US3823559A

US3823559A - Heat exchanging apparatus

Info

Publication number: US3823559A
Application number: US00364680A
Authority: US
Inventors: C Foret
Original assignee: Individual
Current assignee: Individual
Priority date: 1972-12-20
Filing date: 1973-05-29
Publication date: 1974-07-16
Anticipated expiration: 1991-07-16

Abstract

A flexible vane rotor eccentrically rotates within a stator cavity to operate as a compressor and expander of a sealed working gas. The rotation of the eccentrically mounted rotor having circumferentially secured flexible vanes alteratiely compresses and expands the internally sealed working gas to cause a heat transfer to and from the respective portions of the apparatus.

Description

United States Patent [191 Foret 1 July 16, 1974 HEAT EXCHANGING APPARATUS [76] Inventor: Claude H. Foret, 10015 Culver Blvd., Culver City, Calif. 90230 [22] Filed: May 29, 1973 21 Appl. No.2 364,680

Related US. Application Data [63] Continuationin-part of Ser. No. 316,663, Dec. 20,

1972, abandoned.

' [52] US. Cl. 60/508, 62/401 [51] Int. Cl. F03g 7/06 [58] Field of Search 60/36, 59 T, 508, 509,

[56] References Cited UNITED STATES PATENTS 1,469,729 10/1923 Myers 62/401 3,141,309 7/1964 Gesell 62/401 Primary ExaminerEdgar W. Geoghegan Assistant Examiner-Allen M. Ostrager Attorney, Agent, or Firm-Spensley, Horn & Lubitz [57] ABSTRACT A flexible vane rotor eccentrically rotates within a stator cavity to operate as a compressor and expander of a sealed working gas. The rotation of the eccentrically mounted rotor having circumferentially secured flexible vanes alteratiely compresses and expands the internally sealed working gas to cause a heat transfer to and from the respective portions of the apparatus.

21 Claims, 20 Drawing Figures PATENIEBJUHBIHN 3.823.559

sum a or g 1 HEAT EXCHANGING APPARATUS ously filed application entitled Heat Exchanging Apparatus, filed Dec. 20, 1972 and designated as Ser. No. 316,633, now abandonned.

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention generally relates to heating and cooling systems and, particularly, to those systems employing a non-flow process.

2. Prior Art There are many devices disclosed in the prior art which would constitute air conditioning systems. The present invention provides a heat exchanging apparatus for simultaneously providing a source of warm air and cold air. One of the devices disclosed in the prior art which provides the dual function utilizes an axial flow compressor to deliver simultaneously two separate currents of air, one being heated, the other being chilled. This device utilizes a stator which is a laminated structure comprised of a large number of thin, discoid laminae. The stator has alternate piles of thin laminae having high heat conductivity, certain of the piles forming stator blading stages, and other piles forming spacers therebetween. Heat is removed by conduction through the stator blades and by radiation from the rotor shaft rotating within the stator to provide a source of heated air. An expansion motor is used to recover a portion of the compressed energy to reduce the temperature of a stream of cooling air. This device disclosed by the prior art is highly complex and requires structural elements which'are substantially overcome by the present invention. The device necessitates a large number of stages to be built into the compressor to provide a sufficient temperature difference between the heated compressed air and the ambient air to provide a sufficient heat transfer. The present invention substantially solves the deficiencies inherent in this device by providing a simplified structure utilizing an eccentrically disposed rotor having flexible vanes which, when rotating within a stator cavity, inherently provide for compression and expansion of a working gas in a nonflow process.

With regard to devices used for cooling ambient air, systems currently in use typically employ a two-phase refrigeration system. The components used by these devices are complicated and also expensive. As an example, the vapor-liquid compressor of a conventional system can be found in several variations, all of which must be constructed to close tolerances as well as requiring high pressure seals against contamination and refrigerant leakage. The systems also necessitate the use of two liquid-pressure-to-air heat exchangers which have similar problems. The systems disclosed in the prior art also utilize an expansion valve and a number of high pressure refrigerant lines and suitable fittings. In these prior art devices, means must also be provided on the evaporative side to conduct the air from the refrigerated space surrounding the evaporator coils to the area being cooled. The inherent deficiencies of these devices are evident. The required use of refrigerants and the equipment needed therewith is totally over- Another device disclosed by the prior art is one typically used for air refrigeration. This device utilizes an This application is a continuation-in-part of my previelliptic stator having a rotor axially disposed therein, the rotor having resiliently mounted vanes coupled radially therein. The vanes maintain contact with the stator wall bearing rotation of the rotor. Ambient air is introduced into a chamber formed by an adjacent pair of the resiliently mounted vanes, the ambient air being compressed, cooled and then expanded to recover the compression work. This device operates on the air as it passes through the device requiring that the work of compression be sufficient to compress heated air, the expansion process being adiabatic thereby resulting in the loss of the heat expansion. Although this device does not require the use of a refrigerant, there are many deficiencies in its construction. The close tolerances of the radially mounted vanes and their contact with the stator wall and rotor impose serious problems. This device uses a working cycle similar to a Brayton cycle which basically requires adiabatic processes, whereas the present invention utilizes a cycle which is similar to an Ericsson cycle which utilizes isothermal compression and expansion.

The present invention substantially solves the problems left unresolved by the devices disclosed in the prior art through the utilization of an eccentrically mounted rotor having axially mounted flexible vanes. The rotor is eccentrically disposed within a cylindrical shell, the rotation of the rotor resulting in the alternate compression and expansion of a working gas to yield the desired heating and cooling effects.

SUMMARY OF THE INVENTION The present invention comprises a heat exchanging apparatus which provides means for simultaneously producing a source of heated and cooled air. An inner shell or stator is provided, the inner shell having a uniform cylindrical bore therethrough, the outer surface having thermally isolated finned sections to facilitate heat transfer processes to and from ambient air streams passing over the respective surfaces. A rotor having a substantially cylindrical section is eccentrically mounted within the bore of the inner shell. Flexible vanes are axially mounted about the circumference of the rotor, the flexible vanes obliquely contacting the inner surface of the bore of the inner shell. Rotation of the rotor causes alternate compression and expansion of the chambers intersticial with each pair of adjacent vanes. A working gas such as helium or air is present in the volume between the inner shell and the rotor surface, the working gas being alternately compressed and expanded due to the rotation of the rotor within the shell bore.

Air streams thermally isolated from one another are directed over the surface areas associated with the respective compression and expansion of the working gas. The air stream adjacent the surface of the shell or stator which is adjacent the working gas being compressed will be heated as a result of the heat transfer from the compressed working gas to the air stream. The second air stream adjacent the surface of the stator as sociated with the expansion of the working gas will be cooled as a result of the transfer of heat from the air stream to the working gas.

The present invention operates by a non-flow process resulting from the alternate compression and expansion of the working gas sealed within the stator bore. As a result, the present invention will also operate as a heat engine where one of the air streams passing adjacent the stator shell is heated. The use of a heated air stream will result in a pressure differential causing the rotor to rotate and thereby reduce a source of power.

It is therefore an object of the present invention to provide an improved heat exchanging apparatus.

It is another object of the present invention to provide a heat exchanging apparatus utilizing a non-flow process.

It is still another object of the present invention to provide a compact device for cooling ambient air without the use of a refrigerant.

It is yet another object of the present invention to provide an improved heat exchanging apparatus which is simple and inexpensive to fabricate.

The novel features which are believed to be characteristic of the invention both as to its organization and method of operation, together with further objectives and advantages thereof, will be better understood from the following description considered in connection with the accompanying drawing in which a presently preferred embodiment of the invention is illustrated by way of example. It is to be expressly understood, however, that the drawing is for the purpose of illustration and description only and is not intended as a definition of the limits of the invention.

BRIEF DESCRIPTION OF THE DRAWING FIG. 1 is a perspective view of a heatercooling system in accordance with the present invention.

FIG. 2 is a side elevation, cross-sectional view of the present invention heat exchanging apparatus taken through line 2-2 of FIG. 1.

FIG. 3 is a perspective view of a stator shell fabricated in accordance with the present invention.

FIG. 4 is a perspective view of a rotor incorporating flexible vanes in accordance with the present invention.

FIG. 5 is a side elevation, cross-sectional view taken through line 5-5 of FIG. 2.

' FIG. 6 is a cross-sectional view of the working gas input line taken through line 6-6 of FIG. 2.

FIG. 7 illustrates an exemplary mounting of the flexible vanes to the rotor assembly of FIG. 4.

FIG. 8 is a schematic view illustrating the flexible vanes of FIG. 4 in compression and expansion.

FIG. 9a and FIG. 9b are schematic views illustrating the forces on the flexible vanes in states of compression and expansion respectively.

FIG. 10 is a graphic representation of the pressurevolume curve of the present invention heat exchanging apparatus operating in the heating/cooling mode.

FIG. 11 is a graphic representation of the pressurevolume curve of the present invention heat exchanging apparatus operating in the heat engine mode.

FIG. 12a and FIG. 12b are graphic representations of the differential pressures applied on the rotor of the present invention heat exchanging apparatus operating in the heat engine mode.

FIG. 13 is a partial, side elevation, cross-sectional view of another form of heat exchanging apparatus in accordance with the present invention.

FIG. 14 is a cross-sectional view of the rotor assembly shown in FIG. 13 taken through line 14-14 of FIG. 13.

FIG. 15 is an axial cross-sectional view of the rotor assembly shown in FIG. 13 and FIG. 14 taken through line 15-15 of FIG. 14.

FIG. 16 is a partial, top plan view of another form of a stator in accordance with the present invention.

FIG. 17 is an enlarged graphical representation of the flexible vanes coupled to the rotor assembly shown in FIG. 14.

FIG. 18 is another form of a rotor-stator assembly for a heat exchanging apparatus in accordance with the present invention.

DESCRIPTION OF THE PRESENTLY PREFERRED EMBODIMENT An understanding of a form of the present invention can be best gained by reference to FIG. 1 wherein a perspective view of an assembled heat exchanger in accordance with the present invention is shown. For the purpose of example, the configuration shown in FIG. 1 is operating as a heater/cooler unit as will be explained in detail hereinbelow. Motor 10 drives pulley 11 which is in turn coupled to endless belt 12 which supplies the rotary motion for the present invention. Belt 12 is disposed about and in frictional contact with pulley 13 which is securely mounted upon rotor shaft 14. Rotor shaft 14 is rotatably mounted within end plate 15 which, in combination with end plate 16, seals the axial confines of the present invention heat exchanging apparatus.

End plates

15 and 16 are coupled by spacers 18 in a manner which will be described in detail hereinbelow.

Ducts

19, 20, 21 and 22 extend from outer housing 17 and provide access to the interior cavity within outer housing 17. As will be explained hereinbelow,

ducts

19, 20, 21 and 22 provide the inlet and outlet passages for the streams of air which will be the recipients of the heating and cooling effects produced by the present invention heat exchanging apparatus.

Blowers

23 and 24 are shown for the purpose of example to illustrate means for providing ambient temperature sources to the present invention heat exchanging apparatus. The manner in which the air inlet means can be altered will be explained'in detail hereinbelow. The embodi: ment of the present invention heat exchanging apparatus shown in FIG. 1 is mounted upon base 25 by front and rear mounting

brackets

26 and 27 respectively. The placement of the heat exchanging apparatus as shown in FIG. 1 is for the purpose of example only, there being no attempt to limit the manner in which the assembly can be installed for operation.

An understanding of the component elements of the present invention can be best gained by reference to FIG. 2 wherein a cross-sectional view of the heat exchanging apparatus of FIG. 1 is shown. Rotor shaft 14 (FIG. 1) is eccentrically disposed within rotor shell 35 in a manner which will be described in detail hereinbelow. Rotor shell 35 is a substantially cylindrical-like member having a uniform inner and outer diameter. Referring briefly to FIG. 4, the axial ends of rotor shell 35 are partially enclosed by hub 36. In one form of the present invention, hub 36 has disposed therethrough apertures 37 which provide access to an interior plenum chamber of rotor shell 35. The position of hub 36 along the rotor shaft 14 is secured by collar clamp 38. As will be described hereinbelow, it is within the scope of the present invention to utilize other hub structures to secure rotor shell 35 to rotor shaft 14, the use of collar clamp 38 being for the purpose of illustration only.

Flexible vanes 39 are axially disposed within the outer surface of rotor shell 35 in a uniform manner about the circumference of rotor shell 35. As can be seen in FIG. 7, flexible vanes 39 can be disposed within axial slots 40 in the outer surface of rotor shell 35 after which the edge thereof can be crimped to thereby secure flexible vanes 39 within axial slot 40. Axial slots 40 are uniformly distributed about the circumference of rotor shell 35 Referring again to FIG. 2, rotor shell 35 is eccentrically mounted within stator 41. In FIG. 3, a perspective view of a simplified construction of stator 41 is best seen. Stator 41 has a cylindrical inner-bore 42 within which rotor 35 is mounted. The outer surface of stator 41 is thermally divided into two sections separated by cavities 43. In the form of the present invention shown in FIG. 3, cavities 43 are uniformly distributed along the top and bottom of stator 41 and are diametrically opposed from one another.

Walls

44 and 45 are extremely thin and thereby provide a high thermal resistance to heat transfer between the diametrically 0pposed sections separated by cavities 43.

The outer surface of stator 41 circumferentially intermediate cavities 43 has

heat dissipating fins

46 and 47 separated by spaced channels. As will be described hereinbelow, fins 47 are adjacent the compressed vanes 39 and will thereby conductively transfer the heat of compression from the working gas whereas fins 46 are adjacent vanes 39 undergoing expansion and will thereby transfer heat to the working gas undergoing expansion.

As will be described, stator 41 is to provide a conduit for the transfer of heat. To best accomplish this purpose, stator 41 is fabricated from steel or cast iron although stator 41 can be fabricated of other conventional metals such as aluminum which have good heat conductive properties. Cavities 43 separating

fins

46 and 47 provide a thermal barrier against substantial transfers between the respective air streams passing

adjacent fins

46 and 47. Walls 44 are therefore approximately one-eighth inch thick. Since stator 41 is preferably fabricated of steel or cast iron, wall 45 can be very thin, namely one thirty-second inch. In addition to providing a thermal barrier, cavities 43 also act to compensate for thermal expansion of stator 41. The semicircular portion of stator 41 including fins 46 will be adjacent the working gas being expanded whereas the semcircular portion of stator 41 including fins 47 will be adjacent the working gas being compressed. The compression and expansion processes will result in significant temperature differentials which will cause stator 41 to flex based upon the coefficient of thermal expansion of the material used to fabricate stator 41. Since steel or cast iron has a coefficient of thermal expansion of approximately one-half that of aluminum, the dimensional changes due to the temperature differential will not be as great as would be encountered if aluminum were used to fabricate stator 41. The thermal expansion of stator 41 is compensated for by the thin wall cavities 43 which will permit flexing of walls 44 while providing thermal conductivity which is sufficiently high to accomplish the object of the present invention.

Outer housing 17 is disposed about stator 41, outer housing 17 having an inner cylindrical surface

adjacent fins

46 and 47 as well as cavities 43.

Ducts

19 and 211, and 21 and 22 provide a pair of ducting channels within which air streams can be directed. In a form of the present invention, the heat exchanging apparatus operates as a heat engine. Operating in the heat engine mode, the air stream directed into duct 19 is heated by the entry of hot gasses from a manifold (not shown) having a plurality of nozzles (not shown) communicating with duct 19. Operation of the present invention heat transfer apparatus as a heat engine will be described in detail hereinbelow. Although a preferred form of the present invention utilizes

centrifugal blowers

23 and 24 to supply ambient air streams

adjacent fins

46 and 47 respectively, it is within the scope of the present invention to provide a fixed heat source and a fixed heat sink adjacent the appropriate surfaces of

fins

46 and 47.

The assembly of a form of the present invention heat transfer apparatus can be best seen by reference to FIG. 5 and FIG. 6. The form of the present invention shown in FIG. 5 and FIG. 6 utilizes a rotor assembly having an inner plenum chamber 67 which communicates with the volume intermediate the rotor assembly and stator 41. For the purpose of clarity, elements corresponding to that discussed in connection with FIGS. 1 4 will bear the same reference numerals. Rotor shaft 14 is disposed through and journeled within end plate 15, end plate 15 being supported upon mounting bracket 26. Shaft 14 is suitably journeled within end plate 15 by appropriate bearings 55. Pulley 13 is securely mounted to the end of rotor shaft 14 by a suitable set screw 56 or like coupling. Pulley 13, as shown in FIG. 1, is adapted to receive endless belt 12 for rotation of rotor shaft 14 when operating in the heater/- cooler mode. Rotor shell 35 is secured to rotor shaft 14 via hub 36 which is secured to the inner wall of rotor shell 35 by cap screws 57. As can be seen from FIGS. 4 and 5, hub 36 has apertures 37 disposed through the radial wall thereof, the radial wall of hub 36 having depending side wall 58 adjacent rotor shell 35. The top of wall 58 of hub 36 is extended axially to form a circular protrusion about the top thereof. As will be discussed hereinbelow, protrusion 59 is within a receiving cavity of end plate 15 forming a suitable labyrinth seal in which there is always some clearance but otherwise restricts the passage of working gas from the sides of the flexible vane chambers to plenum chamber 76. Rotor end play is controlled by the ends of rotor 35 bearing against respective Teflon coated faces of

end plates

15 and 16.

Collar clamp 38 is integral with the radial base of hub 36 and is securely coupled to rotor shaft 14. As mentioned, rotor shaft 14 is suitably journeled within end plate 15 so that rotor shaft 14, is vertically eccentric with respect to the center line of stator 41, the eccentricity of rotor shaft 14 with respect to stator 41 is graphically represented by the separation 60 between the designated center lines. As will be explained hereinbelow, the eccentricity of rotor shell 35 and the attached flexible vanes 39 provide for the mechanical action constituting compression and expansion of the volumetric chambers intermediate each adjacent pair of flexible vanes 39.

Stator 41 is in a cooperative relationshp with the rotor assembly as well as being coupled to

end plates

15 and 16 in a manner which will secure the total structure. Referring now to FIG. 5, stator 41 is in abuttment with end plate 15, resilient annular ring 61 being disposed between end surface 62 of stator 41 and the cooperating surface of end plate 15. Although only end plate 15 is shown, a resilient annular ring 61 is disposed intermediate the opposed surfaces of stator 41 and end plate 16 in a like manner. Resilient annular ring 61 provides for axial expansion of stator 41 as a result of the thermal properties of stator 41. To prevent stator 41 from radially turning, dowel pin 63 is axially disposed throgh end plate 15, resilient annular ring 61 and through surface 62 of stator 41. A dowel pin 63 is disposed through end plate 16 in the same manner as shown in FIG. 5. Annular channel 64 is circumferentially disposed about the inner portion of end plate 15 adjacent wall 45 of stator 41. O-ring 65 is disposed within annular channel 64 to provide a sealing surface between end plate 15 and wall 45 of stator 41. Outer housing 17 is a thermally insulated sleeve disposed about the outer surface of stator 41, the axial ends thereof being received within axially disposed grooves 66 in

end plates

15 and 16. As can be seen in FIG. 5, outer housing 17 is adjacent the respective fins 46 of stator 41. Although FIG. does not show the opposed side of stator 41, outer housing 17 is adjacent finned area 47 in a like manner. As can be seen in FIG. 5, duct 19 is coupled through outer housing 17 to form the respective air passes over the finned walls of stator 41. This can be best seen by reference to FIG. 2.

As previously mentioned, the present invention heat transfer apparatus performs a non-flow process. One of the devices disclosed in the prior art utilizes a flow process whereby radially resilient vanes coupled to a rotating rotor operate on a flowing gas to compress andexpand the gas to carry out the process. The present invention gains substantial advantages over the prior art devices by performing a non-flow process. As was previously described, the inner bore 67 of rotor shell 35 is designated the plenum chamber. In a form of the present invention, aperture 37 through hub 36 provides access to plenum chamber 67 and allows same to be filled with the working gas. As can be best seen in FIG. 2 and FIG. 6, valve 70 provides an inlet to add working gas. Valve 70 is coupled to a radially directed input line 71 which communicates with axial input line 72 disposed radially inwardly from the portion of end plate receiving protrusion 59. In this manner, gas can be input through valve 70 into the plenum chamber 67. Similarly, a valve on the other end plate 16 (not shown) allows evacuation of the vessel prior to filling. The working gas used with the present invention is a heat conductive fluid such as air, helium or even a wet vapor. Plenum chamber 67 is pressurized at or above atmospheric pressure. The volume intermediate rotor shell 35 and radial wall 45 and coextensive with the flexible vanes 39 is designated as the working chamber. Compensation line 73 is axially disposed through the interior wall of end plate 15 communicating with input line 71. Compensation line 73 provides for the input of working gas to the working chamber, the initial pressurization of the working chamber being carried out merely by rotation of rotor shaft 14. It should be noted that the O-ring seal mentioned hereinabove provides for sealing the working gas from the ambient environment but leakage will occur between plenum chamber 67 and the working chamber intermediate the flexible vanes 39 through the labyrinth seal formed by rotor protrusion 59 and end plate receiving cavity. In this manner, the quiescent condition of the rotor assembly working chamber and plenum chamber 67 since leakage will occur along the side of the flexible vanes 39 and through the labyrinth seal. Compensation line 73 also provides an important function at the time the rotor shaft is initially rotated. This will be explained in detail hereinbelow.

Before discussion, the principles of operation of the form of the present invention heat transfer apparatus shown in FIGS. 1 6, some discussion of the theory involved in the use of flexible vanes 39 is warranted. As shown in FIG. 4 and FIG. 7, flexible vanes 39 are axially inserted into the outer surface of rotor shell 35 and secured thereto. The flexible vanes 39 are uniformly distributed about the circumferential surface of rotor shell 35. Prior to insertion within stator 41, the voume defined by each pair of adjacent flexible vanes is equal. As was discussed in connection with FIG. 5, rotor shaft 14 is journeled within

end plates

15 and 16 in a manner which will provide that the rotor assembly is eccentric with respect to the interior bore defined by stator 41. As can be seen in FIG. 2, the eccentricity created by the oH-set interval 60 results in a minimum interval be tween the outer surface of rotor shell 35 and the upper wall 45 of stator 41 with a maximum interval occuring between rotor shell 35 and the bottom portion of wall 45 of stator 41. The result is that insertion of the rotor assembly within stator 41 in a manner shown in FIG. 2 and FIG. 5 results in a quiescent condition whereby the volume defined between respective pairs of adjacent flexible vanes 39 are no longer equal. The eccentricity of the rotor assembly provides a means whereby the compression and expansion processes can be unitarily carried out by the present invention. Flexible vanes 39 are typically one-half inch wide. As will be explained hereinbelow, each vane 39 is structured like a cantilever beam and will react to imposed forces in the same manner. The dimensions of the flexible vanes 39 are chosen to compromise the flexibility needed to negate the creation of undue friction while requiring the necessary strength to withstand the differential pressures created during the compression and expansion cycles. To lower the friction created and to provide good heat conductivity, the interior surface of stator 41 of a form of the present invention is typically nickel coated, with flexible vanes 39 being coated with an insulating material which will provide for dry lubrication and thermal insulation. The outer portion of rotor shell 35 is also coated with an adequate insulating material for thermal insulation as well as inside surfaces of end plates 15 and proper heating and cooling operation. In the absence of such an insulating surface, the metallic parts would act as heat sinks.

An understanding of the forces imposed on flexible vanes 39 during the rotation of the rotor assembly can be best gained by reference to FIG. 8, FIG. 9a and FIG. 9b. FIG. 8 diagramatically illustrates the maximum and minimum deflection of flexible vanes 39. The initial deflection between a tangent to surface of rotor shell 35 and the secured portion of flexible vane 39 is designated by the angle 81. The dotted projection 39a of flexible vane 39 illustrates the initial condition of flexible vane 39 in the absence of any imposed force. The minimum deflection of flexible vane 39 would occur at the position shown in FIG. 2 where the interval between the outer surface 80 of rotor shell 35 is at its maximum distance from the bottom portion of wall 45 of stator 41. In FIG. 8, this is represented by the vane reference numerals 39b. It can be seen that flexible vanes 39 are oriented at an oblique angle with respect to the contacted bore of stator 41, the minimum deflection being designated as d,. The maximum deflection will occur where the interval between the outer surface 80 of rotor shell 35 and the upper portion of wall 45 (FIG. 2) is a minimum. In FIG. 8, this is represented by the reference numerals 39c, the deflection being designated as d The pogression for minimum deflection to maximum deflection will result in the incremental reduction of the volume between adjacent flexible vanes 39 from the position 3917 to 396. As can be seen from FIG. 8, the volume intermediate the adjacent flexible vanes 39 will be incrementally reduced thereby compressing the working gas therebetween. Where the volume between adjacent vanes at minimum deflection is designated V, and where the volume at maximum deflection is designated V the compression ratio is equal to V,/V During a complete compression cycle, all chambers will pass through a series of intermediate or differential states between V, and V and therefore the flexible vanes must be of sufficient strength to resist the incremental differential pressure dP. The change in the differential pressure dP intermediate vantes 39 will be substantially logarithmic during the compression and expansion cycles. As a result of the logarithmic relationship, the differential pressure dP is quite small at the initiation of the compression cycle, the differential pressure increasing logarithmically to a large value at the termination of the compression cycle.

In a preferred form of the present invention, the compression ratio is approximately equal to 5, that is, there will be a 5:1 reduction in the volume between adjacent flexible vanes 39 during the compression cycle. This will be explained in detail hereinbelow. It is to be noted from FIG. 8 that when flexible vanes 39 are compressed to position 390 the thickness of the vanes becomes a material portion of the total volume. The crowding of the vanes will cause even greater reduction in the volume between the adjacent flexible vanes thereby causing a greater increase in the compression ratio.

The theoretical forces which are imposed upon flexible vanes 39 can be better understood by reference to FIG. 9a and FIG. 9b wherein the forces arising during compression and expansion are shown respectively. FIGS. 9a and 9b graphically represent the forces which will be directed upon flexible vanes 39 during the rotation of the rotor assembly. As mentioned previously, flexible vanes 39 act substantially like cantilever beams and therefore the imposition of forces thereon and the reaction of the vanes are analogous. Referring now to FIG. 9a, the forces incident upon vane 39 are shown diagramatically and are designated as F,, F, and F As mentioned previously, as a flexible vane 39 is increasingly subjected to compression forces, a differential pressure will be created across the flexible vane, the greater force appearing as shown in FIG. 9a. The force created by the differential pressure d? is shown to be uniformly distributed across the surface of flexible vane 39. The force F, is conventionally selected to represent a concentrated force created by the distributed differential pressure across the surface area of flexible vanes 39. The direction of the differential pressure has arbitrarily been chosen to be negative since the rotation of the rotor assembly and the flexible vanes 39 is counterclockwise. The force F, creates a negative torque about the center of rotor shell 35 which is designated by the reference numeral 92. As mentioned previously, the strength of flexible vane 39 must be sufficient to adequately resist the differential pressure represented by force F,. The alternate position of flexible vane 39 is represented by the reference numeral 39 to illustrate a safety valve effect that will be created where the differential pressure and therefore force F, exceeds the strength of flexible vane 39.

Since flexible vane 39 has finite dimensions, forces F and P, will create measurable torque about center 82. Because of the manner in which flexible vane 39 is secured within rotor shell 35, force F will constitute the particular chamber pressure imposed on surface 83. The pressure existing within a particular chamber will produce force F along the axis of flexible vane 39, the torque created thereby being dependent upon the curvature of flexible vane 39. For the purpose of simplification, the negative torque created by force F will be considered to be equal and opposite to the positive force created by force F 3 and thereby cancel the opposing effects of each.

FIG. 9b represents the forces imposed upon a flexible vane 39 during expansion. Neglecting the efforts of forces F and F during expansion, the polarity of the pressure applied to the surface of flexible vane 39 is alternated in a manner shown in FIG. 9b. As stated, the safety valve effect shown in FIG. 9a exists when the strength necessary to maintain contact between the edge of flexible vanes 39 and the bore of stator 41 is exceeded. In FIG. 9b, it is seen that the vane 39 is no longer acting as a cantilever beam but is supported at both ends thereby providing increased strength to resist the imposed differential pressure. The positive pressure creating force F, will cause flexible vanes 39 to slide along the inner bore of stator 41 with a resulting rotor rotation as depicted by the directional arrow.

The present invention heat transfer apparatus will simultaneously produce sources of heated and cooled air. Referring to FIG. 2, a first air stream passage is created by

ducts

19 and 20 and along the intermediate surfaces of fins as and the spaces intermediate fins 46. A second passage for ambient air is created by

ducts

21 and 22 and the volume intermediate the fins 47. Centrifugal blowers are typically connected to

ducts

19 and 22. It is to be noted that the blowers could be connected to

ducts

19 and 21 to provide for counterclockwise flow of the ambient air, but reverse flow will provide more efficient operation because'the rate of heat transfer depends on the pressure of the working gas and maximum heat transfer occurs near the end of the compression cycle and near the beginning of the expansion cycle. Therefore, to obtain the most efficient operation, the blowers should be coupled to

ducts

19 and 22. Where centrifugal blowers are connected to

ducts

19 and 22, the air streams will be exhausted at

ducts

20 and 21 respectively. For the purpose of definition, the air entering at duct 19 will be designated the air conditioned stream, the ambient air entering at duct 22 being designated the air cooling stream.

Under the initial conditions, the plenum chamber 67 and working chamber will be filled with a heat conductive working gas such as air or helium at or above atmospheric pressure. Under initial conditions, all working gas within the volumetric chambers intermediate the flexible vanes 39 will be at substantially the same pressure and temperature as that contained in plenum chamber 67. Motor 10 (FIG. 1) will start rotation of rotor shaft 14 in a counterclockwise direction. Referring to FIG. 2, the initial rotation of rotor shaft 14 will cause a compression cycle to commence, the compression cycle constituting the compression of the working gas contained within the volume intermediate the working vanes 39 shown on the right side of the drawing. As is well known, during the compression cycle, the heat of compression will tned to build up in the absence of means for dissipation of the heat. During the compression cycle of the form of the present invention shown in FIG. 2, heat will be transferred through the wall of stator 41 at fins 47 causing the temperature of the air cooling stream to be increased. The compression cycle is substantially an isothermal process as shown in FIG. 10.

Flg. 10 is a graphic representation of the typeical pressure-volume relationship being carried out during the operation of the form of the present invention shown in FIG. 2. The curve 90 represented by points 123 represents the isothermal compression cycle of the present invention heat transfer apparatus. Point 1 represents the initiation of the compression cycle and comprises a point where the working gas is at a temperature of approximately 500 Rankine. From point 1 to point 2, the pressure within the working chamber rises very rapidly due to compression of the working gas between adjacent flexible vanes 39. The curve 91 from point a to point b represents a pure isothermal curve at 530 Rankine (70F). The curve 92 from a to d represents the air conditioned stream, the curve 93 from b to being the air cooling stream. As can be seen from FIG. 10, the temperature of the air cooling stream at point b is 530 Rankine, the ambient temperature, the final temperature a point c being 550 Rankine. It is to be noted that to illustrate efficiency, the streams of ambient air are shown in reverse flow. In a like manner,

the initial temperature of the air conditioned stream is 530 Rankine, the final temperature being 510 Rankine.

As stated previously, the curve from point 1 to point 2 illustrates the rapid increase in pressure due to compression of flexible vanes 39 as well as an initial transfer of heat from the air cooling stream to the working gas just emerging from an expansion cycle. From point 2 to point 3, curve 90 is substantially close to isothermal compression with the temperature dropping slightly in the vicinity of point 3. During this portion of the curve, a substantially constant temperature differential is established with the air cooling stream since the heat of compression is absorbed by the air cooling stream pursuant to the heat transfer process through the wall of stator 41. As can be seen, the termination of curve 90 in the vicinity of point 3 illustrates a substantially logarithmic increase in pressure. From point 3 to point 4 on the curve 94 of FIG. 10, the working chamber pressure will drop very rapidly due to expansion of the working gas as well as a partial heat loss to the air conditioned stream.

Referring briefly to FIG. 2, the expansion cycle is initiated at the top interface between flexible vanes 39 and wall 45 of stator 41, the expansion cycle continuing during the time the flexible vanes 39 traverse the left side of stator 41. From point 4 to point 1, the curve 94 is substantially an isothermal expansion with the temperature rising slightly toward point 1. During this portion of the curve, a substantially constant temperature differential is established with the air conditioned stream and heat is transferred from the air conditioned stream to the expanding working chambers. The effect of the substantially isothermal compression/expansion cycles of FIG. 10 are that in the reverse flow, the air cooling stream will enter at point b at approximately 530 Rankine and exhaust at point e at a temperature of 550 Rankine. The air conditioned stream will enter at point a at a temperature of approximately 530 Rankine and exhaust at point d at approximately 510 Rankine. The heating effect occuring during the isothermal compression cycle comprises the net isothermal work of compression less the partial heat loss by the air cooling stream to the working chamber at the beginning of compression. The cooling effect compris ing the net isothermal work of expansion less the partial heat gain by the air conditioned stream from the working chamber at the beginning of expansion. The horsepower required from motor 10 is equal to the difference between the work of compression and the work of expansion.

Before discussing other aspects of the present invention heat transfer apparatus, reference must again be directed to FIG. 6 in connection with initiation of rotation of the rotor assembly and the equalization of pressure in the working chambers initially undergoing expansion. As will be explained hereinbelow, this discussion related only to the use of a rotor assembly having a plenum chamber 67 and hub 36 such as that shown in FIG. 5. Referring to FIG. 6, compensation line 73 is substantially at the bottom of the inner face between flexible vanes 39 and wall 45. The positioning of compensation line 73 coincides with the start of the compression cycle. Compensation line 73 is large enough to permit rapid pressure compensation of all working chambers intermediate the adjacent flexible vanes'39. At the initiation of the operation of the present invention heat transfer apparatus, all portions of the working chamber will be at substantially the same pressure. The working chambers on the left side of the rotor assembly (FIG. 2) will initially be at the base pressure and then undergo expansion which will by its nature reduce pressure below that desired. Compensation line 73 provides means whereby each working chamber will be brought back to the base pressure from plenum chamber 67 to compensation line 73.

The result of the present invention heat transfer apparatus operating as a heater/cooler unit is to produce heating and cooling effects which are readily usable for such applications as automobile heating and air conditioning units. The heating effect is approximately equal to the isothermal compression work. The cooling effect is approximately equivalent to the isothermal expansion work.

As mentioned previously, the present invention heat transfer apparatus is also capable of operating as a heat engine. A heat engine cycle is a thermodynamic cycle in which there is a net heat flow to the system and a net work flow from the system. The system which executes a heat engine cycle is a heat engine. In order to operate as a heat engine, hot gases must be introduced into the air conditioned stream to provide an ambient stream of heated air. To accomplish this, a manifold having nozzles projecting into duct 19 is provided, hot gasses being provided through the manifold and mixed with the air stream prior to blowing over the finned area 46 of stator 41. Referring to FIG. 2, where hot gasses are admitted at duct 19, the stream of air blowing across fins 46 and exhausting at duct 20 will be referred to as the hot air stream. To provide the air cooling stream, centrifugal blowers are coupled to duct 21 to provide a stream of air across fins 47 exhausting at duct 22. The air streams are assumed to be in the same direction as the rotation of rotor shaft 14, i.e., counterclockwise because it is desirable that the hot gas stream heats the working gas in its most compressed state to initiate the heat engine cycle. On the other hand, the centrifugal blowers used to generate the air cooling stream can be coupled to duct 22 to provide a reverse flow for the air cooling stream.

As stated hereinbelow, a heat engine cycle is one where there is a net heat flow to the system and a net work flow from the system. The hot air stream used in connection with the present invention heat transfer apparatus is described hereinabove. Referring to FIG. 1, the operation of the present invention heat transfer apparatus as a heater-cooler unit utilized power derived from motor 10. When operating as a heat engine, endless belt 12 can be coupled to a generator or like device for utilizing the power produced by the work output from the heat engine. The compression cycle of the present invention heat transfer apparatus operating as a heat engine is substantially the same as that described in connection with the heater/cooler application.

Referring to FIG. 2, assuming initialization of the rotation of the rotor assembly, the working gas intermediate adjacent pairs of flexible vanes 39 will be substantially logarithmically compressed with the heat of compression being dissipated at fins 47 to the air cooling stream. The working gas will be compressed by the compression ratio factor, the temperature of the working gas at the end of the compression cycle being substantially the same as the ambient temperature of the air cooling stream. In a manner which is similar to that shown by curve 90

intermediate points

2 and 3 thereof, the compression cycle of the present invention heat transfer apparatus operating as a heat engine is substantially an isothermal compression process. Upon reaching the upper interface between flexible vanes 39 and wall 45 (FIG. 2), expansion of the working gas will commence. The compressed working gas will be heated at a high rate by the hot gas stream which will result in expansion of the working gas at a substantially constant pressure over a portion of the cycle after which heat will be absorbed at a decreasing rate by the working gas by the transfer of heat from the hot gas stream via fins 46. Following the initial expansion at constant pressure, the remainder of the expansion process is similar to the isothermal expansion process illustrated by curve 94 of FIG. between points 4 and 1 thereof.

The operation of the present invention heat transfer apparatus as a heat engine has several modes of operation, the modes being dependent upon the temperature of the hot-gas stream. Where the hot gas stream is at a temperature of approximately 100 120F, there will be a heat transfer from the hot-gas stream to the working gas during the expansion cycle, the transfer being at a rate which is relatively slow, but higher than that which occurs during the heater/cooler operation. The

expansion cycle is a substantially isothermal expansion cycle and since the temperature of the hot gas stream will support expansion at constant pressure for only a short period of time. the pressure-volume expansion curve would be relatively close to the pressure-volume compression curve. Under these circumstances, neglecting effects of friction and power required for the centrifugal blowers, the expansion work could be made to equal the compression work thereby eliminating any horsepower requirements.

The use of a warm gas stream does not substantially meet the definition of a heat engine cycle since there is no net work output from the engine. Where the hot gas stream applied at duct 19 (FIG. 2) reaches a temperature in the range of 200F, a net work output can be achieved. Where the temperature of the hotgas stream has an absolute temperature in the range of 640 660 Rankine, there will be a net heat transfer from the hot gas stream to the working gas during the expansion cycle, the transfer occuring at a faster rate than that which would occur at a lower hot-gas stream temperature. The increased rate would particularly arise at the beginning of the expansion cycle due to the large temperature differential between the hot-gas stream and the working gas which would have a temperature close to the ambient temperature of the air cooling stream. As the working gas intermediate flexible vanes 39 expanded, the working gas temperature would rise and, for a short period of time, would maintain expansion at a constant pressure. When the temperature of the hot gas stream and working gas commenced to drop, the process would become polytropic as heat would be transferred from the hot gas stream to the working gas but not a rate which would produce an isothermal expansion process. Under these circumstances, the expansion work would exceed the compression work yielding a net horsepower gain from the system. A heating effect would occur at the air cooling stream outlets at duct 21 (or duct 22 where counterclockwise flow is used) since heat must be dissipated from the working chamber. A cooling effect would result in connection with the hot-gas stream, but obviously the hot gas stream would have a temperature which is in excess of the ambient air temperature at duct 20. The output from rotor shaft 14 would be capable of driving a DC generator or other energy absorbing means.

The most efficient operation of the present invention heat transfer apparatus as a heat engine would occur where the hot gas stream is in excess of that discussed hereinabove. Without the use of special materials, the present invention heat transfer apparatus is capable of operating with hot gas stream temperatures up to 500F. An understanding of the operation of the present invention utilizing a hot gas stream having a temperature of 240F (700F) can be best gained by reference to FIG. 12 wherein the pressure-volume curve of the heat engine is shown, the curve being compared to a theoretical Ericsson cycle. The Ericcson cycle is shown in dotted lines and is designated as curve 100.

The operation of the present invention heat transfer apparatus shown in FIG. 1 1 assumes the hot gas stream is operating in a parallel flow with respect to the rotor assembly, the hot gas entering at duct 19 and exhausting at duct 20. In addition, the air cooling stream is assumed to operate in a reverse flow entering at duct 22 and exiting at duct 21. Neglecting the initial effects, point 1 represents the initiation of the compression cycle, the working gas being at a temperature of 580 Rankine as compared to the temperature of the air cooling stream at point b, namely, 560 Rankine.

Curve 101 from point 1 to point 2 is substantially an isothermal compression cycle, the pressure in the working chambers intermediate adjacent pairs of flexible vanes 39 logarithmically increasing as shown. During the compression cycle, the heat of compression and that resulting at the end of the expansion cycle, are progressively removed by a heat transfer to the air cooling stream. The temperature of the working gas at the end of the compression cycle is approximately 550 Rankine which is relatively close to the ambient temperature of the air cooling stream and far below the temperature of the hot gas stream. Curve 102 represents the air cooling stream, the cross-hatched interme diate curves 101 and 102 being defined by points l2ab1 represents the heat transferred from the working gas to the air cooling stream during the compression cycle. Curve 103 represents the pressure-volume curve of the working gas during the expansion cycle. Curve 104 depicts the characteristics of the hot gas stream, the inlet temperature shown at point 0 being approximately 700 Rankine.

As the volume intermediate the adjacent pairs of flexible vanes 39 begins to expand, the pressurevolume characteristics of the working gas will exhibit expansion at a substantially constant pressure due to expansion of the volume being accompanied by a high rate of heat transfer from the hot-gas stream. This is depicted by curve 103

intermediate points

2 and 3. As stated hereinabove, the temperature of the working gas at the end of the compression cycle is approximately 550 Rankine, the temperature of the working gas at point 3 at curve 103 is approximately 620 Rankine. As the temperature of the hot gas stream decreases, the working chamber pressure during the expansion cycle will drop to approximately an isothermal expansion cycle, this being depicted by the portion of curve 103 intermediate points 3 and 1. There is a continuous heat transfer from the hot gas stream to the working chamber, the heat transfer continuing at a decreasing rate. At the terminationof the expansion cycle, as shown at point 1, the temperature of the working gas will be approximately 580 Rankine. The temperature of the hot gas stream as it exhausts at duct is approximately 600 Rankine as shown at point d of curve 104. The heat transfer from the hot-gas stream to the working gas during the expansion cycle is represented by the cross-hatched area

intermediate curves

103 and 104 and defined by points 132 cdl.

As can be seen from FIG. 11, the expansion work is substantially greater than the compression work which yields a net horsepower gain which is represented by the area

intermediate curves

101 and 103. The heating effect on the air cooling stream is greater than where the temperature of the high gas stream is lower since a greater amount of heat must be dissipated from the working chamber during the compression cycle. It is within the scope of the present invention to utilize the regenerative effect of the increased temperature of the air cooling stream and recirculate same to the hot gas stream to increase the thermal efficiency of the present invention operating as a heat engine. Based upon the discussion hereinabove, the output of rotor shaft l4 will deliver power to a DC generator or other power absorbing means.

One of the advantages of the present invention heat transfer apparatus operating as a heat engine is that the engine is self-starting and requires no direct rotational power to be supplied to rotor shaft 14 to commence rotation of the rotor assembly. The self-starting feature of the present invention can be best seen by reference to FIG. 12a and FIG. 12b. Under quiescent conditions, the base pressure of the working gas within the plenum chamber 67 and the working chamber intermediate the outer surface of rotor shell 35 and the bore of stator 41 are equal. In addition, the temperature of the working gas is at the ambient temperature of the environment. The initial conditions occur when the air cooling stream and hot gas stream are activated. FIG. 12a graphically illustrates the torque generated by the activation of the got gas stream and the air cooling stream. Since the temperature of the working gas on the compression side of the rotor assembly is substantially the same as that of the air cooling stream, the pressure of the working gas will remain equal to the base pressure. On the other hand, the activation of the hot gas stream will increase the temperature of the working gas on the expansion side of the working assembly increasing the pressure of the working gas by an amount AP. As can be seen in FIG. 12a the incease in temperature in working chamber 110 will impose a negative differential pressure (-AP), the pressure generating torque T in the direction shown about the center of the rotor assembly. By heating the working chambers on the expansion side of the rotor assembly, a positive differential pressure (+AP) will generate torque T in the direction shown about the center of the rotor assembly. There are a series of progressive temperature increments on the expansion side of the rotor assembly terminating at the maximum hot gas stream temperature at working chamber 111. On the other hand, the working chambers on the compression side of the rotor assembly are maintained substantially at the base temperature by the air cooling stream. This effect will produce the maximum starting torque depicted in FIG. 12b. The torque created by heating the working gas on the expansion side of the rotor assembly creates a positive torque T which exceeds the value of the negative torque represented by T The positive torque will tend to rotate the rotor assembly counterclockwise as required by operation of the present invention heat transfer apparatus.

The prior discussion has dealt with a theoretical heat exchanging apparatus utilizing rotary flexible vanes. Referring now to FIG. 13, another form of a heat exchanging apparatus in accordance with the present invention is shown, the heat exchanging apparatus being generally designated by the reference numeral 100. Heat exchanging apparatus utilizes a simplified rotor assembly 101 which substantially increases the efficiency of the present invention. Rotor assembly 101 comprises rotor shell 102 which is in the form of a substantially cylindrical member. Referring briefly to FIG. 15, a detailed axial cross-sectional view of rotor assembly 101 is shown. Concentrically disposed about rotor shell 102 is annular insulating member 103 which is typically fabricated of fiberglass reinforced epoxy or similar conventional insulating materials. Annular insu- Iating member 103 is secured to rotor shell 102 cy wrapping or other conventional steps of mounting used for much materials. As will be described hereinbelow, the flexible vanes utilized in the form of the present invention are lodged within the outer surface of insulating member 103.

Rotor assembly 101 is eccentrically mounted within stator 104 in a manner similar to that described previously. Referring briefly to FIG. 16, a detailed view of stator 104 can be best seen. Stator 104 is a substantially cylindrical member having a uniform cylindrical inner bore 105 (FIG. 14), rotor assembly 101 being eccentrically within inner bore 105 in a manner to be described hereinbelow. As described hereinbefore, the outer surface of stator 104 has two sets of diametrically opposed

heat dissipating fins

106 and 107, the fins being separated by

thermal barriers

108 and 109. In the embodiment of the present invention shown in FIG. 14, each of the heat dissipating surfaces of

thermal barriers

108 and 109 constitutes a single axial channel in the outer surface of stator 104 rather than the multiple thermal barriers 33 utilized in the alternate form of the present invention previously described. In addition,

thermal barriers

108 and 109 have undercut regions 110 and 111 respectively to increase the effective length of the thermal barriers. The volume defined by each of the

thermal barriers

108 and 109 is substantially filled with a suitable thermal insulating material such as fiberglass or asbestos. Stator 104 is preferably fabricated from aluminum because of the ease of workability. To insure proper cooperation between rotor assembly 101 and inner bore 105 of stator 104, inner bore 105 is typically chrome plated and coated with a dry lubricant such as molybdenum disulfide or like compounds. In addition, to increase the surface area of inner bore 105, inner bore surface 105 is subject to shot peening which will result in a lower differential temperature between the working gas and air streams and thus result in greater efficiency. It is also within the scope of the present invention to utilize a stator inner bore surface 105 which is porous which can be accomplished through the use of conventional powder metallurgy techniques.

Rotor assembly 101 is eccentrically mounted within stator 104 in a manner which is best shown in FIG. 13 and FIG. 14. End plate 112 is centered with respect to bore 105 of stator 104. Although only one end plate 112 is shown, the opposite end of stator 104 and rotor assembly 101 are sealed by a mating end plate. End plates 112 are centered with respect to stator 104 and coupled through the use of spacers 113 and bolts 114. End plate 112 includes a cylindrical inner bore 115 in off-set registration. Rotor shell 102 is suitably journeled within bore 115 through the use of needle bearings 116, the off-set registration of bore 115 providing the proper eccentricity 121 between the axis of rotor assembly 101 and stator bore 105. Tie rod 117 is coupled to rotor shell 102 through the use of supporting block 118. In the form of the present invention shown in FIG. 14, support block 118 is a triangular shaped member, each apex of support block 118 being coupled to a portion of the inner surface of rotor shell 102. Tie rod 117 is extended through bore 115 of end plate 112 and terminated at pulley wheel 1 19 and hand knob (not shown).

The working chamber used for carrying out the objectives of the present invention is intermediate the outer surface of rotor shell 102 and inner bore 105, the working chamber being generally designated by the reference numeral 122. The working chamber is filled with a suitable working gas such as helium, air or even a wet vapor. To effectively seal the working chamber from the exterior environment, stator locating insulating ring 123 has its inner diameter concentrically disposed about the annular flange 124 of end plate 112 and Iodgedin abuttment with inner bore of stator 104 and surface 125 of end plate 112. To effect the seal, annular channel 126 is disposed in the outer surface of insulating ring 123, the portion of the outer surface of insulating ring 123 in abuttment with surface 125 being chamfered. The final seal of working chamher 122 is accomplished by disposing O-ring 127 within channel 126 to provide the seal between insulating ring 123 and inner bore 105. In addition, O-ring 128 is placed in abuttment with the chamfered surface of insulating ring 123 and responds as a resilient member intermediate insulating ring 123 and end plate 112. O- ring 128 keeps stator 104 in an axially floating condition and keeps it from coming in contact with end plates 112. The air gap created by O-ring 128 is maintained for heat insulation and thermal expansion of stator 104.

To complete mounting of rotor assembly 101 within end plates 112, thrust ring 129 is disposed about the terminus of rotor shell 102 intermediate the axial surface of end plate 112 and the inner surface of pulley 119. Although only one end of the rotor assembly 101 is shown, the other axial end is mounted in a similar manner. Pin a disposed through pulley 119 into rotor shell 102 provides for mechanical coupling between pulley wheel 119 and rotor assembly 101. Radial set screws 120b are radially disposed through a portion of pulley wheel 119 to bear against the outer surface of rotor shell 102. Radial set screws 120b permit elimination of any interval between the outer diameter of rotor shell 102 and the inner diameter of pulley wheel 119 after the correct axial thrust clearance has been set.

The simplified form of rotor assembly 101 permits elimination of the labyrinth seal and plenum chamber described in connection with the form of the present invention set forth hereinabove. It can therefore be seen that the admission of working gas into working chamber 122 will provide for a non-flow operation on the working gas to carry out the objectives of the present invention. In order to sense the pressure within working chamber 122, aligned

channels

131, 132 and 133 permit communication between working chamber 122 and valve 130. The connection shown is for the disposition of a pressure gauge (not shown). Channel 133 is disposed intermediate the termination of the compression cycle and the initiation of the expansion cycle.

The form of the present invention shown in FIG. 6 utilized compensation channels to equalize the pressure between the plenum chamber 67 and the working chamber. In the preferred embodiment of the present invention, each end plate 112 includes a compensation orifice (not shown) which can be either open to the atmosphere or connected to a sealed container to provide for regulation of the base pressure. The compensation orifice is disposed through the end plates 112 and communicates with working chamber 122 intermediate the termination of the expansion cycle and the initiation of the compression cycle. The compensation orifices serve the same function as previously described.

The preferred embodiment of the present invention shown in FIGS. 13 and 14 permits a substantial reduction in the number of flexible vanes to be utilized by the present invention. Referring to FIG. 14, a plurality of channels are obliquely disposed along the axial length of annular insulating member 103 of rotor assembly 101. An enlarged view of the flexible vane assembly can be best seen by reference to FIG. 17. Each axial channel 140 is adapted to receive a flexible vane 141 along the full axial length thereof as well as wedging member 142 to secure vane 141 within channel 140. As stated previously, the isothermal compression cycle initiated by the present invention results in a substantially logarithmic increase in pressure. The logarithmic increase of pressure is matched by a vane strength which is a result of several factors. Firstly, the vane strength increases in a semi-logarithmic manner due to the progressive reduction of the effective vane width. Secondly, the compressive force acting against the vanes increases only in a semi-logarithmic manner because of the progressively reduced flexible vane surface. To match the logarithmic increase in pressure, channel 140 is formed by having its inner surface 143 rounded along a set radius 145 so that the flexing length of vane 141 is progressively reduced resulting in a semilogarithmic increase in strength. As shown in FIG. 17, points a, b, c, d, e, f and g correspond to the effective flexing length aa', bb, cc, dd, ee', ff and gg' respectively. To secure vane 141 and wedge 142 within channel 140, epoxy 144 is disposed intermediate wedge 142 and the outer surface of insulating member 103.

The ability to utilize flexible vanes 141 which increase in strength in substantially the same relationship as the pressure increases allows for the use of a minimal number of flexible vanes 141. In the preferred embodiment of the present invention, only twelve vanes 141 are needed, each vane having a thickness of approximately 0.004 inches. In addition to reducing the number of flexible vanes 141, the progressive reduction of force-reactive surfaces requires that there be less vane strength and therefore the friction against stator bore 105 is also reduced. Higher compression ratios and safer working stresses are possible due to the design of vane channel 140 which produces less dead space as the volume intermediate adjacent vanes 39 is reduced. Overstressing vanes 39 is eliminated. FIG. 17 illustrates the progressive reduction in working space intermediate rotor shell 103 and stator bore 105, the maximum space being depicted at point bb, the minimum working space being at point gg. As was described hereinabove, as rotor assembly 101 rotates counterclockwise, flexible vanes 141 will be progressively depressed compressing the working gas intermediate each adjacent pair of flexible vanes 141. When flexible vane 141 is forced into position gg, maximum compression will have been achieved.

The operation of the preferred form of the present invention can be best seen by reference to FIG. 14. Blowers (not shown) are coupled by

air ducts

150 and 151, the exhaust therefrom appearing at

ducts

152 and 153 respectively. As described, with the air being input at duct 150 and exhausting at duct 152 is designated the air conditioning stream. The air being input at duct 151 and exhausting at duct 153 being designated as the air cooling stream. The curves illustrated in FIG. substantially illustrate the substantially isothermal compression and expansion cycles which can be carried out in the form of the present invention shown in FIG. 14. As rotor assembly 101 is rotated counterclockwise, the flexible vanes 141 substantially adjacent fins 107 will be subjected to increasing forces because of the eccentricity 121. The working gas disposed intermediate pairs of vanes 141 adjacent fins 107 will be compressed in a substantially logarithimic relationship. As described hereinabove, heat will be transferred through the wall of stator 104 being dissipated at fins 107. The air exhausted at duct 153 will be heated by the transferred heat in a manner which has been described hereinabove. In a like manner, flexible vanes 141 substantially adjacent fins 106 will be subjected to decreasing forces since the eccentricity between rotor assembly 101 and stator 104 will increase the working space intermediate rotor assembly 101 and stator bore 105. The expansion of the working gas intermediate adjacent pairs of flexible vanes 141 will result in the transfer of heat from the air being input at duct to the working gas thereby causing the exhausted air at duct 152 to be at a reduced temperature.

The preferred form of the present invention heat transfer apparatus shown in FIGS. 13 17 provides a system whereby simultaneous sources of heated and cooled air can be provided in a manner which is improved over the devices disclosed in the prior art. The following constitutes an exemplary form of the present invention and the operating characteristics of same.

Free working gas volume per revolution 23 cubic inches Compressed working gas volume per revolution 4.6 cubic inches Compression ratio 5:l

Compression or expansion heat transfer bore surface Compression or expansion heat transfer finned surface Heating Capacity at 1500 RPM 5,000 Btu/hour Cooling Capacity at I500 RPM 4,500 Btu/hour Heater temperature output increase 20F Cooler temperature output decrease 20F 0.6 square feet 3 square feet Referring to FIG. 18, a schematic view of an alternative form of a rotary, flexible vane heat exchanging apparatus is shown. The forms of the present invention described hereinabove utilize a cyclindrical stator bore and an eccentrically disposed rotor. The eccentricity of the rotor assembly provides for cyclical compression and expansion of the working gas because of the compression and expansion of the flexible vanes along each half of the cylindrical bore. The form of the rotary heat exchanging apparatus shown in FIG. 18 utilizes a stator having an elliptical bore 161. The rotor 162 is centrally disposed within the elliptical bore 161 thereby providing for four alternate expansion and compression cycles, the four cycles occuring within the four quadrants of the elliptical bore 161. As an example, starting at point 163, flexible vanes 164 would be under minimum force. As rotor 162 is rotated counterclockwise, flexible vanes 164 passing from point 163 to 165 would undergo compression. As can be seen in FIG. 18,

blowers

166, 167, 168 and 169 are coupled to ducts 170,

171, l72 and 173 respectively. During the compression cycle from point 163 to 165, the air input at duct 170 and output at duct 174 will be heated as a result of the isothermal compression cycle. As flexible vanes 164 pass from point 165 to point 175, the working gas intermediate flexible vanes 164 will be expanded and the air output at duct 176 will be cooled as a result of the substantially isothermal expansion cycle. The compression cycle occuring from point 175 to 177 is substantially similar to that occurring from points 163 to 165. The expansion cycle occuring from point 177 to point 163 is substantially similar to that occuring from point 165 to point 175.

The form of the present invention heat exchanging apparatus shown in FIG. 18 eliminates problems arising out of the unbalanced forces which are imposed upon a rotor assembly which in each rotation of the rotor encounters a single compression cycle and a single expansion cycle. The unbalanced forces on the rotor will impose large forces on the bearings. The form of the present invention shown in FIG. 18 will permit use of low capacity bearings since the radial forces on the bearings will be reduced.

The present invention heat transfer apparatus provides a device which can operate as a heater/cooler or a heat engine. Extremely large heat transfer surfaces are provided with relation to the compressed volume of the working chamber. This will maintain a small temperature differential between the air streams being used and the working chamber thereby providing greater thermal efficiency. The present invention substantially utilizes isothermal compression and expansion processes which economizes on the work of compression by transferring the heat which would otherwise be developed. In addition, the transfer of heat from ambient air extends the work of expansion in a manner not achieved by the devices disclosed in the prior art.

It can therefore be seen that the present invention constitutes a flexible vane, rotary heat transfer apparatus which is substantially simpler and more economical than the devices disclosed in the prior art. The present invention can be fabricated without requiring close tolerance components and substantially resolves the difficulties encountered by the devices disclosed in the prior art. The present invention provides a self-starting apparatus to be used as a heat engine for the generation of power while also providing a heater/cooler apparatus which is usable with lower power input requirements.

The above discussion has identified preferred forms of the present invention. Although a preferred form of the present invention utilizes circumferentially finned areas along the outer surface of stator 41, it is well within the scope of the present invention to utilize longitudinally directed fins coupling same with axially disposed ducting for providing passages for the air streams used therewith. In addition, the sources of ambient air or hot gas streams can be replaced by other conventional sources for such fluids where fluids can comprise suitable liquids or gasses.

I claim:

1. A heat transfer apparatus comprising:

a. first and second fluid source means for supplying fluid at predetermined temperatures; and

b. a sealed unitary compression and expansion means for alternately compressing and expanding a heat conductive fluid, said unitary compression and expansion means including:

i. a stator having a substantially uniform inner bore and first and second heat exchanging means for transferring heat, said first and second heat exchanging means being disposed on the outer surface of said stator, said first and second heat exchanging means being thermally coupled to the fluid supplied by said first and second fluid source means respectively; and

ii. a rotor assembly rotatably coupled within said inner bore and being in off-set registration therewith, said rotor assembly including a plurality of flexible vanes uniformly distributed thereabout, each of said vanes having a portion thereof in slidable contact with said inner bore and being in oblique registration therewith.

2. A heat exchange apparatus as defined in claim 1 wherein said rotor assembly includes a rotor shaft extending exterior to said sealed unitary compression and expansion means, said heat transfer apparatus further including power source means for providing rotary power to said shaft, said power source means being coupled to said rotary shaft whereby said flexible vanes compress and expand the heat conductive fluid exchanging heat through said first and second heat exchanging means with said fluid supplied by said first and second fluid source means respectively.

3. A 'heat transfer apparatus as defined in claim 1 wherein said first fluid source means further includes means for heating said fluid to a temperature in excess of the temperature of the fluid supplied by said second fluid source means whereby a differential pressure is created across said rotor assembly causing said rotor assembly to rotate within said stator.

4. A heat transfer apparatus comprising:

a. heat source means for providing a body from which heat may be transferred;

b. heat sink means for providing a body to which heat may be transferred;

c. a heat conductive stator having a uniform cylindrical inner bore and an outer surface segmented into first and second thermally isolated portions, said first and second thermally isolated portions being thermally coupled to one of said heat source means and said heat sink means respectively;

(1. sealing means for hermetically sealing said inner bore, said sealing means coupled to the ends of said stator;

e. a rotor assembly disposed within said inner bore comprising a shaft extending through and rotatably coupled to said sealing means, said shaft being in offset registration to said inner bore, said shaft consists of a cylindrical shell, a thermally insulating member concentrically secured about said cylindrical shell and a plurality of spaced flexible vanes securely mounted circumferentially about the outer surface of said thermally insulating member, a like surface of all of said flexible vanes in slidable contact with said inner bore and being in oblique registration therewith; and

f. heat conductive fluid disposed within said inner bore filling the space intermediate said stator and said flexible vanes.

5. A heat transfer apparatus as defined in claim 4 wherein said heat source means and said heat sink 23 means comprise separable streams of air respectively.

6. A heat transfer apparatus as defined in claim including power source means for rotating said shaft whereby said flexible vanes compress and expand said heat conductive fluid exchanging heat through said first and second thermally isolated portions with said heat sink means and said heat source means respectively.

7. A heat transfer apparatus as defined in claim 5 further including means for heating said heat source air stream to a temperature in excess of the temperature of said heat sink air stream whereby a differential pressure is created across said rotor assembly causing said rotor assembly to rotate within said stator.

8. A heat transfer apparatus as defined in claim 5 wherein said sealing means comprises a pair of end plates each being coupled to an end of said stator, each of said end plates including an aperture therethrough from the ambient environment to the volume intermediate said inner bore and said rotor assembly.

9. A heat transfer apparatus as defined in claim 8 further including a compensation opening axially disposed in said end plates between said valved aperture and the sealed surface of said end plate, said compensation opening being adjacent said flexible vanes in the maximum interval between said cylindrical shell and said inner bore.

10. A heat transfer apparatus as defined in claim 5 wherein said first and second thermally isolated portions each comprise a plurality of finned sections defining channels between adjacent parts thereof, said finned sections being separated by diametrically opposed longitudinal sets of thermal barriers.

11. A heat transfer apparatus as defined in claim 10 wherein each of said finned sections are spaced axially along said stator, each of said streams of air being independent of one another and flowing through the plurality of axial channels defined by one of said finned sections respectively.

12. In a heat transfer apparatus having a heat source and a heat sink, a unitary compressor and expander comprising:

a. a heat conductive stator having a cylindrical inner bore and an outer surface segmented into a pair of heat conducting sections thermally isolated from one another, one of said heat conductive sections in thermal communication with the heat source, the other of said heat conductive sections in thermal communication with the heat sink;

b. sealing means for hermetically sealing said inner bore, said sealing means coupled to said stator;

c. a rotor assembly disposed within said inner bore comprising a shaft extending through and rotatably coupled .to said sealing means, said shaft being in off-set registration to said inner bore, said shaft consisting of a cylindrical shell, a thermally insulating member being concentrically disposed about and secured to said cylindrical shell, and a plurality of uniformly spaced flexible vanes securely mounted circumferentially about the outer surface of said thermally insulating member and extending longitudinally the length of said thermally insulating member, a like surface of all of said flexible vanes in slidable contact with said inner bore and being in oblique registration therewith; and

of end plates, each being coupled to an end of said stator, each of said end plates including an aperture therethrough from the ambient environment to the interior of said stator.

15. A unitary compressor and expander as defined in claim 14 further including a compensation opening axially disposed in said end plates between said valved aperture and the sealed surface thereof, said compensation opening being adjacent said flexible vane in the maximum interval between said cylindrical shell and said inner bore.

16. A heat transfer apparatus comprising:

a. first and second air stream sources;

b. a heat conductive stator having a uniform cylindrical inner bore and an outer surface segmented into a pair of thermally isolated portions comprising axially disposed finned sections separated by diametrically opposed longitudinal sets of thermal barriers, each of said finned sections being spaced axially along said stator, the air streams from said first and second air stream sources being in thermal communication with one of said pair of finned sections respectively whereby a separable air stream is caused to flow over each of said finned sections;

0. first and second end plates each being coupled to an end of said stator, each of said end plates including an aperture therethrough from the ambient environment to the volume enclosed by said inner bore;

d. a rotor assembly disposed within said inner bore comprising a shaft rotatably coupled to said end plate, at least one end of said shaft extending through an end plate, said shaft being in off-set registration to said inner bore, said shaft consisting of a cylindrical shell, a thermally insulating member being concentrically disposed about and secured to said cylindrical shell, and a plurality of uniformly spaced flexible vanes securely mounted circumferentially about the outer surface of said thermally insulating member and extending the axial length thereof, a like surface of all of said flexible vanes in slidable contact with said inner bore and being in oblique registration therewith; and

heat conductive fluid disposed within said inner bore filling the spaced intermediate said stator and said flexible vanes.

17. A, heat transfer apparatus as defined in claim 16 including rotary power source means for rotating said shaft, said rotary power source means being coupled to said shaft whereby said flexible vanes compress and ex pand said heat conductive fluid exchanging heat through said pair of finned sections with the air streams supplied by said first and second stream sources respectively.

18. A heat transfer apparatus as defined in claim 16 further including means for heating said air streams in thermal communication with the expanding heat conductive fluid to a temperature in excess of the temperature of the other of said air streams whereby a differential pressure is created across said rotor assembly causing said rotor assembly to rotate within said stator.

19. A heat transfer apparatus as defined in claim 16 further including a compensation opening axially disposed in both end plates between said aperture and the sealed surface thereof, said compensation opening being adjacent said flexible vanes in the maximum interval between said rotor assembly and said inner bore.

20. A heat transfer apparatus as defined in claim 16 wherein each of said thermal barriers comprises a botbarriers.

Claims

1. A heat transfer apparatus comprising: a. first and second fluid source means for supplying fluid at predetermined temperatures; and b. a sealed unitary compression and expansion means for alternately compressing and expanding a heat conductive fluid, said unitary compression and expansion means including: i. a stator having a substantially uniform inner bore and first and second heat exchanging means for transferring heat, said first and second heat exchanging means being disposed on the outer surface of said stator, said first and second heat exchanging means being thermally coupled to the fluid supplied by said first and second fluid source means respectively; and ii. a rotor assembly rotatably coupled within said inner bore and being in off-set registration therewith, said rotor assembly including a plurality of flexible vanes uniformly distributed thereabout, each of said vanes having a portion thereof in slidable contact with said inner bore and being in oblique registration therewith.

3. A heat transfer apparatus as defined in claim 1 wherein said first fluid source means further includes means for heating said fluid to a temperature in excess of the temperature of the fluid supplied by said second fluid source means whereby a differential pressure is created across said rotor assembly causing said rotor assembly to rotate within said stator.

4. A heat transfer apparatus comprising: a. heat source means for providing a body from which heat may be transferred; b. heat sink means for providing a body to which heat may be transferred; c. a heat conductive stator having a uniform cylindrical inner bore and an outer surface segmented into first and second thermally isolated portions, said first and second thermally isolated portions being thermally coupled to one of said heat source means and said heat sink means respectively; d. sealing means for hermetically sealing said inner bore, said sealing means coupled to the ends of said stator; e. a rotor assembly disposed within said inner bore comprising a shaft extending through and rotatably coupled to said sealing means, said shaft being in offset registration to said inner bore, said shaft consists of a cylindrical shell, a thermally insulating member concentrically secured about said cylindrical shell and a plurality of spaced flexible vanes securely mounted circumferentially about the outer surface of said thermally insulating member, a like surface of all of said flexible vanes in slidable contact with said inner bore and being in oblique registration therewith; and f. heat conductive fluid disposed within said inner bore filling the space intermediate said stator and said flexible vanes.

5. A heat transfer apparatus as defined in claim 4 wherein said heat source means and said heat sink means comprise separable streams of air respectively.

6. A heat transfer apparatus as defined in claim 5 including power source means for rotating said shaft whereby said flexible vanes compress and expand said heat conductive fluid exchanging heat through said first and second thermally isolated portions with said heat sink means and said heat source means respectively.

12. In a heat transfer apparatus having a heat source and a heat sink, a unitary compressor and expander comprising: a. a heat conductive stator having a cylindrical inner bore and an outer surface segmented into a pair of heat conducting sections thermally isolated from one another, one of said heat conductive sections in thermal communication with the heat source, the other of said heat conductive sections in thermal communication with the heat sink; b. sealing means for hermetically sealing said inner bore, said sealing means coupled to said stator; c. a rotor assembly disposed within said inner bore comprising a shaft extending through and rotatably coupled to said sealing means, said shaft being in off-set registration to said inner bore, said shaft consisting of a cylindrical shell, a thermally insulating member being concentrically disposed about and secured to said cylindrical shell, and a plurality of uniformly spaced flexible vanes securely mounted circumferentially about the outer surface of said thermally insulating member and extending longitudinally the length of said thermally insulating member, a like surface of all of said flexible vanes in slidable contact with said inner bore and being in oblique registration therewith; and d. heat conductive fluid disposed within said inner bore filling the spaces intermediate said stator and said flexible vanes.

13. A unitary compresser and expander as defined in claim 12 including power source means for rotating said shaft whereby said flexible vanes compress and expand said heat conductive fluid exchanging heat through said pair of heat conducting sections with said heat sink and heat source respectively.

14. A unitary compressor and expander as defined in claim 12 wherein said sealing means comprises a pair of end plates, each being coupled to an end of said stator, each of said end plates including an aperture therethrough from the ambient environment to the interior of said stator.

16. A heat transfer apparatus comprising: a. first and second air stream sources; b. a heat conductive stator having a uniform cylindrical inner bore and an outer surface segmented into a pair of thermally isolated portions comprising axially disposed finned sections separated by diametrically opposed longitudinal sets of thermal barriers, each of said finned sections being spaced axially along said stator, the air streams from said first and second air stream sources being in thermal communication with one of said pair of finned sections respectively whereby a separable air stream is caused to flow over each of said finned sections; c. first and second end plates each being coupled to an end of said stator, each of said end plates including an aperture therethrough from the ambient environment to the volume enclosed by said inner bore; d. a rotor assembly disposed within said inner bore comprising a shaft rotatably coupled to said end plate, at least one end of said shaft extending through an end plate, said shaft being in off-set registration to said inner bore, said shaft consisting of a cylindrical shell, a thermally insulating member being concentrically disposed about and secured to said cylindrical shell, and a plurality of uniformly spaced flexible vanes securely mounted circumferentially about the outer surface of said thermally insulating member and extending the axial length thereof, a like suRface of all of said flexible vanes in slidable contact with said inner bore and being in oblique registration therewith; and e. heat conductive fluid disposed within said inner bore filling the spaced intermediate said stator and said flexible vanes.

17. A heat transfer apparatus as defined in claim 16 including rotary power source means for rotating said shaft, said rotary power source means being coupled to said shaft whereby said flexible vanes compress and expand said heat conductive fluid exchanging heat through said pair of finned sections with the air streams supplied by said first and second stream sources respectively.

20. A heat transfer apparatus as defined in claim 16 wherein each of said thermal barriers comprises a bottom surface substantially aligned with the inner bore of said stator and a pair of outwardly directed, parallel, side surfaces longitudinally disposed along said stator surface, said bottom and side surfaces defining an open channel, the interface between said side surfaces and said bottom surfaces being undercut whereby the width of said bottom surface is greater than the distance between said parallel side surfaces.

21. A heat transfer apparatus as defined in claim 20 wherein thermally insulating means for retarding the transfer of heat is disposed within each of said thermal barriers.