-stage compressor  384 ; a heat exchanger  383  that cools gases between compression stages and heats gases between expansion stages; an intake choke  432 ; an air intake filter  385 ; an air intake  390 ; an exhaust  392 ; an exhaust choke  430 ; a high-pressure tank  322 ; a gas drive  329  which outputs shaft power; output shaft  428 ; a discharge check valve  272 ; a low-pressure tank  318 ; an intake check valve  268 ; a three-port, two-position, solenoid cooler valve  340 ; a regenerator  254 ; a fuel tank  396 ; a fuel pump motor  398 ; a fuel pump  400 ; a fuel filter  402 ; an adjustable, fuel flow control valve  404  which is regulated by the controller to maintain a constant temperature in the hot chamber  276 ; an oxygen sensor  361 ; a center-rod, electromagnetic-displacer-drive TC  410 ; a hot chamber  276 ; a cold chamber  270 ; a fuel-injection nozzle  406 ; an igniter  408 ; and a hot-chamber temperature transducer  412  which the controller monitors.  
       FIG. 5B  gives conceptual, temperature versus specific-entropy curves for a continuous-internal-combustion, uncoupled TC. By modulating the fuel flow into the combustion chamber, the TC hot chamber maintains a constant temperature. In  FIG. 5B  the area: 
          1. Q 1 +Q 5  corresponds to the mechanical work extracted from the gases,     2. Q 2  corresponds to the heat extracted from the gases by the cooler,     3. Q 3 +Q 5  corresponds to the heat added to the gases by fuel in the combustion chamber, and     4. Q 3 +Q 4  corresponds to the heat extracted from the gases by the regenerator in one cycle and delivered to the gases in the next cycle.        
      This cycle more closely approximates the Ericsson cycle than that in  FIG. 5A . The efficiency of this cycle is closer to the Carnot cycle since the efficiency of the Ericsson cycle is the same as the Carnot cycle.  FIG. 64  gives the related calculations.  
      The heat engine in  FIG. 7B  is suitable for automotive applications where engine speed varies and often operates for long periods at a small fraction of rated output. Three or four thermal compressors operating in parallel are ideal in such an application since individual units can be turned on or off. Section 3.19 below gives a detailed description of a related automotive drive system.  
      The continuous-internal-combustion, uncoupled, thermal-compressor, gas-turbine engine has the lowest specific weight of the three engine systems in this specification. Because it performs back work, it can have a lower efficiency than the other two engine systems; however, this loss of efficiency can be more than offset by operating at a higher temperature, lowering the ambient pressure and more closely approximating the ideal Ericsson engine by maintaining a constant hot chamber temperature.  
       FIG. 80  is another version of the continuous internal combustion, uncoupled, thermal compressor, gas turbine engine configured for an automotive application. The engine uses high-pressure air in a closed cycle in which a thermal compressor compresses air and a turbine converts compressed air energy into mechanical energy. The engine uses a three stage reciprocal compressor, a three stage reciprocal expander  1583  and thermal compressor displacer  1620 , all on one shaft. An approximate isothermal compression and expansion results by removing heat in each stage of compression and transferred it to gas in each stage of expansion. Three heat exchangers  1544 ,  1610  and  1620 , and an oil spray system described below implement this heat transfer. The reciprocal compressor brings fresh air into the closed cycle and the reciprocal expander removes energy from the products of combustion before discharging then into the atmosphere. The isothermal process minimizes work required to compress air and maximizes energy extracted from the products of combustion during expansion.  
      The engine uses a noise mitigator  1532  to transform the pulsating intake and exhaust gases into a near continuous and thus near noise-free intake  1572  and exhaust  1524  flow processes. The engine uses a cooler  1500  that operates in the closed cycle, condenses water  1646  from the products of combustion and uses the water to absorb pollutants such as sulfur dioxide from the products of combustion.  
      The compressor-expander  1583  uses an oil spray system that lubricates the system and as noted above performs a heat transfer function. Two oil separators  1536  and  1548  discharge oil  1538  and  1550  that is delivered to the oil storage tank  1602 . The oil pump  1600  powered by a motor  1606  receives oil from the oil tank  1602  and delivers it under pressure to spray nozzles in the four compressor chambers  1584 ,  1590 ,  1592  and  1598 , and in the four expansion chambers  1586 ,  1588 ,  1594  and  1596 . Oil entering the compression chambers works its way through and enters the oil separator  1548  where it separates and returns to the oil tank and oil entering the expansion chambers works its way through and enters the oil separator  1536  where it separates and returns to the oil tank. Oil in the compression chambers heats up and oil in the expansion chamber cools down. In the oil tank, the cool and warm oil mix.  
      The compression-expansion pistons and thermal compressor displacer are on a common push rod  1581 . The push rod undergoes translational motion and it spins. The translational motion is required for it to perform its compression and expansion motion. Spin minimizes friction and wear. For a lubricated metal-on-metal surface, the coefficient of friction at low surface velocities is typically near 0.1 where as above 250 cm/s it typically is less than 0.001.  
      The regenerator  1624  is integrated into the close container  1622  so that no part of the closed container structure is subjected to both high temperatures and high tensile stresses. This allows the portion of the structure subjected to tensile stresses to be configured of metals that operate only at low temperature. The high temperature portion of the displacer is subjected only to compressive stresses.  
      Air enters the engine at the intake  1572 , moves through the filter  1570  and flow control valve  1569 , and then enters a noise-mitigator intake chamber  1654 . The products of combustion leave the engine by first entering the noise-mitigator exhaust chamber  1650 , then passing through the exhaust-flow control valve  1531  and exhaust filter  1526 , and then discharged into the atmosphere  1524 . The noise mitigation subsystem simultaneously discharges air to the compressor chamber  1584  and receives products of combustion from the expander chamber  1586 . This occurs when the displacer is moving to the right. When the displacer is moving to the left, the noise mitigator does not receive gas from the expander or discharge it to the compressor. When the noise mitigator discharges and receives gas the noise mitigator piston  1652  moves down, and when it is not receiving or discharging gas the piston  1652  is moving up under the force of the noise-mitigator return-spring  1656 . This motion of the noise mitigator piston largely eliminates a pulsating intake flow and pulsating exhaust flow. If the intake-flow control valve is replaced with a variable flow control valve, upstream pressure sensor and related control system, the variable-flow control valve can be modulated so the intake flow noise is essentially eliminated. In an analogist way, exhaust-related flow noise is essentially eliminated.  
      As the push rod moves to the right air is discharged from the noise, mitigation, intake chamber; moves through piping; the intake check valve  1539 ; and then into the first compressor chamber  1584 . When the push rod begins moving to the left, air in the first compressor chamber  1584  with volume V 1  starts moving through the discharge check valve  1608 , through the heat exchanger  1610 , and then into the second compressor chamber  1590  with volume V 2 . As gas moves from the first compressor chamber to the second compressor chamber the gas with volume V 1  is continuously compressed to a volume V 2  and heat of compression is removed as the gas passes through the heat exchanger. When the push rod again moves to the right the gas in the second compressor chamber  1590  moves through the heat exchanger  1544  and then into the third compressor chamber  1592 . When the push rod again moves to the left the gas in the third compressor chamber  1592  moves through the heat exchanger  1612  and then into the fourth compressor chamber  1598 . Finally, when the push rod moves to the right, again the gas in the fourth compression chamber  1598  is discharged and moves through the check valve  1552 , through the oil separator  1548 , into a pipe with gas coming from the low pressure tank  1649 , through the thermal compressor intake check valve  1556 , and then into the cold chamber  1634  of the thermal compressor.  
      When the displacer  1620  is at the far left the cooler divert valve  1564  is in the position shown in  FIG. 80 . In this position, the pressure in both the hot chamber  1632  and cold chamber  1634  is at the system high-pressure level and the two chambers are connected. As the displacer moves to the right the system pressure drops and when it approximately reaches the system low pressure level the cooler divert valve  1564  is activated. Then gas in the hot chamber  1632  moves through the regenerator  1624 , into the plenum  1630 , through the cooler divert valve  1564 , and then into the cooler  1500 . As the displacer moves to the right the pressure in the cold chamber  1634  drops and when if falls below the pressure of the low pressure tank  1649  gas from the low pressure tank moves first through the intake check valve  1556  and then into the cold chamber  1634 . When the displacer reaches the far right the cooler divert valve  1564  actuates returning it to the configuration shown in  FIG. 80  thus providing a clear passage between the cold chamber and the hot chamber. As the displacer moves to the left, cold gas moves through the regenerator and into the hot chamber, and thus the internal pressure starts rising and when it exceeds the pressure in the high pressure tank  1651  gas is discharged first through the discharge check valve  1554  and then into the high pressure tank  1651 .  
      The hot chamber can be maintained at a near constant temperature by modulating the fuel flow into the hot chamber. The fuel system consists of a fuel tank  1558 , fuel pump  1562 , fuel pump motor  1560 , fuel filter  1566 , variable flow control valve, hot chamber fuel nozzle  1628 , igniter  1626  and thermal couple  1629 .  
      The cooler removes heat from the system and incorporates a water exhaust scrubber. Combustion of automotive fuels and air produces carbon dioxide, CO 2  and water H 2 O among other compounds. For system pressures above 5 MPa condensation of CO 2  is an issue; however, we assume that the system will operate at nominal values of 3 MPa. Due to high system pressures water condensation will occur. This condensation predominately occurs when the thermal compressor hot chamber discharges gases. The cooler  1500  receives almost all of this gas and consequently the cooler will collect this water. The products of combustion, which flow into the cooler, will consist of both liquids and gas. These products of combustion entering the cooler first flow through the heat exchanger  1644 , then discharge through cooler nozzles  1647  at the base of the scrubber chamber  1646  which contains condensed water, then through the liquid-gas separator  1643  which passes the gas and diverts the liquid back into the scrubber chamber. This process partially removes particulates, water-soluble gases and liquids. The liquid is discarded  1534  and the gas moves on to the low pressure tank  1649 .  
      The expander consists of four piston chambers  1596 ,  1594 ,  1588  and  1586  which intake gas, perform three expansion stages and discharge the gas. The expander extracts as much energy as possible from the system gases before discharging them into the atmosphere. It receives gas from both the high-pressure tank  1651  and the low-pressure tank  1649 . Partly receiving gas from the high-pressure tank allows the expander to operate as a drive motor for the engine push rod. The expander high-pressure valve  1542  controls the flow of high-pressure tank gas into the first stage expander. When the push rod is at the far right and just starts moving to the left the expander high-pressure valve is opened allowing gas from the high-pressure tank to flow into the first expander chamber  1596 . After the push rod nominally moves ¼ of a stroke, the expander high-pressure valve closes. Then the gas in the first stage expander chamber expands and the pressure drops. When the pressure in this expander chamber drops below the pressure of the low temperature tank  1649  gas from this tank moves first through the expander intake check valve  1543  and then into the first stage expander chamber. When the push rod reaches the far left position, the pressure in the first stage expander chamber is nominally very close to the pressure of the low-pressure tank. To increase the speed of the engine, the portion of the stroke during which the expander high-pressure valve is open is increased and to decrease the engine speed the portion of the stroke during which the expander high-pressure valve is open is decreased. The configuration of the four-expander valves  1540 ,  1546 ,  1605  and  1616  as shown in  FIG. 80  is for the case when the push rod is moving to the right. When the push rod reaches the far right, these four valves change positions. When the push rod moves to the left gas in the first expander chamber  1596  moves first through the heat exchanger  1612  and then into the second expander chamber  1594 . In this process the gas continuously expands, cools and absorbs heat from the heat exchanger. Similarly, the gas continues to expand as it moves from the second expansion chamber  1594  to the third expansion chamber  1588  and from the third expansion chamber to the fourth expansion chamber  1586 . When the gas discharges from the fourth expansion chamber, it first passes through the expansion valve  1540  and into the oil separator  1536  where oil and gas separates. Oil  1538  returns to the oil tank  1602  and the gas moves to the exhaust chamber  1650  of the noise mitigator  1532 .  
      The push rod is kinematically constrained by a crank  1576  and crank rod  1580  system. The pushrod interfaces with the crank drive with an integral thrust bearing and spin motor  1582 . This interface allows the push rod together with the attached pistons and displacer to spin. This allows maintenance of full lubrication for all bearing and seal surfaces throughout the cycle. This reduces both friction and wear by an order of magnitude.  
      The engine uses four wheel mounted turbine drive subsystems  1502 . Each unit has a forward constant power turbine  1514 ; a reverse constant power turbine  1518  (see Section3:19); two variable flow rate turbines  1512  and  1516 ; a planetary gear speed reduction gear set  1504 ,  1506  and  1520 ; a system pressure containment structure  1503 ; a high pressure seal  1522 ; and a wheel and tire assembly.  
      3.6 Internal-Combustion-Heater, Uncoupled, TC, Gas-Turbine Engine  
       FIG. 8  is a schematic for a continuous internal-combustion-heater, uncoupled, TC, gas-turbine engine. This heat engine comprises a controller  310 ; a cooler  302  discharging heat into a cooling media  300 ; a liquid-gas separator  391  that separates out water from the products of combustion and discharges the water  393 ; a three-stage expander  380 ; a three-stage compressor  384 ; a heat exchanger  383  that cools gases between compression stages and heats gases between expansion stages; an intake choke  432 ; an air intake filter  385 ; an air intake  390 ; an exhaust  392 ; an exhaust choke  430 ; a high-pressure tank  322 ; a gas drive  329  which outputs shaft power; output shaft  428 ; an oil-gas separator  360 ; a discharge check valve  272 ; a low-pressure tank  318 ; an intake check valve  268 ; a three-port, two-position, solenoid cooler valve  340 ; a regenerator  254 ; a fuel tank  396 ; a fuel pump motor  398 ; a fuel pump  400 ; a fuel filter  402 ; an adjustable, fuel flow control valve  404  which is regulated by the controller to maintain a constant temperature in the heater  305 ; a fuel-injection nozzle  406 ; an igniter  408 ; and a heater temperature transducer  411  which the controller monitors; a pressurized, gravity, oil sump tank  362 ; an oil heater  363  that drives off water; an oxygen sensor  361 ; a two-port, two-position oil valve  364  which responds to commands to add oil to the crankcase; crankcase  366 ; crank drive motor  312 ; crankcase oil  368 ; crank arm  319 ; connecting rod  320 ; a pressure-equalizing bellow and low-oil sensor  370  which activates the solenoid oil valve  364  that restores lost crankcase oil; displacer pushrod  357 ; pushrod seal  258 ; a closed container  264 ; a displacer  262 ; a displacer seal  260 ; a hot chamber  276 ; and a cold chamber  270 .  
      The engine version in  FIG. 8 , with a continuous internal combustion heater, has advantages for some applications over the engine versions in  FIGS. 7A and 7B , where continuous internal combustion occurs in the closed container hot chamber. Generally, the continuous internal combustion heater is preferable for large systems. With proper modifications, the use of a “dirty” fuel such as coal in a continuous internal combustion heater is possible. Necessary modifications include a bag filter that removes ash from the products of combustion before they exit the heater and a method of removing condensates such as K 2 O and P 2 O 5  from the regenerator.  
      3.7 Solar-Energy, Thermal-Compressor, Gas-Turbine Power System  
      A schematic of a solar energy based thermal compressor power system is shown in  FIG. 9  and is comprised of solar radiation  1520 ; solar receiver and thermal compressor  1522 ; sun tracking parabolic mirror  1524 ; natural gas heater and thermal compressor  1526 ; low pressure tank  318 ; high pressure tank  327 ; hot water tank and heat exchanger  1538 ; generator  324 ; variable flow turbine nozzle  327 ; and turbine  328 . Section  3 . 18  gives details of the solar receiver shown in  FIG. 10 .  
      The solar-energy thermal compressor operates in parallel with the natural-gas thermal compressor. This allows maximum solar energy use while natural gas provides energy as required. Energy delivered by the thermal compressor is in the form of warm compressed gas and warm gas. In a typical design a pressure ratio of 1.4 is used, compression causes the gas to heat up from an ambient temperature and cooling the compressed gas to the ambient temperature results in a loss of energy. This loss, however, is less than 4% and consequently transferring the compressed gas over long distances does not entail large losses. The warm gas that has not been compressed removes heat engine rejected heat (Q 2  in  FIG. 5A ). Typically, the uncompressed warm gas discharged by the solar receiver transfers about half the solar energy.  
      3.8 Integral Small-Clearance Seal and Gas Bearing  
      The displacer for the system shown in  FIG. 4A  does not use a lubricant other than the working fluid. This is possible since the displacer incorporates a gas bearing. The gas bearing serves the additional role of being a small-clearance seal. A tertiary role to that of gas bearing is a small-capacity regenerator that acts in parallel with the main regenerator. Thus, counter intuitively a small amount of leakage through the seal can improve efficiency. Clearances required for the differential thermal expansion and clearances resulting from pressurization far exceed acceptable leakage values for the seal.  
      A unique, integral, small-clearance seal and gas bearing is shown in  FIG. 11 .  FIGS. 12 and 13  are enlargements defined in  FIG. 11 . These figures describe the seal and bearing concept and do not incorporate a displacer drive system. The seal and bearing concept in  FIGS. 11, 12  and  13  comprise a closed container  264 ; a displacer structure  262 ; a stationary seal cylinder  440 ; a displacer-mounted seal cylinder  442 ; a closed-container back plate  444 ; a displacer back plate  446 ; a cold chamber  270 ; a hot chamber  276 ; a gap between the closed-container structure and the stationary seal cylinder  450 ; a gap between the displacer structure and the displacer mounted seal cylinder  454 ; a gap between the stationary seal cylinder and the displacer mounted seal cylinder  452 ; an entrance for gas to flow between the displacer structure and the displacer seal cylinder  456 ; and an entrance for gas to flow between the closed container and the stationary seal cylinder  458 .  
      This arrangement allows gas pressure to equalize on both sides of the stationary seal and both sides of the displacer seal. Let D designate the diameter of the mean point between the two seal cylinders  440  and  442  as shown in  FIG. 11 . D is the system diameter. The closed-container pressure fluctuates between a high and low value, causing its diameter to fluctuate between a high and low value. Let δ C  designate the difference between the high and low value of this diameter. For many applications, values of δ C /D&gt;0.004 will occur. Similarly, δ D  designates the difference between the high and low value of the displacer structure diameter and δ D /D&gt;0.004 can be expected. Slow-speed, high-value pressure fluctuations, however, will not cause the gap between the two seal cylinders to change significantly. When the displacer moves rapidly, the differential pressure between the hot and cold chambers will increase; however, for most applications this differential pressure should not exceed 30 psi. This pressure will cause the two seal cylinders to reduce the gap between them by an amount designated by δ P . For a 0.2 MPa differential pressure and a typical design, δ P /D&lt;0.0001 with the maximum occurring at a region near the cold chamber.  
      The requirements for the seal material selection process include (1) a near zero coefficient of thermal expansion, (2) good refractory properties, (3) good dimensional stability over time, (4) have good honing and lapping qualities for high-volume precision manufacturing and (5) a tensile stress of at least 140 MPa. The family of glass ceramics offers the most promise; more specifically Pyroceram of the Corning Glass Works can meet these requirements. Pyroceram can have a coefficient of thermal expansion of 0.36×10 −6  cm/cm-° K., a softening point of 1300° C., a flexural tensile strength of 207 MPa and is used for telescope mirror blanks that require a high degree of dimensional stability and good honing and lapping qualities.  
      In a typical design, the maximum temperature difference between the moving displacer seal cylinder and the stationary seal cylinder can be 200° K. Let δ T  designate the maximum difference in the change in diameters between the displacer seal and the stationary seal due to thermal expansion. Then δ T /D&lt;0.0001 and the maximum value of δ T /D occur at the end of the hot chamber.  
      The seal leakage is laminar flow and easily calculated. Let δ S  designate the difference in diameter between the two seal cylinders. With normal high volume production an out of round, taper and waviness tolerance of 0.001&gt;δ S /D&gt;0.0001 can readily be achieved. This value results in a leakage rate that will not cause a significant loss of efficiency for most applications.  
      The gas-bearing design is different from most gas-bearing applications by virtue of the very high-pressure gas used. As the pressure goes up the viscosity of the gas changes only a small amount; however, the compressibility of the gas increases proportionally to the pressure. At 100 atmospheres, the gas, in terms of how it affects the bearing characteristics, behaves more like an incompressible fluid than a gas. A smooth bushing type bearing should remain stable for most applications. If bearing stability is an issue, a herringbone-groove journal bearing design of a type shown in  FIG. 14  is used. This design is very stable down to much lower pressures. The herringbone-groove journal bearing shown in  FIG. 14  is comprised of a journal or shaft  460  with helically shaped grooves  464  at each end and a bearing  462 .  
      3.9 Displacer Center Rod Support  
      A low cost alternative to the integral small clearance and gas bearing as a means of minimizing TC wear is the use of a displacer center rod. This engine component, described in more detail below and in  FIGS. 22A and 22B , is a version that uses a displacer labyrinth seal that has a small differential pressure and does not support the displacer relative to the closed container. The bushing that supports the displacer center rod can wear so that the labyrinth seal will in turn wear. By using the motorized bushing shown in  FIG. 22B  the wear problem is overcome. This motorized bushing maintains an oil film between the bushing and center rod and maintains a centering force on the center rod. Therefore, bushing wear and labyrinth seal wear is not significant after an initial wear in period. The center-rod oil seal can experience wear; however, the pressure across this seal is essentially zero so that maintaining a continuous oil film between the seal surface and center rod is not difficult.  
      The center rod and displacer drive in  FIG. 22A  comprises a displacer  523  supported by a lubricated slender center rod  528  with center channel  525 , and end ports  512  and  520 ; center rod bushing  514  with busing base cavity  513 ; magnet  516  and magnetic circuit iron elements  515  and  517 ; displacer drive coil  519  with leads  518 ; displacer base  526 , seal  521 , and insulation  522 ; and closed container  524  with cold chamber  270  and hot chamber  276 . The center rod  528  has a center channel  525  with openings at both ends  512  and  520  so that the gas pressure at the base of the center rod  513  equalizes with the pressure of the closed container cold chamber  270 . The drive coil leads have a second function by being nonlinear springs that define the stroke length by bouncing the displacer (not shown). Sensors at each end of the displacer drive coil (not shown) cause current reversal when the displacer reaches the end of the stroke. A power supply powers the displacer drive coil.  
      The motorized bushing shown in  FIG. 22B  comprises a slender center rod  528  with center channel  525 , a cylindrical bushing and integral permanent magnet  531 , motor winding and iron core structure  532 , displacer magnetic circuit iron elements  515 , motor winding and iron core structure threaded fastener  539 , center rod oil seal  535 , and center rod oil seal threaded fastener  534 .  
      An advantage of the center rod version is that it is simple and low cost. A disadvantage is that it is not maintenance free and contamination of closed systems is possible.  
      3.10 Permanent Magnet, Displacer-Drive System  
      This specification addresses three linear electromagnetic drive systems, e.g., two low-cost ones based on permanent magnets and a high-performance one that is completely electromagnetic. In volume production use of low-cost permanent magnets such as strontium ferrite results in a low-cost device; however, high-cost permanent magnets, such as ones based on neodymium-iron-boron alloys, may result in devices that are more costly than completely electromagnetic devices. Ultimately, electromagnetic drives can provide higher performance and do not suffer from potential demagnetization problems.  
       FIG. 15  is a conceptual drawing of a permanent magnet, displacer-drive system. It comprises a stationary, spin-coil assembly  480 ; stationary, linear-drive-coil assembly  482 ; Hall sensors  483 ; displacer-permanent-magnet cylinder  484 ; cold chamber  270 ; closed-container structure  264 ; displacer structure  262 ; and hot chamber  276 .  FIGS. 16, 17 ,  18  and  19  describe the thick-walled-cylinder permanent magnet magnetized to create magnetic poles for the spin motor on the inside and the drive motor on the outside.  FIG. 16  comprises a section view of the permanent magnet  484  and  FIG. 17  is a section view defined in  FIG. 16 .  FIG. 18  is an enlargement defined by  FIG. 16  and  FIG. 19  is an enlargement defined by  FIG. 17 . These figures show the magnetized region consisting of ring-shaped poles  490  and the magnetized region of bar-shaped poles  492 .  
      Drive-coil windings and position sensors include a three-phase winding or a two-phase, push-pull winding for the linear drive force and these include one or more electrical position transducers or electro-optical position transducers.  FIG. 20  shows a segment of the stationary, linear-drive coil assembly with a two-phase, push-pull winding that induces longitudinal motion.  FIG. 20  comprises a stationary, drive-coil assembly  482 ; a permanent magnet with linear-drive magnetized regions  490  and spin-motor magnetized regions  492 ; leads to the phase-A coils  506 ; leads to the phase-B coils  508 ; leads to position transducer  510 ; phase-A drive coils with a positive polarity  494 ; phase-A coils with a negative polarity  496 ; phase-B drive coils with a positive polarity  498 ; phase-B coils with a negative polarity  500 ; and a position-transducer coil  502 .  FIG. 21A  shows a two-phase, push-pull, spin-drive-motor configuration using 6 permanent-magnet rotor poles, 12 stator poles and 2 Hall sensors.  FIG. 21B  is a basic circuit for a two-phase, push-pull, brushless DC motor.  
      Another type of displacer drive uses a sound speaker like coil and magnet system as shown in  FIGS. 22A and 22B . Section 3.9 gives a description of this displacer drive.  
      3.11 Electromagnetic, Displacer-Drive System  
      The electromagnetic, displacer drive is in essence a brushless, linear motor when it accelerates the displacer and a brushless, linear generator when it decelerates the displacer. Gas-dynamic forces associated with compression at either end of the stroke decelerates the displacer; however, electromagnetic, decelerating forces can shape the overall, decelerating force into a sinusoidal shape as part of a vibration-control system described below.  FIG. 23  gives a conceptual description of the electromagnetic, displacer drive and comprises a set of nested, stationary cylinders  547 ,  557 ,  555 ,  552  and  549 ; a set of nested cylinders attached to the displacer and interlaced with the stationary cylinders  558 ,  556 ,  553  and  551 ; and displacer position sensor including a lamp  548 , longitudinal slit  551 , light receiver  550  and Hall sensors  559 .  
      The electromagnetic, displacer-drive system has one moving part, e.g., the displacer. The displacer, levitated on a gas bearing, uses electromagnetic forces to induce oscillatory longitudinal motion. Therefore, no surface contact is required. The displacer speed is thermal limited, i.e., overheating of the drive coils limits the system speed. The drive system of the displacer consists of a set of nested cylinders, which during operation causes high velocity helium or hydrogen or nitrogen to move across the surface, resulting in a high capacity, coil-cooling system. Thus, this arrangement is suited to high-performance operation. This cooling function also retards the displacer; however, the largest retarding force is typically associated with the pressure drop across the regenerator. As described above, the displacer uses gas-dynamic forces to bounce off the ends of the closed-containment structure. Displacer retarding forces include the pressure drop across the regenerator and drive coil forces. Here the coils dissipate regenerator energy only as part of a vibration-control system. The resulting electrical energy can be stored in a capacitor and reused, or dissipated in a resister. It follows that minimizing displacer weight reduces displacer kinetic energy and improves efficiency.  
       FIG. 24  is a wiring and position-sensor schematic for the electromagnetic, displacer-drive system. The elements in the figure are grouped into five blocks, e.g.: (1) the stationary-exciter circuit group  554  that includes the exciter winding  566  which induces power in the displacer circuits; (2) the displacer-mounted circuit group  558  comprising a three-phase exciter winding  568 , a solid-state rectifier bank  570 , a displacer-mounted, translation-drive winding  572 , and a displacer-mounted, spin-drive winding  574 ; (3) the stationary, translation-circuit and spin-circuit group  560  comprising phase A translation drive windings  576 , phase B translation drive windings  577 , phase A spin drive windings  578 , and phase B spin drive windings  579 ; (4) the group comprising the external controller, power-conditioning circuits, power supply and spin-drive circuits similar to those in  FIG. 21562 ; and (5) displacer-position-sensor group  556  comprising a lamp  588 , a light receiver  587  and a translation slit  589 .  
      The stationary-exciter coil  566  in  FIG. 24  corresponds to the cylinder  552  in  FIG. 23 . The stationary exciter forms a set of magnetic poles on the cylinder  552 . It has two modes of operation, e.g., the spin-starting mode and the spin-running mode. If the displacer is rotating, a DC current in the stationary-exciter coils  566  will induce a current in the three-phase, displacer exciter winding  568  which is rectified through the rectifier bank  570  and then delivered as a uniform DC current through displacer, spin windings  574  and displacer, drive windings  572 . An AC current in the exciter winding  566  induces a current in the displacer spin and drive windings when the displacer is not spinning. Thus, an AC exciter current initiates displacer spinning and is then switched to a DC current. The stationary-spin windings  578  and  579  form part of a two-phase, push-pull spin motor and the stationary-translation windings  576  and  577  form part of a two-phase, push-pull translation drive.  
      The capacitor  582  stores electrical energy extracted from the displacer kinetic energy during deceleration and then reused. A resister  584  also dissipates kinetic energy. By shaping the acceleration curve to a sinusoidal shape, the task of noise and vibration reduction is simplified (no harmonic frequencies). Generally, any acceleration curve shaping will reduce operating efficiency slightly; however, this can sometimes be justified in order to achieve optimal noise and vibration reduction.  
       FIG. 25  presents a section view of an electromagnetic-drive, thermal-compressor assembly. This figure is descriptive and not an engineering drawing.  FIG. 25  includes the hot, closed-container segment  218 ; the cold, closed-container segment  220 ; valves  212 ,  216 ,  203 ,  204 ,  205 ,  222  and  224 ; regenerator  208 ; pressurization vessel  228 ; the discharge to the heater  214 ; the return from the heater  210 ; the discharge to the cooler  206 ; the intake from the low-pressure tank  202 ; the discharge to the high-pressure tank  226 .  FIGS. 26 and 30  are section views and enlargements defined in  FIG. 25 .  FIGS. 27, 28  and  29  are enlargements defined in  FIG. 26 .  FIGS. 25, 26  and  30  callouts correlate to those in  FIG. 4A . Thus the gas intake  226  receives gases coming from the low-pressure tank  318 ; the gas discharge  202  sends gases to the high-pressure tank  322 ; the cooler, solenoid poppet valve  204  corresponds to the schematic valve  340 ; the cold-chamber, bounce poppet valve  224  corresponds to the schematic valve  347 ; the intake, check valve  203  corresponds to the schematic valve  268 ; the discharge, check valve  205  corresponds to the schematic valve  272 ; the return check valve  222  coming from the heater corresponds to the schematic valve  342 ; the heater, solenoid poppet valve  212  corresponds to the schematic valve  344 ; the regenerator, solenoid, poppet valve  216  corresponds to the schematic valve  348 ; the gas discharge  214  discharges gas from the hot chamber  276  to the heater  304 ; and the heater gas return is  210 . The pipe outlet in  FIGS. 26 and 30  designated by  230  are connected. A description of valve and regenerator details appears below.  
       FIG. 27  is an enlargement defined by  FIG. 26  and comprises the ambient, closed-containment structure  590 ; the cylindrical, stationary, integral, small-clearance seal and gas bearing  592 ; displacer-mounted, seal and bearing cylinder  610 ; outer, stationary, translation-drive iron core  612  and windings  614 ; displacer-mounted, winding support structure  594 ; displacer-mounted winding  596 ; inner, stationary, translation-drive iron core  616  and winding  618 ; displacer-mounted, coreless-spin-motor winding  598 ; displacer-mounted, spin-motor-winding support structure  620 ; spin-motor stator winding  622 ; displacer-mounted, coreless, exciter-winding support structure  600 ; displacer-mounted, coreless, exciter winding  602 ; stationary, exciter winding  606 ; stationary, exciter iron core and lamp support  608 ; cylindrical, optical-slit support structure  626 ; and light-receiver support rod  628 .  
       FIG. 28  is an enlargement defined by  FIG. 26  and comprises the cold, closed-container segment  590 ; hot, closed-container segment  648 ; the cylindrical, stationary, integral, small-clearance seal and gas bearing  592 ; displacer-mounted, seal and bearing cylinder  610 ; displacer-mounted, seal and bearing cylinder, attachment ring  654 ; displacer-mounted, drive-attachment structure  656 ; displacer-mounted, translation-drive-winding support structure  596 ; displacer-mounted, coreless, spin-motor-winding support structure  620 ; displacer-mounted, exciter-winding support structure  602 ; displacer-mounted, optical-position-sensor, slit cylinder  626 ; a radiation cone  652 ; and a displacer, ceramic, vacuum bulb  650 .  
       FIG. 29  is an enlargement defined by  FIG. 26  and comprises a cold-chamber, bounce valve  224 ; a valve cover  670 ; upper, corrugated, disk spring  672 ; valve-stem, magnetic iron  674 ; solenoid winding  676 ; solenoid back iron  678 ; lower, corrugated disk spring  680 ; poppet valve  716 ; valve casing structure  714 ; cold-chamber port  700 ; closed-container, cold-chamber back plate  702 ; ambient-temperature, closed-containment structure  590 ; cylindrical, stationary, integral, small-clearance seal and gas bearing  592 ; the gap that accommodates the displacer-mounted, seal cylinder  704 ; the gap that accommodates the displacer-mounted, translation, drive winding  706 ; the gap that accommodates the displacer-mounted, spin-motor winding  708 ; the gap that accommodates the displacer-mounted, exciter winding  710 ; the gap that accommodates the displacer-mounted, position-sensor, slit cylinder  712 ; back-plate-mounted rod which supports the position-sensor light receiver  628 ; back-plate-mounted, exciter-winding support structure  606 ; back-plate-mounted exciter winding  608 ; back-plate-mounted, spin-motor-winding support structure  624 ; back-plate-mounted, spin-motor winding  622 ; back-plate-mounted, translation, inner-drive, core iron  616 ; back-plate-mounted, translation, outer-drive, core iron  612 ; back-plate-mounted, translation, drive winding  614 ; and back-plate-mounted, translation, inner-drive, core iron  616 .  
       FIG. 30  is a section view of the high-temperature end of the displacer and closed container, comprising a vacuum bulb  219 ; an integral, small-clearance seal and gas bearing  610 ; two radiation cones  722 ; high-temperature end of the closed container  648 ; low-temperature end of the closed container  590 ; heater return check valve  222 ; regenerator solenoid poppet valve  212 ; heater solenoid poppet valve  216 ; discharge-to-heater pipe  214 ; regenerator, solenoid, poppet-valve stem  752 ; and heater, solenoid, poppet-valve stem  756 . The vacuum bulb uses a vacuum in order to eliminate convective heat losses and requires a refractory material with a low coefficient of thermal conductivity such as refractory mullite. The two radiation cones with highly reflective surfaces essentially eliminate radiation heat losses. The radiation cones also require refractory material, although the temperature and stress requirements are less demanding, and Pyroceram is used. To insure low radiation losses the interior of the vacuum bulb and both sides of the radiation cones should receive a vapor deposited metallic coating. The interior temperature of the vacuum bulb will typically be below 1000° C. and consequently the reflective surface is copper since almost all the reflective energy will be in the infrared portion of the spectrum.  
      3.12 Balancing Systems for Electromagnetic Drive TCs  
       FIG. 31A  is a schematic of an active, vibration-mitigation system that can be used with any of the electromagnetic-drive thermal compressors. In essence, the active damper induces a vibration that cancels the vibration caused by the displacer. The schematic in  FIG. 31A  comprises a closed container and drive coil structure  264  that interacts with the displacer and is subjected to longitudinal motion  816 , a displacer and drive coil assembly  262  that interacts with the closed container structure and is subjected to longitudinal motion  818 , a vibration isolation spring  814  that is attached to the closed container structure  264  and system support plate  812 , an active damper drive coil and structure  806 , and an active damper armature  810  that is subjected to longitudinal motion  808 . By making the spring  814  soft, the force transferred from the closed container structure  264  to the system support plate  812  is small and the mass of the active damper armature  810  required to nullify system vibrations is small.  
       FIG. 31B  is a section view for an integrated, gas-compressor, system assembly using four electromagnetic-drive TCs; a heater; a pressurization vessel; and a tilted-disk, vibration-mitigation subsystem. This figure is a concept drawing and does not include many details. This figure shows the invisible lines for three displacers.  FIG. 31B  comprises a tilted-disk, drive-motor chamber  760 ; vibration-mitigation-system, containment vessel  762 ; tilt-disk drive motor  764 ; angular-position sensor and slit disk  766 ; tilt-plate chamber  768 ; aft, tilt-plate, shaft, bushing support plate  804 ; front, tilt-plate, shaft, bushing support plate  800 ; tilt plate  802 ; aft clamp  798 ; closed-container structure  778  and  796 ; displacers  770 ,  792  and  794 ; insulation  774 ; aft section of pressurization vessel  776 ; front clamp  780 ; front section of pressurization vessel  782 ; heater  784 ; heater intake  786 ; heater exhaust  788 ; and pressurized-chamber intake port  790 . If the four displacers have a sinusoidal motion and are circumferentially sequenced  900  out of phase, then the tilt plate can effectively nullify displacer-based vibrations.  
      3.13 Crank-Drive, TC Assembly  
       FIG. 32  describes a TC together with its valves, regenerator and drive system. The system incorporates a symmetric-double-crank displacer drive. The configuration eliminates vibrations associated with the primary frequency and the second harmonic. The displacer seal uses seal rings and pressurized oil. An oil pump that uses two check valves and is integral to the push rod is used. The push rod uses seal rings. This configuration uses a pressure balance subsystem described in Section 3.4, which effectively prevents a pressure differential between the cold chamber  827  and the crankcase pressure  837 . The elements in  FIG. 32  conform to the schematic in  FIG. 6 .  
       FIG. 32  comprises an oil sump  834 ; a high-pressure tank  832 ; gas-liquid separator screen  830 ; high-pressure tank intake  829 ; high-pressure tank  831 ; high-pressure, tank discharge  836 ; crank-case chamber  837 ; thermal-compressor, cold chamber  827 ; displacer insulation  826 ; high-temperature, displacer shell  824 ; high-temperature, displacer chamber  825 ; high-temperature, closed-container shell  832 ; discharge-to-heater  214 ; heater, poppet valve  838 ; heater-return pipe  210 ; heater-return check valve  222 ; regenerator  208 ; discharge-to-cooler  206 ; cooler poppet valve  204 ; gas-intake pipe  226 ; intake check valve  227 ; discharge check valve  828 ; and discharge pipe  202 .  
       FIG. 33  is an enlargement defined in  FIG. 32  and comprises a crankcase, gas-discharge check valve  904 ; bellow sensor  906  used to activate the oil crankcase, refill, solenoid valve when the bellow contacts the sensor; bellow chamber  902 ; pressure-balance bellow  900 ; bellow gas intake  910 ; bellow-chamber oil intake  898 ; right crank and counter weight  896 ; left crank and counter weight  864 ; right-crankshaft  897 ; left-crankshaft  867 ; right-crank drive gear  892 ; left-crank drive gear  870 ; transfer gear  894 ; motor-drive gear  866 ; drive motor  868 ; right-connecting rod  890 ; left-connecting rod  874 ; crank cross bar  876 ; crankcase casting  856 ; displacer seal ring  858 ; cold chamber  827 ; ring-seal, oil-feed path  913 ; ring-seal, oil-return path  911 ; closed-container, high-temperature shell  854 ; displacer, high-temperature shell  852 ; displacer insulation  850 ; crankcase chamber  878 ; refill, solenoid-valve yoke  880 ; solenoid-valve coil  881 ; valve stem  882 ; valve-return spring  877 ; crankcase-oil intake  883 ; intake from oil sump  884 ; oil-pump-intake check valve  885 ; oil-pump-discharge check valve  895 ; oil-pump plunger and oil-feed pipe  886 ; and oil-pump interior cavity  887 .  
      3.14 High-Performance, Ceramic Regenerator  
      A regenerator is a vessel containing a porous media and a path though the media. It absorbs heat from one cycle and uses it for the subsequent cycle. The porous media is a solid with a surface area portion A 1  over where hot gas enters the media and a surface area portion A 2  over where cold gas exits the media. Let A=(A 1 +A 2 )/2, i.e., A is the average of the two areas and will be referred to as the regenerator area. Let L, called the regenerator length, designate the average distance between the point at which a gas particle enters the porous media and the point at which it exits. Let ΔP designate the pressure drop across the regenerator. Let v 1  be the gas velocity just before it enters the media at surface A 1  and let v 2  be the gas velocity just as it leaves the media at surface A 2 . The absolute temperature of the hot side of the regenerator can be four times that of the cold side. The velocity of a gas particle as it moves through the regenerator is approximately proportional to its temperature. Let v be the average velocity of the gas particle as it passes through the regenerator. The equation in  FIG. 65  gives the pressure drop in porous materials due to gas flow. If we let q be the average volume flow rate of the gas through the regenerator, then q=Av. It follows that for a fixed flow rate ΔP decreases as L decreases and A increases. As L decreases, thermal-conduction heat loses increase, and an optimal design balances these loses.  
       FIG. 34  is a section view of a high-performance regenerator and  FIG. 35  is a section defined by  FIG. 34 .  FIGS. 34 and 35  comprise a spherical ceramic shell  922 , externally pressurized so that the stresses in the sphere are always in compression; insulation  924 ; low-temperature inlet  920 ; high-temperature inlet  928 ; eight segments of heat-absorbing media  926 ; and the high-temperature inlet fan  930 . The heat absorbing media is a square plate of thickness L and area A/8. Thus, the eight heat-absorbing plates have a total gas throughput area of A.  
      The regenerator can perform the additional function of an oxidation catalytic converter by depositing particles of catalytic material on the heat absorbing material. Typically, noble metals are used for this purpose. A mixture of platinum and palladium is most commonly used.  
      3.15 Valve Assembly  
       FIGS. 36 and 37  are enlargements of valve assemblies shown in  FIG. 25 .  FIG. 32  is a ceramic valve assembly and corresponds to the  FIG. 32  valve callout  838 , and  FIG. 37  is a metallic valve assembly and corresponds to valve callout  204  in  FIGS. 25 and 32 .  
      The ceramic valve assembly in  FIG. 36  minimizes heat flow along the valve stem  950  and stem case  952  by minimizing the area through which heat flows, increasing the length of the stem and using materials with a low coefficient of thermal conductivity.  FIG. 36  comprises a valve cover  942 ; upper, corrugated, disk spring  944 ; valve-stem iron  940 ; solenoid winding  945 ; solenoid back iron  941 ; lower, corrugated, disk spring  946 ; upper, steel, valve-stem rod  948 ; ceramic, valve-stem rod  950 ; valve case  952 ; heater, discharge pipe connector  954 ; thermal-compressor, hot-chamber pipe connector  956 ; valve-access cover  958 ; poppet valve  960 ; check valve  962 ; regenerator pipe connector  964 ; and heater, return-pipe connector  966 .  
      The metallic valve assembly in  FIG. 37  comprises a solenoid winding  980 ; back iron  982 ; valve-stem iron  984 ; upper, corrugated disk spring  988 ; lower, corrugated disk spring  990 ; valve stem  992 ; valve-assembly body  994 ; pipe-connector for pipe that discharges to cooler  996 ; pipe-connector for pipe that discharges to regenerator  998 ; poppet valve  1000 ; valve-access cover  1002 ; pipe-connector for pipe that discharges to low-pressure tank  1004 ; thermal-compressor, discharge check valve  1006 ; pipe-connector for pipe that connects to thermal-compressor cold chamber  1008 ; thermal-compressor, intake check valve  1010 ; and pipe-connector for pipe that discharges to high-pressure tank  1012 .  
      3.16 Efficient Low-Emission Heater  
       FIG. 38  is a schematic for a heater with very low emissions, high efficiency (heat transferred to working fluid)/(heat available in fuel), and low specific volume (heater volume)/(rated power).  
      Emissions are controlled by (1) limiting the combustion temperature to a value below where significant NO x  compounds are formed and below where dissociation of CO 2  occurs in large amounts, (2) requiring combustion to occur with a significant excess of oxygen, (3) maintaining products of combustion at an elevated temperature for a significant time so that combustion of all fuel elements is essentially complete, and (4) lowering the temperature of the products of combustion slowly so that dissociation of CO 2  does not result in significant residual CO in the exhaust products.  
      To estimate relative exhaust NO levels as a function of combustion temperature consider the equation given in  FIG. 66 . A nominal operating combustion temperature of the engine in this specification is 1300° K. and a nominal combustion temperature for an automotive internal combustion engine is 2700° K. The N 2  and O 2  concentration levels for both engines are nominally the same, therefore, the initial NO rate of formation for the engine in this specification relative to an automotive internal combustion engine as shown in  FIG. 66  is 1.55×10 −12 . Thus, the NO formation rate is insignificant and, unlike an automotive internal combustion engine, no catalytic converter is required to control NO x  compounds.  
      To achieve a high efficiency, exhaust gases are used to preheat the intake air to a temperature above that required for spontaneous combustion with fuel. Adding fuel increases the temperature to the operating combustion temperature. Each combustion stage typically uses less than 10% of available oxygen. These heated gases then pass through a heat exchanger module, which will heat the thermal compressor working fluid as it cools the products of combustion. The repeated sequence adds fuel to increase the combustion temperature and then transfers heat with a heat exchanger module, to the working fluid, consuming 80% of the oxygen.  
      The schematic in  FIG. 38  comprises a fuel subsystem with a fuel tank  1020 , fuel pump  1024 , fuel pump motor  1022 , fuel filter  1026 , and fuel adjustable flow control  1028 ; a fuel electrical heater  1031  which is used at start up to heat fuel to a high enough temperature so that instant combustion occurs in the combustion chamber; a switch  1025  that initiates electric fuel heating; a battery  1033  that powers the electric fuel heater; an air-feed subsystem with an intake  1052 , air filter  1050 , positive displacement air pump  1054  and air pump motor  1056 ; recuperator  1044  that transfers heat from the exhaust gases  1062  to the intake air and fuel  1060 ; a combustion chamber  1030  that receives preheated air from the regenerator, receives vaporized and preheated fuel also from the regenerator and through the fuel nozzle, and discharges products of combustion through the heat exchange module  1034 ; a series of nine additional and identical combustion chambers and heat exchange modules through which the products of combustion pass and discharge into an ignition chamber  1035 ; an ignition chamber with an igniter  1036  which initiates the combustion process and which receives products of combustion from the last heat exchange module and delivers them to the regenerator; a temperature sensor  1042  which measures the temperature of the products of combustion as they exit the ignition chamber and is used by the controller to regulate the fuel flow rate; an oxygen sensor  1046  which measures the oxygen level of the products of combustion just before they are discharged in the exhaust  1048  and which is used to regulate the air flow rate; and an intake  1038  for the working fluid from the thermal compressor which passes through the heat exchange modules and then is discharged back to the thermal compressor  1040 .  
       FIG. 39  is a section view of a heater subassembly characterized by the  FIG. 38  schematic and comprising an exhaust-gas, heat recuperator; ten heat exchange modules, combustion chambers and fuel injectors; and an igniter.  FIG. 40  is a section of the exhaust-gas, heat recuperator defined in  FIG. 39 .  FIG. 41  is a section view of the heater defined in  FIG. 39 . This heater is a monolithic structure formed by sintering a set of component parts.  
      The  FIG. 39  heater is comprised of 10 heat exchange modules  1080 ; 10 baffle plates  1082  that direct gas products of combustion exiting a heat exchanger module to the fuel nozzles; 10 combustion chambers  1088 ; 10 fuel feed pipes  1083 ; 10 fuel nozzle sets  1084 ; 10 combustion chamber ceramic shell segment  1086 ; 10 heat exchanger end wedge block  1081 ; exhaust gas heat recuperator structure  1094 ; heater recuperator exhaust  1096 ; heater recuperator intake  1098 ; recuperator exhaust baffles  1102 ; recuperator exhaust cross channel  1100 ; igniter  1104 ; recuperator cross feed start chamber  1105 ; an exit chamber that feeds the recuperator  1106 ; an intake chamber that receives gas from the recuperator  1107 ; a separation plate  1120  that separates the intake; and exhaust gases; combustion-chamber and an intake tube that receives working fluid from the thermal compressor  1108 .  
       FIG. 40  comprises a separation plate  1120  that separates the intake and exhaust gases; combustion-chamber, ceramic shell segment  1086 ; fuel feed pipe  1083 ; recuperator, exhaust cross channel  1100 ; intake longitudinal channels  1122 ; exhaust-gas-heat recuperator structure  1094 ; heater recuperator intake  1098 ; and igniter  1104 .  
       FIG. 41  comprises a heat exchanger module  1080 ; a heat-exchanger, working-fluid intake channel  1160 ; a heat-exchanger, working-fluid discharge channel  1146 ; end plate  1162 ; intake pipe, exhaust pipe and end plate structure  1144 , working-fluid exhaust pipe  1148 ; working-fluid intake pipe  1150 ; combustion-chamber ceramic shell segment  1086 ; heat-exchanger end wedge block  1081 ; baffle plates  1082 ; baffle-plate channel  1152 ; fuel nozzles  1084 ; and fuel feed pipes  1083 .  
      The heat-exchanger module is a heat exchanger that transfers heat from the high temperature products of combustion typically at a pressure of one atmosphere to the thermal compressor working fluid, which is at a much higher pressure and can nominally be at 100 atmospheres. In addition, working fluids such as hydrogen can transfer heat more efficiently than the products of combustion at identical pressures. Ideally, such a heat exchanger will have a much higher area for the products of combustion to transfer heat compared to the working fluid area (nominally by a factor of 100). A unique monolithic ceramic structure formed from layers of ceramic cloth and ceramic pipe provides a large difference in area and operates at temperatures that may exceed 1100° K. The ceramic cloth fiber provides a very high surface area in a small volume.  
       FIGS. 42 through 46  give detailed descriptions of the heat exchanger module. This module comprises a two-element sequence of layers characterized as a ceramic cloth layer, a ceramic pipe layer, a ceramic cloth layer, a ceramic pipe layer, etc., which start and end with a ceramic cloth layer and which are sandwiched between a front and back ceramic layer. These layers are compressed and sintered to form a monolithic ceramic structure.  
       FIG. 42  is an isometric explosion of the heat exchanger structure comprising a front plate  1170 ; a back plate  1170 ; three cloth plates  1172 ; and two pipe plates  1174 . Products of combustion flow normal to the plate in the direction  1178 , and the thermal compressor working fluid flows in the orthogonal direction  1176 .  FIGS. 43, 45 , and  46  are platelayer enlargements defined in  FIG. 42 .  FIG. 43  is an enlargement of a small segment of the front plate, and comprises the ceramic structure  1190  and rectangular shaped gas passages  1192 . The rectangular gas passages will typically be quite small with nominal dimensions of say 2 mm by 8 mm.  FIG. 44  is a section view defined in  FIG. 43 .  FIG. 44  comprises a ceramic front plate  1190 ; rectangular-shaped gas passages  1192 ; a cloth fiber structure  1200  formed from a number of ceramic fiber cloth layers; contact surface between 2 pipe plates  1202 ; a pipe plate  1206 ; a pipe plate pipe  1210 ; and another pipe plate pipe  1212 .  FIG. 45  is a section defined by  FIGS. 42 and 44 , and comprises contact surfaces  1202  between 2 pipe plates and cloth-fiber-structure surface  1200 .  FIG. 46  gives a face view of the pipe plate defined in  FIG. 42 , and comprises one pipe sequence  1210 , another pipe sequence  1212 , rectangular shaped gas passages  1192 , and contact surfaces  1202 .  
      3.17 Monolithic Ceramic Heater  
      A unique heater concept called a monolithic ceramic heater, configured by modifying the ceramic heat exchange module ( FIG. 42 ) by including fuel pipe platelayers. The fuel pipe platelayer is either porous so that fuel diffuses though the pipe wall, or each fuel pipe has a set of holes through the pipe wall so that fuel diffuses throughout the monolithic structure. Thus, the monolithic ceramic heater is comprised of a three-plate sequence characterized as a cloth layer, a working fluid pipe plate layer, a fuel pipe plate layer, a cloth layer, etc, together with a front plate and an aft plate. Other equivalent sequences using the three types of interior plates exist.  
      3.18 Solar Receiver with Thermal Compressor  
      A solar receiver is a heater that uses solar energy as the heat source.  FIG. 9  describes a solar-energy, thermal-compressor, and gas turbine power system.  FIG. 10  describes a thermal compressor integrated with a solar receiver. The receiver uses ceramic elements in order to operate at high temperatures and uses a unique design concept that maintains all ceramic components stresses in compression.  FIG. 10  comprises concentrated sunlight  1224  from a parabolic mirror; solar-receiver, inner ceramic sphere  1232  that has a circular window, absorbs sunlight energy on the inner surface and is pressurized on the outer surface by the working fluid; outer ceramic sphere  1228  which is pressurized on the inner surface by the working fluid and pressurized on the outer surface by the working fluid to the peak system pressure; ceramic-fiber-filled, working-fluid cavity  1220  which diffuses energy absorbed by the inner ceramic sphere into the working fluid in the volume between the inner and outer ceramic spheres; outer, tensile stress sphere  1222  which uses a material with good tensile stress properties; thermal insulation  1226  which minimizes heat loses to the atmosphere; insulating compression cone  1230  which seals gases and transfers the unbalanced load on the inner ceramic sphere due to the circular window; a/thermal compressor  346 ; a high-pressure discharge check valve  272 ; a high-pressure discharge pipe  1238 ; an intake check valve  268 ; an intake pipe  1240 ; a three-port, two position valve  340 ; a low-pressure, heat-discharge pipe  1242 ; a regenerator  254 ; and a solar-receiver, pressurization check valve  1236  which insures that the outer tensile sphere is pressurized to the peak operating pressure.  
      3.19 Constant-Power Turbine  
      The TC of the type under consideration in this specification outputs gases at temperatures that are typically below 300° C. Consequently, material strength degradation due to heating of elements subjected to these gases is not an issue. In addition, these TCs will typically input gases at a pressure of 100 atmospheres, which permits small turbines to have a high power output. The constant-power turbine is an innovative device that is especially useful for low-temperature high-pressure systems. This device outputs a constant power over a broad speed range. It exploits the concept of “velocity compounding” described in  FIGS. 47 and 48 .  
      The right side of  FIG. 47  shows three turbine blades  1252 ,  1256 , and  1260  each belonging to a different turbine wheel but connected to a common shaft; two stator blades  1254  and  1258  each belonging to a different stator wheel; a converging diverging nozzle  1250 ; and the trajectory of a particle  1264  from the nozzle and through the blades. The left side of  FIG. 47  is a velocity diagram for a gas particle moving through the blades. The turbine-blade velocity vector is designated by U, the velocity vector of a gas particle as it exits the nozzle is designated by V 1 , the velocity vector of a gas particle as it exits the first turbine blade is designated by V 2 , the velocity vector of a gas particle as it exits the first stator blade is designated by V 3 , the velocity vector of a gas particle as it exits the second turbine blade is designated by V 4 , the velocity vector of a gas particle as it exits the second stator blade is designated by V 5 , and the velocity vector of a gas particle as it exits the third turbine blade is designated by V 6 .  
      Note that a bold letter designates a vector. A scalar component tangent to the turbine blade trajectory V jt  and a scalar component normal to the turbine disk plan V jn  can represent each vector V j . Then (V jt , V jn ) is equivalent to V j . The scalar V j  designates the scalar magnitude of vector V j .  
      The velocity diagram in  FIG. 47  shows that the tangential component of the nozzle-jet gas particle V 1t  is diminished by an amount 2U as the particle moves through a turbine wheel and that the normal component V 1n  is unchanged. Then the number of turbine disks N D  required to null out the tangential component of the nozzle-jet, gas particle is equal to V 1t /2U. In  FIG. 47 , the number of disks required to null out the tangential component of the nozzle jet velocity is three and in  FIG. 48 , the number is one. In  FIG. 47  the energy per unit mass discharged by the nozzle is equal to V 1   2 /2, and the energy per unit mass discharged by the turbine system is V 1n   2 /2=V 6   2 /2. The power of the nozzle gases is qV 1   2 /2, where q is the mass flow rate of the nozzle gases. Note that V 1   2 =V 1t   2 +V 1n   2 . The power available to drive the turbine is qV 1t   2 /2 regardless of the turbine speed if N D  does not exceed the number of available disks. P=Tω, where T is shaft torque and ω is turbine angular velocity, relates the power P at the turbine shaft. It follows that Tω=qV 1t   2 /2=constant=C if the turbine operates with 100 percent gas dynamic efficiency. Then T=C/ω.  
       FIG. 49  gives a normalized torque, T/T O , versus normalized angular velocity, ω/ω O , curve, and a normalized power, P/P O , versus normalized angular velocity curve for a 10-disk turbine system. The normalization parameters are the angular turbine velocity ω O , at which one turbine wheel absorbs all the available nozzle jet energy, the maximum turbine power P O  and turbine stall torque T O . In the range, 0.1&lt;ω/ω O &lt;1.0, the turbine has a constant power output. The two points  1282  and  1284  indicate this range. The  FIG. 49  points  1284  and  1288  correspond to the turbine condition defined in  FIG. 48  and the point  1286  corresponds to the turbine conditions defined in  FIG. 47 . The point  1280  in  FIG. 49  corresponds to the stall-torque conditions.  
       FIGS. 50 through 57  describes a constant-power turbine. This turbine is for applications where a wide speed range is required and where a wide power range at any given speed is required. The configuration is similar to that of the Terry turbine, a small steam turbine dating back to 1906. The Terry turbine is an inefficient inexpensive impulse turbine that can use more than five stages of velocity compounding in order to accommodate a pressure drop of 25-to-1. The constant-power turbine is an impulse turbine that uses a single stage most of the time except during hard acceleration when velocity compounding multiplies torque. The pressure drop when used with a thermal compressor is typically no more than 2-to-1 and often only 1.1-to-1, and consequently a single impulse stage is used. A single stage impulse turbine can operate at an efficiency of 90% and consequently a constant-power turbine can operate as an efficient turbine.  
       FIG. 50  is a view normal to the turbine disk of a constant-power turbine and comprises a turbine disk  1304 , a stator ring  1302 , 12 nozzle feed pipes  1300 , and a turbine shaft  1306 .  FIG. 51  is a section view defined in  FIG. 50  and comprises a nozzle feed pipe  1300 , a stator ring  1302 , a turbine disk  1304 , stator return bucket  1308 , turbine bucket  1310 , and a turbine shaft  1306 .  FIG. 52  is a section view defined in  FIG. 51  and comprises three converging-diverging nozzles  1316 , three nozzle feed pipes  1300 , 9 stator buckets  1308 , a set of turbine buckets  1310 , and a turbine shaft  1306 .  FIG. 53  is a section view of a curved surface flattened as defined in  FIG. 52  and comprises a stator ring  1302 , 3 nozzle outlets  1312 , and 9 stator buckets  1308 .  FIG. 54  is a section view of a curved surface flattened as defined in  FIG. 52  and comprises a turbine disk  1304  and a set of turbine buckets  1310 .  
      A gas particle exiting a nozzle of the constant-power turbine described above will follow a trajectory characterized as a helix with the axis of the helix lying on an arc.  FIGS. 55, 56  and  57  shows the particle trajectories for three different turbine angular velocities. In these figures the solid line corresponds to that portion of the gas particle trajectory in the turbine buckets and the dashed lines corresponds to that portion of the gas particle trajectory in the stator buckets.  FIG. 55  is the gas particle trajectory for a stationary turbine,  FIG. 56  is for the case where the turbine blade velocity is one-sixth the nozzle gas velocity and  FIG. 57  is for the case where the turbine blade velocity is one-half the nozzle gas velocity. The most energy efficient condition is a single pass as described in  FIG. 57 . Note that the condition in  FIG. 57  can correspond to a high turbine velocity with a high nozzle gas velocity or to a low turbine speed with a low nozzle gas velocity. Multiple passes through the turbine of the gas particle allows the turbine torque to increase as the turbine speed decreases and thus approximate the constant-power curve shown in  FIG. 49 .  
      In  FIGS. 55, 56 , and  57  the particle enters the turbine from the nozzle at  1324  and exits at  1330 . The path through the first, second and third turbine blades is respectively designated by  1336 ,  1338  and  1340 . The callouts  1320  and  1322  respectively designated the path through the first and second stator blades.  
      The constant power turbine may be configured so it can operate in reverse, i.e., it can convert mechanical energy into compressed gas energy.  FIG. 78  describes how velocity compounding is used to increase the jet velocity of the discharge nozzle before injecting it into the intake nozzle. Thus the pressure of the gas at a point just before it enters the discharge nozzle can be somewhat less than the pressure of the gas at a point downstream of the intake nozzle. Note that in  FIG. 78 , as opposed to  FIG. 47 , the direction of the gas exiting the discharge nozzle is opposite to the direction of the turbine blade. The following discussion assumes an inviscid fluid (gas dynamic friction loses can be ignored).  
      The right side of  FIG. 78  shows three turbine blades  1252 ,  1256 , and  1260  each belonging to a different turbine wheel but connected to a common shaft; two stator blades  1254  and  1258  each belonging to a different stator wheel; a converging diverging discharge nozzle  1250 ; a converging diverging intake nozzle  1251 ; and the trajectory of a gas particle  1264  from the discharge nozzle, through the blades and then into the intake nozzle. The left side of  FIG. 78  is a velocity diagram for a gas particle moving through the blades. The turbine-blade velocity vector is designated by U, the velocity vector of a gas particle as it exits the discharge nozzle is designated by V 1 , the velocity vector of a gas particle as it exits the first turbine blade is designated by V 2 , the velocity vector of a gas particle as it exits the first stator blade is designated by V 3 , the velocity vector of a gas particle as it exits the second turbine blade is designated by V 4 , the velocity vector of a gas particle as it exits the second stator blade is designated by V 5 , and the velocity vector of a gas particle as it exits the third turbine blade and just before it enters the intake nozzle is designated by V 6 .  
      The velocity diagram in  FIG. 78  shows that the tangential component of the nozzle-jet gas particle V 1t  is increased by an amount 2U as the particle moves through a turbine wheel and that the normal component V 1n  is unchanged. Then the tangential component can increase by an additional 2U for each additional turbine disk. In  FIG. 78  the energy per unit mass exiting the discharge nozzle is equal to V 1   2 /2, and the energy per unit mass discharged by the turbine system is [(V 1t +6U) 2 +V 1n   2 ]2=V 6   2 /2.  FIG. 79  gives the associated relations for pressure and velocity assuming a perfect gas and a reversible adiabatic compression.  
      3.20 Integrated Automotive Turbine Drive  
      The  FIG. 58  schematic describes a unique drive that exploited the characteristics of the thermal compressor and incorporates functions required of an automotive drive. This drive performs the forward, reverse and retard functions.  FIG. 2 ,  FIG. 4A  or  FIG. 6  excluding the turbine  328  or  329  and generator  324 , characterizes the thermal compressor block  1378  in  FIG. 58 . It uses two constant-power turbines with variable-flow controls, one turbine  416  for forward motion with variable-flow control  412  and one turbine  418  for reverse with variable-flow control  414 . It incorporates an electromagnetic, torque-transfer system  1378  that transfers torque from a high-pressure chamber  1375  to atmospheric pressure without the need for a high-pressure seal. The turbine output rotor  1365  transfers its torque through a magnetic window  1362  to the atmospheric-pressure rotor  1367 . The atmospheric-pressure rotor  1367  drives a planetary gear train with sun gear  426 , planet gear  422 , link arm  424  and output shaft  428 . The electromagnetic, torque-transfer system uses an inner stationary exciter  1376  that excites an electromagnetic torque ring  1364  attached to the turbine-output rotor  1365 , and it uses an outer exciter  1358  which excites an electromagnetic torque ring  1360  attached to the atmospheric-pressure rotor  1367 .  
       FIG. 59  is an electrical schematic of the electromagnetic, torque-transfer system. This system can be divided into five groups, e.g., inner, stationary, exciter circuit  1376 ; inner, electromagnetic, ring circuits  1364 ; magnetic window  1362 ; outer, electromagnetic ring circuits  1360 ; and outer, exciter circuit  1358 . Both the inner and outer exciters  1376  and  1358  form two-poll stators, and the inner and outer electromagnetic ring circuits  1382  and  1392  form three-phase, alternating-current dynamos whose currents are respectively rectified by means of rectifier banks  1384  and  1390  which in turn respectively drive the coupling coils  1386  and  1388 . The inner and outer coupling coils  1386  and  1388  transfer torque across the magnetic window  1362 . This figure shows the inner and outer coupling coils  1386  and  1388  respectively embedded in iron cores  1387  and  1389 .  
       FIG. 60  shows the geometric arrangement for both coupling coils by showing a view of the coil unfolded from the surface of a cylinder and placed on a flat plane. In  FIG. 60 , the 0° line and the 360° line coincide on the cylinder. The coil has two contacts  1396  and  1398  through which current enters and exits the coil. The flow of current induces magnetic polls in the iron cores as shown in  FIG. 60 . The magnet poles of the inner electromagnetic ring will align itself with the outer electromagnetic ring so that a south pole is adjacent to a north pole across the magnetic window. This insures that there is no slippage between the inner and outer electromagnetic rings as long as the transferred torque is below some defined threshold. Increasing the current in the exciter coils increases the flux density of the magnetic field across the magnetic window and in turn, the threshold torque is increased.  
      If only one exciter is activated then one electromagnetic ring will induce hysteresis currents in the other electromagnetic ring and the system will operate as an electric clutch. Such a clutch will transfer torque with some slippage.  
      3.21 Wheel-Mounted, Automotive, Turbine-Drive System  
       FIG. 61  described an automotive wheel-mounted drive system partly characterized by the  FIG. 58  schematic.  FIG. 62  is an enlargement defined in  FIG. 61  and  FIG. 63  is a scaled down section view defined in  FIG. 61 . In addition, the turbines described in  FIG. 61  have a relationship to the one described in  FIGS. 53 through 57 .  
      In  FIG. 61 , the inner dust cover  1400  prevents dirt or water from contaminating the part of the system at atmospheric pressure; the air bearing  1401  supports the high-speed inner rotor  1402  and outer rotor  1404  that operates at atmospheric pressure, and are attached to the sun gear  1412  and the electromagnetic, torque toroid (see  FIG. 62 ); no contact magnetic seal  1408 ; wheel, inner roller bearings  1406 ; planet gear  1414 ; wheel, outer roller bearings  1416 ; grease seal  1418 ; primary support disk and ring gear  1403 ; output torque and planetary gear support structure  1415 ; disk brake rotor  1422 ; wheel lugs  1420 ; and outer bearing support structure  1419 .  
      In  FIG. 62 , the inner rotor disk and toroidal segment  1402  supports a segment of the atmospheric-pressure, electromagnetic torque toroid  1444 ; the high-pressure, toroidal nonmagnetic and dielectric container  1448  that contains all the high-pressure elements; the high-pressure, electromagnetic, torque-toroid segment  1450 ; forward-drive, stator blade mask  1452 , a device that minimizes gas-dynamic retarding forces when the forward turbine drive is off; forward turbine blades  1474 ; forward drive stator blades  1440 ; turbine-blade gas guide  1461 ; reverse-drive turbine blades  1462 ; reverse-drive, stator-blade mask  1464 ; reverse-drive stator blades  1466 ; stator-blade gas guide  1467 ; an outer, stationary exciter circuit  1460 ; an outer, electromagnetic ring circuits  1458 ; an inner, stationary exciter circuit  1470 ; an inner, electromagnetic ring circuits  1472 ; an inner, turbine gas bearing  1456 ; and a stationary, inner support ring  1468 .  
       FIG. 63  shows the disk brake rotor  1422 ; a disk brake caliper assembly  1492 ; planetary gear  1414 ; sun gear  1412 ; and ring gear  1490 .  
      4. How the Invention Is Used  
      The engine of the present invention has many applications. There are various versions of engine components. The preferred engine configuration depends on application requirements. The following are some requirements and their applicability to various versions of engine and engine component:  
      4.1 Scale—Engine Size  
      Decoupled cooler: Analyses indicate that a decoupled cooler is lower in cost and more efficient than a cooler that is not decoupled for all but the smallest engines that may use passive cooling. The decoupled cooler requires a valve set but self propels gases to the cooler, however, a cooler that is not decoupled also requires a system to remove heat.  
      Decoupled heater (ECH or ICH): Decoupling the heater requires a valve system that adds complexity. When designing a high-output heater it is desirable, with regard to cost and overall system efficiency, to be unconstrained by heater volume. The effect on design of a decoupled heater is to make the TC smaller and the heater larger. The smaller TC has lower thermal losses, lower gas dynamic losses and lower mechanical losses. A larger heater can use lower cost materials, it is easier to manufacture, can operate at a lower combustion temperature and need not have significantly higher thermal losses.  
      ICTC: The ICTC engine does not require a heater and as such is very compact.  
      Scaling law:  FIG. 67  defines a scaling law that correlates well with many physical processes including engines. In addition, this figure presents deduced relationships. This law shows trends and not absolute values since it does not correlate well with some viscous processes. Note that specific power diminishes with scale. The structural design of larger engines is typically more efficient, with regard to weight, but the underlying trend in the scale law does not change. This infers that multiple units are lighter than a single unit of equal power. Complexity and operating efficiency dictates toward fewer units.  
      4.2 Durability  
      Electromagnetic displacer drive: An electromagnetic displacer drive used with the combined small clearance seal and gas bearing operates without wear surfaces. Electrical coils have a very long service life, but are subject to fatigue failures. Thus, this drive is very good for long durability applications.  
      Center-rod-type displacer: The center-rod-type displacer with a motorized bushing and labyrinth displacer seal can operate as a near wear-free configuration after the initial wear-in period. The pressure across the center-rod oil seal is essentially zero so that maintaining continuous oil film between the seal surface and center rod is not difficult. This insures very low, center-rod, oil seal wear.  
      Crank displacer drive: At low speed, the forces acting on the displacer are small. The pressures across the pushrod and displacer seals are small and crank-bearing loads are small. Thus, this drive also is very good for high-durability, continuous-service applications; however, it is not as maintenance-free as the electromagnetic displacer drive.  
      Turbine output drive: Turbines can operate with fluid bearings having a film thickness thick enough to preclude significant wear. The invention powers turbines with warm gas at a temperature typically below 300° C. and consequently high-temperature, turbine-blade creep and fatigue failures are not a problem.  
      Direct drive: A multistage reaction turbine operating with a low-pressure ratio and warm gases lend themselves to designs that directly drive a 60 Hz generator and preclude the need for reduction gears or higher speed operation that require the use of costly power conditioning electronics. Thus, such a system is very durable.  
      4.3 Emissions  
      All versions of the engine can meet stringent emission requirements.  
      4.4 Efficiency  
      Cycle efficiency: The cycle efficiency of the various versions is ideal since they come close to the Carnot cycle efficiency (see  FIG. 64 ).  
      Thermal compressor: In a typical design, TC efficiency peaks near one-half rated output. The most important element in limiting TC efficiency is the regenerator thermal and gas dynamic losses. At low speeds, thermal losses dominate and at high speeds, gas dynamic losses dominate by requiring high displacer drive power. A system that operates at low output for long periods will preferably use multiple TCs of different sizes. Efficiency maximizes by starting and stopping TCs as required.  
      Multistage reaction turbine: Reaction turbines can operate at efficiency above 90%. The low-pressure-ratio gas outputted by the TC is ideal for driving a multistage reaction turbine operating at a low speed. The low-temperature gases allow a low-cost turbine. This approach is better suited to large direct drive systems. For constant speed systems, varying system pressurization varies output torque.  
      Single-stage, impulse turbines: Single-stage, impulse turbines can operate near 90% efficiency. The low-pressure-ratio gas outputted by the TC allows this design. This approach is better suited to small higher speed applications; however, a 0.5 m diameter impulse turbine with N 2  or air as the working fluid can directly drive a 60 Hz generator at optimal efficiency.  
      Constant-power turbine: Multi-pass impulse turbines can operate at efficiencies as low as 50%. The constant-power turbine can operate as a single-stage impulse turbine or as a multi-stage impulse turbine. A good design will operate the constant-power turbine as a single-stage impulse turbine as much as practical.  
      System efficiency:  FIG. 68  characterizes and estimates system efficiency.  
      4.5 Specific Power  
      ECH verses ICTC: For engines below 100 kW, the ECH engine is competitive with the ICTC engine in terms of specific power; however, for large engines the ICTC engine will be much lighter. In essence, this is a tradeoff between the weight of the heater and the weight of the compressor-expander module.  
      Decoupled heater: The innovative concept of decoupling the heater significantly improves ECH engine specific power and makes practical its use in large power systems.  
      Monolithic ceramic heater: The monolithic heater is much smaller and lighter then its alternative and thus improves specific power.  
      4.6 Cost  
      Pressurized ceramic structure: The innovative concept of using pressurized ceramic structures for high temperature elements allows the use of low cost ceramics.  
      60 Hz direct drive: Small gas turbines usually drive generators at high speeds and use costly, power-conditioning electronics to transform high frequency power into a 60 Hz output. The low-pressure ratio and low gas temperatures of the current invention can directly drive a 60 Hz generator with a turbine even for small engines.  
      Decoupled heater: A decoupled heater can be designed independent of heater interior volume and does not affect volumetric efficiency. This simplifies design and manufacture and consequently reduces cost.  
      Warm gas turbines: The fact that turbines experience only warm gases allows low-cost manufacture.  
      Solar power: By integrating a solar energy system into a cogeneration system the cost of solar energy is substantially lowered, i.e., the incremental cost of adding solar energy to a power system is much less then a stand alone solar system.  
      4.7 Fuel Type  
      Coal: An external combustion heater best processes a dirty fuel like coal. A coal heater is a device similar to a coal boiler for a steam power plant.  
      Distillates: An external combustion heater best processed distillates that leave a residue after combustion. A continuous internal combustion engine using a replaceable regenerator can use a fuel that leaves a residue.  
      Clean distillates and natural gas: Continuous internal combustion engines can use these fuels as can external combustion engines.  
      4.8 Output Variability  
      Variable system pressurization: One very efficient method of varying turbine output is to vary system pressurization. This method is applicable to reaction turbine, impulse turbines or positive displacement drives.  
      Variable flow nozzle: A variable flow rate nozzle is applicable to single stage impulse turbines. Such a system is applicable to either fixed or variable pressurization.  
      Constant power turbine: A constant power turbine is ideal when very high low speed torque is required for short periods.  
      Multiple TCs: The use of multiple TCs is ideal when the system operates for long period at an output that is a small fraction of rated output.  
      4.9 Start Up Time  
      ICTC: The hot chamber of the ICTC ignites almost instantaneously and for small engines achieves full power in a small fraction of a second. This engine does not have low speed combustion instabilities as does spark ignition or compression ignition engines. Thus, it operates at a very slow speed, and is stopped or started almost instantly.  
      ECH: The ECH engine requires the heater to heat up before the engine can power up. Minimizing heater mass and maximizing heater heat-diffusion rate minimizes the start-up time. Small engines will typically require almost a minute to start while very large engines may require 15 minutes or more.  
      Monolithic ceramic heater: The monolithic ceramic heater is an innovative concept that minimizes heater mass and maximizes heater heat-diffusion rate.  
      4.10 Output Responsiveness  
      Decoupled displacer drive and output drive: Since a motor independently drives the displacer, the engine can speed up or slow down very quickly.  
      Variable system pressurization: Varying pressurization varies output torque. Systems, which use variable pressurization, are very responsive. The system can quickly pressurize by opening a valve and quickly reduce output torque by slowing the TC. Depressurizing is slow but brings the system to optimal operating efficiency.  
      5. Specific Embodiments and Examples  
      Given below are examples of some specific applications for these devices.  
      5.1 Home-Cogeneration, 20 kW, TC, Turbo-Generator  
      The table in  FIG. 69  contains the specifications for a small home cogeneration unit that generates electricity and uses rejected heat to provide space heating and hot water. It uses a linear electromagnetic displacer drive, which obviates the problem of the pushrod seal. As a result, this electromagnetic drive thermal compressor (EMTC) system can operate at a high, 7.0 MPa, system pressure. In addition, it uses a monolithic ceramic heater with an interior heater volume that is small relative to its heat throughput.  
      A long, maintenance-free service life while operating continuously is an important requirement for a home cogeneration unit. Annual replacement of an air filter does not significantly influence this requirement; however, annual or regular servicing can add costs that limit the utility of the unit. The TC turbo-generator with a linear electromagnet drive can have a very long, continuous-use service life. It does not have any wear surfaces. The turbine blades do not experience high temperatures.  
      The combination of a 30 cm diameter turbine, nitrogen as the working fluid and a system pressure ratio of 1.4 allows the turbine to drive the generator directly at 60 Hz. This is desirable because it lowers the system cost and enhances system durability by eliminating the need for power-conditioning electronics.  
      To be useful as part of a home cogeneration system an engine needs to be very durable, nominally have an efficiency of more than 30% at 10% of rated output, be low cost and have very low emissions. The EMTC can meet these requirements.  
      5.2 Solar Receiver, 2 kW, TC System  
      The table in  FIG. 70  contains the specifications for a solar receiver TC system that can operate with the home cogeneration unit described above.  FIG. 10  is a schematic of the solar receiver and TC and described in Section 3.18. One or more solar TCs operate in parallel with the natural gas heated TC and all drive one turbo-generator as shown in  FIG. 9 .  
      A solar collector that uses a parabolic mirror that tracks the sun collects and concentrates the solar energy in the solar receiver. The solar receiver and TC converted the solar energy into warm gas and compressed warm gas as discussed in Section 3.7. Solar energy costs are minimized by making the solar energy system an adjunct to a natural gas, TC turbo-generator.  
      A clean parabolic mirror gives a solar energy collection efficiency of 80%. A value of 70% may be more realistic for most applications. Warm gas used to heat water (90% transfer efficiency to the water) transfers 40% of the energy; and compressed gas that is used to generate electricity (75% turbine and generator conversion efficiency) transfers 60% of the energy. Thus of the solar energy striking the parabolic mirror, 31% can be converted to electricity and 25% can be used to heat water. Therefore, the overall efficiency of the solar thermal system is approximately 56%.  
      5.3 Space Solar Thermal Power System  
      The table in  FIG. 71  contains the specifications for a space solar thermal power system. The primary objective of this system is a very long, continuous-use, maintenance-free service life. It uses an electromagnetic thermal compressor drive. Heat not required for other purposes is radiated into space using the back of the parabolic mirror as a radiating surface.  
      5.4 Continuous-Internal-Combustion, TC, 200 kW, Gas-Turbine Cogeneration System  
      The table shown in  FIG. 72  contains the specifications for a continuous internal combustion TC (ICTC) turbo-generator used as a cogeneration unit. It uses a displacer center rod with electromagnetic drive. This unit uses a labyrinth displacer seal and natural gas as its fuel.  
      The turbo generator is a variable load constant velocity (60 Hz) system. It uses a reaction turbine to drive the generator. At rated output, back work consumes 13% of fuel energy and only 5% at 30% of rated output. This system uses variable pressurization as a means of varying the turbine torque. It uses a 10 cm radius turbine, a 1.4 system pressure ratio and a four-stage reaction turbine, which directly drive a 60 Hz generator.  
      5.5 Central Power 100 MW ICTC, Base-Load Turbo-Generator  
      The table in  FIG. 73  contains the specifications for a base-load (constant output) central-power-plant, turbo-generator that uses continuous internal combustion TCs. This system is the same as the  FIG. 7A  schematic with the gas drive  329  replaced with a turbine and generator. In addition, use is made of multiple TCs operating in parallel. To achieve high efficiency axial flow reaction turbines are used that can operate at 91% efficiency and axial flow compressors that can operate at 89% efficiency. As shown in  FIG. 7A , air compression uses three stages of cooling and uses three stages of heating when the products of combustion expand back down to atmospheric pressure. This reduces back work by half. Without the cooling and heating the compression and expansion would be approximately adiabatic and with the cooling and heating the compression and expansion is approximately isothermal. When adiabatic compression is used 10% of energy performs back work, whereas, when isothermal compression is used only 5% of the energy performs back work. Back work is much less than the back work required of a gas turbine, which can require 40% of the energy. Incorporating cooling and heating as used in this system increases system energy efficiency from 58% to 61%. A modern natural-gas-fired steam turbo-generator plant can operate at an efficiency of 42% and a modern gas turbo-generator with intercooler and regenerator can operate at an efficiency of 47%. The difference for the steam plant relative to the ICTC is the advantage of the gas cycle, which for the ICTC is almost equivalent to the Carnot cycle (see  FIG. 64 ), the higher working fluid temperature, 1300° C. verses 600° C. and combustion heat utilization that for the ICTC is 10 to 20 percent better. The difference for the gas turbine plant is due to its departure from the Carnot cycle and greater amount of back work. The steam plant has the benefit of essentially no back work.  
      Durability constraints include wear, creep failure and fatigue failure. Gas turbines and steam turbines can essentially operate without wear if they use fluid (air or oil) bearings. Seal replacement is required for steam turbines. The balanced-pressure-crank displacer drive used with this ICTC system has two notable wear items, e.g., displacer pushrod seal and the displacer seal. The pushrod-balanced-pressure concept insures that the pressure across the pushrod seal is always near zero. The pressure drop across the regenerator is approximately proportional to the pressure drop across the displacer seal. With a well-designed system, this pressure drop can be limited to 0.2 MPa. This is very low when compared to diesel engines, which can operate with a pressure drop of 4.0 MPa. A thick oil film on a lubricated surface follows from a low displacer ring-seal load. Thus, slow wear occurs and a seal service life of 50,000 hours follows.  
      Above a temperature of 700° C., turbine blades can undergo creep failure. For steam turbines, this is not a problem; however, creep failure limits gas turbine service life and typically requires blade replacement before 10,000 hours. Steam boiler tubes are subject to creep failure and require servicing at intervals of 20,000 hours or less. The compressor and turbine blades for the ICTC turbo-generator experience temperatures below 170° C. and consequently creep failure is not an issue.  
      The ICTC uses a combustion temperature (1300° C.), which is low enough to preclude the formation of NO x  compounds and oxygen-rich combustion in order to minimize CO formation. A steam boiler operates oxygen-rich. In principal, they can use the same combustion temperature as the ICTC; however, in practice most gas-fueled steam boilers operate at higher temperatures in order to improve heat transfer efficiency. Gas turbines operate at much higher combustion temperatures, which result in the formation of significant amounts of NO x  compounds.  
      The ICTC is ideally suited as a base-load turbo-generator—offering advantages in efficiency, durability, emissions, cost, space and noise. The material costs to manufacture the ICTC turbo-generator are low. When multiple TCs are used, a TC can be serviced without stopping the rest of the system. The ICTC only requires a low-silhouette, small-footprint building since it does not require discharging of high temperature gases or use of a boiler that tend to be large. The absence of a high-temperature gas discharge simplifies noise reduction.  
      5.6 Auto, 270 kW, ICTC, Gas-Turbine Engine  
      The table shown in  FIG. 74  contains the specifications for a continuous internal combustion auto, 270 kW, ICTC gas-turbine engine.  FIG. 7  shows a schematic of an ICTC system that this automotive engine uses. In this schematic a three-stage, centrifugal compressor, and a three-stage, radial-flow, turbine expander is implied; however, other types of compressors and expanders are available. Both the centrifugal compressor and radial-flow turbine need to operate within a narrow flow-rate range for any given compressor-expander speed. If the flow rate is too high, it will choke and if it is to slow it will surge (be unstable). To maintain ideal flow conditions for both the compressor and the expander over the entire engine operating range, from low output to rated power, both intake and exhaust chokes are used. To control the compressor-expander subsystem, two choke motors, pressure sensors and/or flow sensors (not shown on the  FIG. 7  schematic) are required.  
      Automotive application ideally uses several ICTCs in parallel. Consider an engine with four ICTCs and a rated power of 280 kW. When driving at a steady 100 km/h the power requirements can be as little as 15 kW. One unit can efficiently deliver this power while three units are off. Each unit uses a continuously hot igniter. The units that are off can come up to full power quickly. They can have a time constant (63% of rated speed) of 0.1 second. The engine quickly achieved full power and has small thermal losses. The largest thermal loss is associated with the regenerator, which requires a large regenerator presented area and short regenerator length in order to handle a 70 kW unit. Other important ICTC unit losses include the mechanical loss and the loss related to the drop in pressure across the regenerator. It is not ideal from efficiency point of view to use one unit at rated output. At a 35 kW output, for example it is more efficient to operate two units in order to reduce the losses due to the pressure drop across the regenerator. Optimal efficiency results from the use of different size TCs.  
      Different types of drive systems are available such as a gas turbine or positive displacement motor driving a conventional transmission. The pressure in the turbine cavity can be up to 100 atmospheres and the problem of transferring power from this high-pressure region to a one-atmosphere region poses a difficult seal problem. The automotive ICTC engine will operate most of the time with a turbine cavity pressure of only two or three atmospheres and can be limited to 30 or 40 atmosphere peak pressures so that seal wear is not a serious problem, and since the working fluid, air, is continuous replenished, a small amount of leakage is not a problem.  FIG. 58  shows a drive schematic that uses two constant-power turbines and a planetary type reduction gear. This drive provides the forward, reverse and retard functions. Another type of drive is the wheel mounted drive similar to that shown in  FIG. 61 ; however, for the ICTC engine the turbine can directly drive the planetary gear system and consequently not require a magnetic field to transfer the torque from a high-pressure region to an ambient one.  
      The internal-combustion thermal compressor (ICTC) has a very short startup time since it does not use a heater.  FIG. 7B  is a schematic of an ICTC engine. Other ICTC engine system schematics are possible. This engine is for high power applications requiring a small specific weight. Emissions are not significantly different from that of an external-combustion thermal compressor engine if peak temperatures are identical. Unlike the spark ignition and diesel engines, the ICTC is a continuous combustion system that can operate at a much lower peak temperature.  
      Fuels that leave a residue after combustion or condensates that result when products of combustion cool pose a problem for the ICTC engine because the regenerator can clog up. The ICTC engine can use a replaceable or cleanable regenerator for engines that use fuels with a small amount of residue. Gaseous or clean distillates are preferred fuels for the ICTC engine.  
      The automotive ICTC engine discussed here uses a ceramic design, a displacer center rod, electromagnetic drive and uncoupled cooler. It can have much lower wear characteristics than spark ignition or diesel engines.  
      The ICTC engine achieves very low emissions without the need for a catalytic converter. A catalytic converter is costly and can degrade over time. The combustion temperature is below where NO x  compounds can form and consequently has lower emissions than spark-ignition, diesel or gas turbine engines. The use of a regenerator in the ICTC engine allows it to approach the efficiency of the Carnot cycle. This allows the ICTC engine to be more efficient than a spark ignition engine. Further efficiency advantages for multi-TC engines accrue from the ability to start and stop any one TC as discussed above.  
      There are no pulsating combustion noises like those associated with a spark ignition or compression ignition engine. The ICTC engine can be almost noise and vibration free.  
      The material costs for this engine are low and do not require many high precision components. The compressor and turbine blades only experience low temperatures. Therefore, in quantity production this engine is cost-wise competitive with current spark ignition engines.  
      5.7 Heavy Truck, 500 kW, ICTC, Gas-Drive Engine  
      The heavy truck ICTC engine described in the table in  FIG. 75  is similar to the auto engine described above except that this engine design stresses maximum efficiency, tolerates a higher weight for a given output and is less influenced by cost. In place of a turbine, a helical (Lysholm) type gas motor is used. This drive is both heavy and costly; however, it is more efficient over the power range of this application. Stall torque is high and by varying the pressurization from 0.1 MPa to 5.0 MPa, torque varies proportionately. A large number of gears common in heavy diesel trucks are not required. A three-speed transmission deals with all requirements. The engine uses four TCs. Optimal engine performance for a particular condition can requires some TCs to shut down.  
      The ICTC helical drive engine offers significant advantages over the diesel engine in the area of life-cycle costs and emissions. The cost advantages are in the area of fuel efficiency, maintenance and durability. The material costs for this engine is low and do not require high precision components with the exception of the helical drive. A heavy-duty, truck-size diesel engine will typically not exceed efficiency of 45%, whereas, the ICTC can achieve 54%. This translates into a 20% advantage in fuel mileage. Because the ICTC engine can shut down one or more TCs and optimize fuel consumption for any given driving condition, the fuel consumption advantage is more than 20%. As discussed above, emissions for the ICTC engine are very low and, in particular, do not produce particulates or NO x  compounds as diesel engines do.  
      5.8 Railroad, 5 MW, ICTC, Gas-Turbine Engine  
      An ICTC gas-turbine engine can replace the railroad diesel engine and be used to drive the generator. The table shown in  FIG. 76  describes such an ICTC engine.  
      The ICTC three-stage, reaction-turbine-drive engine offers significant advantages over the diesel engine in the area of life-cycle costs and emissions.  
      5.9 Coal Heater, 200 MW, TC Electric Power System  
      With small modifications, a modern coal fueled boiler for use with steam power plants can be used as a heater for an uncoupled TC electric power system. By varying the pressurization, the TC system can operate as a variable output power plant while maintaining a very high efficiency. The table in  FIG. 77  describes such a TC engine  
      A power system with a variable output can better meet overall power requirements.  
      6. Advantages of the Invention  
      Responsive Engine: The Invention Comprises a Gas Cycle Engine with Responsive Controls. By driving the thermal compressor displacer with a motor ( FIG. 2 ) or linear drive ( FIG. 4 ) that is independently powered, the displacer can quickly speedup or stop. Spark ignition and diesel engines are responsive; however, gas turbine and Stirling engines are not.  
      Very-Low Emission: The invention comprises a gas cycle engine with very-low emission. The engine emissions are controlled by (1) limiting the combustion temperature to a value below where significant NOx compounds are formed and below where dissociation of CO 2  occurs in large amounts, (2) requiring combustion to occur with a significant excess of oxygen, (3) maintaining products of combustion at an elevated temperature for a significant time so that combustion of all fuel elements is essentially complete, and (4) lowering the temperature of the products of combustion slowly so that dissociation of CO 2  does not result in significant residual CO in the exhaust products. Thus, the invention comprises a gas cycle engine with ideal combustion conditions resulting in very low emissions without the need for a catalytic converter.  
      Very-Low Specific Fuel Consumption: The invention comprises a gas cycle engine with very-low specific fuel consumption. The engine cycle closely approximately the Carnot cycle and can operate down to zero speed i.e., the engine can start or stop almost instantly. These features allow it to achieve very-low specific fuel consumption.  
      Excellent Volumetric Efficiency: The invention comprises a gas cycle engine with excellent volumetric efficiency. A system of valves effectively decouples the heater and/or cooler interior gas volumes. The heater and/or cooler are designed for maximum efficiency independent of interior volume. Heat first transferred from the heater to the regenerator directly compresses gas. The concept of decoupling the heater and/or cooler is an important element in achieving small, high-powered gas cycle engines.  
      Unique, Integral, Small-Clearance Displacer Seal and Gas Bearing: The invention comprises a unique, integral, small-clearance displacer seal and gas bearing. Two concentric cylinders with a small clearance between them form an integral seal and gas bearing. The significance of this innovation is that large temperature differences and large pressure fluctuations do not nullify the small clearance. This concept maintains an excellent seal without a wear surface.  
      Very Long Maintenance-Free Service Life: The invention comprises a gas cycle engine without sliding wear surfaces and consequently an engine with a very long maintenance-free service life. By using the integral small clearance seal and gas bearing, together with an electromagnetic displacer drive and spin motor, the thermal compressor can operate without sliding wear surfaces. In essence, the displacer floats in the closed containment structure without contacting the walls. This turbo-generator unit can operate with gas or liquid bearings without contact wear. A very long fatigue life designed for all components insures a very long maintenance-free service life. A gas dynamic bounce at the end of each stroke increases displacer speed and system performance.  
      Problem of a High-Pressure Pushrod Seal Circumvented: The invention comprises a gas cycle engine that uses a displacer pushrod but obviates the problem of a high-pressure pushrod seal. This problem is solved by the use of a crankcase that is completely flooded with oil and use of a bellow that causes the pressure in the crankcase and cold chamber to be equal during operation.  
      Displacer Center Rod Version: The invention comprises a gas-cycle engine version with a displacer center rod and electromagnetic drive resulting in a simple, inexpensive and durable engine. This is the simplest version of the engine. A lubricated center rod supports the displacer and a dry labyrinth displacer seal is used.  
      Low-Cost Ceramic Heat-Exchange Module: This heater uses a set of monolithic ceramic heat-exchange modules formed by sintering a stack of ceramic plates. The heater incorporates the resulting structure and operates at the highest desirable combustion temperature.  
      Low-Specific-Volume Ceramic Heater: The invention comprises a ceramic, low-specific-volume heater for mobile applications. This heater is a sintered monolithic ceramic structure comprised of a three-plate sequence which is repeated numerous times and characterized as a cloth layer, a working fluid pipe plate layer, a fuel pipe plate layer, a cloth layer, etc, together with a front plate and an aft plate.  
      Ceramics High Temperatures Elements: The invention comprises a gas cycle engine that can operate at high temperatures without the use of costly metals or costly ceramics. The invention achieves a configuration that uses low-cost ceramic parts for all parts subjected to high temperature by encapsulating them in a pressurized chamber.  
      High-Performance Regenerator: The invention comprises a high-performance, high-temperature, gas-cycle regenerator with a low-pressure drop. The regenerator uses a large regenerator area, a short regenerator length and monolithic ceramic structure encapsulated in a pressurized structure.  
      Home Cogeneration System with Ten-Year Maintenance-Free Service Life: The invention comprises an external-combustion, gas-cycle engine that is inexpensive and can achieve a ten-year maintenance-free service life for applications such as home cogeneration electric power unit. A home cogeneration unit makes economic sense only if it is maintenance-free for a long period with the exception of simple maintenance such as an annual air filter change. The invention accomplishes this by using an uncoupled TC without surfaces that wear and a turbo-generator without surfaces that wear. The engine-rejected heat provides space heating and heats water.  
      Thermal Compressor with a Solar Receiver: The invention comprises a small thermal compressor and solar receiver with a minimum ten-year maintenance-free service life. The invention uses a sun-following parabolic mirror to concentrate solar radiation and deliver it into a solar receiver that acts as a heater for a TC. This TC receives gas from a low-pressure tank, and discharges compressed gas to a high-pressure tank and low-pressure warm gas to heat water.  
      Coal-Fueled Gas Cycle Engine: The invention comprises a coal-fueled, external-combustion, gas-cycle engine that operates efficiently at rated output and down to a small fraction of rated output. With small modifications, a modern coal-fueled boiler for use with steam power plants converts to a heater for an uncoupled TC electric power system. By varying the pressurization, the TC system can operate as a variable output power plant while maintaining a very high efficiency. Thus, the power plant can operate when peak power is required and reduce its output as demand declines.  
      Internal-Combustion Gas-Cycle Engine: The invention comprises an internal-combustion, regenerative, gas-cycle engine with high performance, high efficiency and low maintenance. An ICTC gas-turbine engine differs from ECTC engines by using air as the working fluid, using continuous internal combustion in place of a heater, using a compressor to bring high-pressure air into the system and using an expander to extract energy from high-pressure exhaust gases before discharging them into the atmosphere.  
      Base-Load Central Power Plant: The invention comprises a natural gas, continuous internal-combustion TC (ICTC), central-power, base-load turbo-generator optimized for efficiency, durability and low emissions. The ICTC is ideally suited as a base-load turbo-generator—offering advantages in efficiency, durability, emissions, cost, space and noise.  
      Low Specific Weight Automobile Engines: The invention comprises a low specific weight internal-combustion, gas-cycle engine with low emissions and significantly higher efficiency than current automobile engines. The engine uses a three-stage, centrifugal compressor and a three-stage, radial-flow, turbine expander.  
      Gas-Cycle Truck Engine: The invention comprises an internal-combustion, gas-cycle truck engine with low emissions and very high efficiency. The heavy truck engine is similar to the auto engine except that this engine concept stresses maximum efficiency and not cost. In place of a turbine, a helical (Lysholm) type gas motor is used.  
      Constant Power Gas Turbine: The invention comprises a gas turbine that does not require a transmission to increase low speed torque, can operate as a constant power motor and fully exploit the thermal compressor. The thermal compressor outputs gases at temperatures that are typically below 300° C. and at pressures of 100 atmospheres, which permits small turbines manufactured from low-temperature materials to have a high power output. The constant-power turbine is an innovative device that is especially useful for low-temperature high-pressure systems. This device outputs a constant power over a broad speed range.  
      Full Function Automotive Drive Turbine: The invention comprises a full function automotive drive turbine including forward, reverse and retard functions and fully exploits the characteristics of the thermal compressor. This unique drive exploits the thermal compressor and incorporates functions required of an automotive drive.  
      Automotive, Wheel-Mounted, Turbine-Drive: The invention comprises a full-function, automotive, wheel-mounted, turbine-drive system for use with thermal compressors. This wheel hub-mounted drive-motor has many advantages including (1) a larger swivel angle than can be obtained with a constant velocity joint, (2) the elimination of constant velocity joints, (3) more road clearance for off-road vehicles and (4) elimination of differentials without a large increase in complexity, cost and unsprung suspension mass.  
      7. Alternatives and the Closing  
      Although the present invention has been described in considerable detail with reference to certain preferred versions thereof, other versions are possible. For example: 
          1. a Stirling engine with a decoupled cooler,     2. a Stirling with decoupled heater, and     3. a Stirling with decoupled cooler and heater.        

      Therefore, the spirit and scope of the appended claims should not be limited to the description of the preferred versions contained herein.  
      The reader&#39;s attention is directed to all papers and documents which are filed concurrently with this specification and which are open to public inspection with this specification, and the contents of all such papers and documents are incorporated herein by reference.  
      All the features disclosed in this specification (including any accompanying claims, abstract, and drawings) may be replaced by alternative features serving the same, equivalent or similar purpose, unless expressly stated otherwise. Thus, unless expressly stated otherwise, each feature disclosed is one example only of a generic series of equivalent or similar features.  
      Any element in a claim that does not explicitly state “means for” performing a specified function, or “step for” performing a specific function, is not to be interpreted as a “means” or “step” clause as specified in 35 U.S.C. § 112. ¶6. In particular, the use of “step of” in the claims herein is not intended to invoke the provisions of 35 U.S.C. § 112, ¶6.