Abstract:
Expansible chamber apparatus for a positive displacement, high volume, low friction, reversible rotary pump for gases and/or liquids. Two or more chambers are provided, defined in part by one or more small abutments that move radially in a slot to switch or redirect fluid flow with minimal loss of fluid mechanical energy. A complementary pair of chambers, formed by a single groove, is radially divided by a concentric land ring and is longitudinally segmented by the abutment, which seals the chambers against reverse leakage. The abutment is slightly smaller in radial dimension than the groove and is floated to avoid hard contact with the inner and outer land ring surfaces of the groove by a balancing of the Bernoulli effects that develop. The apparatus can also be operated as a fluid compressor, as a motor and in other applications, in single cycle or multiple cycle operation. The pump provides substantially non-pulsating fluid flow in one or more stages.

Description:
This application is a Continuation In Part of a patent application, U.S. Ser. No. 08/714,383, filed on Sep. 16, 1996, abandoned. The present invention relates generally to rotary pumps and more particularly to gas and liquid expansible-chamber rotary pumps, motors, compressors and expanders characterized by positive displacement, high volume, and low friction operation. 
    
    
     FIELD OF THE INVENTION 
     Description of Related Art 
     Prior art rotary positive displacement motor and pump devices have directed themselves to improving mechanical, hydraulic, and volumetric efficiencies, increasing longevity, making manufacturing easier, improving the size-to-delivery and weight-to-delivery ratios, and to including the ability to run dry without damage. Many conventional positive-displacement rotary pumps are based on a chamber with expanding and contracting volumes that provide a pumping action. Such devices usually depend on a vane, lobe or piston that is mechanically swept by a cam, crank, lever, or gear. 
     Piston pumps include connecting rods and crankshaft or a swash plate, and are heavy apparatus that require lubrication. This makes the efficiency of the apparatus low, because a heavy weight needs to be repeatedly started and stopped. In addition, the friction of the seals (piston rings, in this case) robs power, and requires lubrication. Vane pumps also have metal parts that are repeatedly started and stopped. These parts have sliding friction as well as momentum, and have a pressure-loaded mechanical or hydraulic seal. In addition, these pumps require complex porting for hydraulic loading of the seals. Gear or lobe pumps are better, as the abutment has a rotary motion. However, precision gearing is required, and there is significant friction as well. Screw type pumps suffer from the same difficulties, and are even harder to manufacture. Nutating and gyrotor pumps require complex gearing or other assemblies to give them the special motion that characterizes them. Finally, flexible impeller pumps, although they are sold in the hundreds of thousands, have serious friction losses, and they cannot be run dry without damaging the impeller. 
     The closest related art found was C. F. Davis, Jan. 1 1946, U.S. Pat. No. 2,392,029, which discloses a rotor with elliptical groove and a land ring and two disk-like cylindrical abutments. However, his device has a number of differences from the present invention. 
     First, the ports are small and of the wrong shape, so that the device will not work well as a practical pump. Second, the Davis et al invention provides only a line contact seal between the land ring and the groove in the rotor. Third, although it is asserted that the Davis et al invention will provide non-pulsating flow, it will not do so. This is because the two sides of the pump are in phase, and so their large discharges and small discharges coincide. Fourth, in the Davis patent the ports are described as being rotary valved. Thus, the ports are completely closed during part of the cycle, so the fluid is forced to stop and start, causing large hydraulic losses. 
     What is needed is a rotary pump that has little or no frictional or other energy losses, has little or no mechanical wear and that can be utilized in many different applications. 
     SUMMARY OF THE INVENTION 
     An object of the present invention is to provide a rotary pump with at least two chambers and a near frictionless abutment balanced between the chambers that does not present a load that rubs on the walls of the chambers. 
     A further object of the present invention is to provide a rotary pump with high mechanical and volumetric efficiencies. 
     Another object of the present invention is to provide a near frictionless rotary pump subject to minimum wear. 
     A further object of the present invention is to provide a reversible rotary, high-speed, positive displacement pump and motor with very little friction that can deliver relatively high flow rates at moderate to high pressures with good overall efficiency while also being able to run dry without damage. 
     A still further object of the present invention is to provide rotary motors and pumps that can easily be multiple staged in a single unit and provide non-pulsating efficient liquid flow. 
     Another object of the present invention is to provide a heat engine using Brayton, or other thermodynamic cycles, that is easy to manufacture, simple, durable, and low-cost. 
     Briefly, a rotary pump embodiment of the present invention comprises a complementary pair of expansible chambers (or lobes) radially divided by a land ring and longitudinally segmented by an abutment that seals the chambers against reverse leakage. The pair of expansible chambers is formed from a single groove in the end face of a rotor. The land ring divides the pair of chambers into complementary expansible chambers, is concentric on the end face of a stator, and extends fully into the groove. The inner and outer land rings of the groove have a constant radial separation dimension, and the abutment is urged to follow the eccentricity of the groove by positioning of the abutment in a slot in the land ring. The abutment is slightly smaller in radial dimension than the groove and is floated to avoid hard contact with the inner and outer land rings of the groove by a balancing of the Bernoulli effects that develop between the abutment and both the inner and outer land rings of the groove. Preferably, the abutment is approximately spherical and can roll in the slot radially inward and outward. 
     An advantage of the present invention is that a rotary pump is provided that is nearly frictionless. 
     Another advantage of the present invention is that a rotary pump is provided that exhibits very high mechanical and volumetric efficiencies. 
     A further advantage of the present invention is that it provides rotary motors and pumps which can easily be arranged to form multiple phased, parallel and serial stages in a single assembly, to thereby provide smoothed, high volume, and high pressure liquid flow. 
     Another advantage of the present invention is that a heat engine is provided using Brayton, and other thermodynamic cycles, that is easy to manufacture, simple, durable, and low-cost. 
     A still further advantage of the present invention is that a tolerance pump or motor is provided with no contacting surfaces. As such, the mechanical efficiency is essentially a function of the viscosity of the pumped liquid or gas. 
     Another advantage of the present invention is that embodiments can be made with only three main parts: the rotor, the stator, and the abutment. Standard milling techniques can be used to fabricate such components by casting, injection molding, sintering, or other high-volume, low cost manufacturing processes. Still further objects and advantages will become apparent from the following description and accompanying drawings. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a perspective view of a first rotary pump embodiment of the present invention. 
     FIGS. 2A-2D are cross sections of the expansible chambers of the rotary pump of FIG. 1 shown during operation at rotor phases of 0°, 90°, 180°, and 270°. 
     FIG. 3 shows an expanded view of the first pump as a double pump with two rotor discs with rotor grooves, and with the grooves 180° out of phase for non-pulsating flow. 
     FIG. 4A is a cross section of the pump through the middle of the abutment. It shows the size of porting which can be achieved by using transitions, including porting through the land ring. The pump has a flat stator plate with protruding land ring and abutment guide slot, cross hatched port area, and pumping zones for inner and outer chambers. FIG. 4B is a cross section through the rotor for FIG. 4A at the height of the middle of the abutment. It shows the rotor with rotor groove having 120° sectors of constant radius and 60° sectors of changing radius. 
     FIG. 5A is a cross section of another embodiment of the pump through the middle of the abutment. It shows the size of porting which can be achieved by omitting two of the three available cycles from the three lobe rotor, allowing a large amount of porting through the land ring. The white crescent zones adjacent to the land ring sector are the pumping zones. 
     FIG. 5B is a cross section through the rotor for FIG. 5A at the height of the middle of the abutment. It shows a three lobed (three cycle) rotor with rotor groove having 10° sectors of constant radius and 50° sectors of changing radius. 
     FIGS. 6A-6C that the land ring sector and associated ports may be rotated to chance the ratio of port area between the inlet and outlet ports. 
     FIG. 7 shows the first pump stator end plate with land ring and ports, showing a curved slot as an abutment guide. This slot is shaped so that when the apparatus is used as a pump, the pressure against the abutment causes the abutment to traverse the slot with motion similar to a pendulum. In this way, the pressure (and hence the force) component against the abutment is in the direction of required accelerating and decelerating motion. 
     FIG. 8 is a cross section of another embodiment of the pump through the middle of the abutment. It shows two abutments, with associated ports, for use with a three lobe (three cycle) rotor. One cycle is dropped to allow an enlarged port area. Because the two remaining cycles are a half cycle out of phase, this pump provides non-pulsating flow by linking the inlet ports and linking the discharge ports. 
     FIG. 8B is a cross section through the rotor for FIG. 8A at the height of the middle of the abutment. It shows a three lobed (three cycle) rotor with rotor groove having 10° sectors of constant radius and 50° sectors of changing radius. 
     FIGS. 9A-9B show a pump as in FIG. 5 where the means to prevent the abutment from rotation is a pin which protrudes from the stator, with a flexible band around the pin and around a spacer which maintains the proper width of abutment. 
     FIGS. 10A-10L is a sequence of views showing how the land ring can be made to trap fluid (cross-hatched region) between the rotor groove, the land ring, and the abutment. The trapped fluid causes the abutment to move in the desired direction due to hydraulic or pneumatic force. 
     FIG. 11A is a cross section of an embodiment of this pump as a compressor, through the middle of the abutment. In this embodiment, the discharge from the outer groove is ducted to the intake of the inner groove as shown by the dotted line. A gas moving from the outer chamber into the inner chamber is compressed. The porting is timed so that all of the ports in the embodiment are closed at the same instant, preventing puffback from one stage to another. 
     FIG. 11B is a cross section through the rotor for FIG. 11A at the height of the middle of the abutment. It shows a single lobed (one cycle) rotor with two concentric rotor grooves having 10° sectors of constant radius on the inside, 40° sectors of changing radius, and a 270° sectors of constant radius on the outside. 
     FIG. 12A is a cross section of a different embodiment of this pump, used as a compressor, through the middle of the abutment. In this embodiment, the outer chamber has no discharge port. Instead, the slot has a clearance that allows passage of gas from the outer chamber to the inner chamber. Thus, a gas is compressed from the outer chamber into the inner chamber. The inner chamber has a rotary valved discharge port for the compressed gas. 
     FIG. 12B is a cross section through the rotor for FIG. 12A at the height of the middle of the abutment. It shows a single lobed (one cycle) rotor with one rotor groove having 10° sectors of constant radius on the inside, 40° sectors of changing radius, and a 270° sectors of constant radius on the outside. 
     FIG. 13A cross section of another embodiment of this pump, through the middle of the shaft. In this embodiment, the outer chamber  76  is of shorter axial length than the inner chamber  75  in order to provide non-pulsating flow. The dotted line shows a recess for the abutment  80 . 
     FIG. 13B is a cross section through the stator for FIG. 13A at the height of the middle of the abutment. It shows a land ring  24  having a slot for the abutment  28 . The slot  80  extends into the raised area of the stator as shown by the dotted line in FIG.  13 A. The purpose of this embodiment is to provide a single rotor, single groove pump-motor which has the inner chamber swept volume equal to the outer chamber swept volume for constant flow, constant torque operation 
     FIGS. 13C and 13D are three-dimensional views of the stator and rotor respectively of the same embodiment. Note that the outer part of the rotor  82  is shorter axially than the inner part  83 , and that the inner part of the stator has a recess  81  to match the inner (longer) part of the rotor  83 .  13 C also shows the slot  80  to allow the abutment  28  to completely seal off the chambers. 
     FIG. 14A is a cross section of another embodiment of the pump through the middle of the abutment. It shows a rotating abutment that is pivoted such that the widths of the inner and outer chambers are different. The inner chamber is wider than the outer chamber, with the pivot point and rotor size chosen so that the swept volume of the inner and outer chambers are equal. The purpose of this embodiment is to provide a single rotor, single groove pump-motor that has the inner chamber swept volume equal to the outer chamber swept volume for constant flow, constant torque operation. 
     FIG. 14B is a cross section through the rotor for FIG. 14A at the height of the middle of the abutment. It shows a two lobed (three cycle) rotor with rotor groove having 10° sectors of constant radius and 80° sectors of changing or transition radius. 
     FIGS. 14C and 14D are cross sections through the rotor for FIG. 14A, showing the rotor in two different positions during a cycle. 
     FIGS. 15A through 15C show two opposing external views and a detail cross section of a heat engine embodiment of the present invention, referred to herein by the general reference numeral  200 . It has a multistage compressor and a multistage expander, referred to herein by the general reference numerals  204  and  220  respectively. 
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     FIG. 1 represents a rotary pump embodiment of the present invention, referred to herein by the general reference numeral  10 . The rotary pump  10  comprises a shaft  12  fitted with a cylindrical coaxial disc rotor  14  with an annular groove  16  and a flat end face  17  that is perpendicular to the main axis of rotation. The annular groove  16  preferably has radial width w and axial depth d that are both constant. In order to provide pumping, the whole of the annular groove  16  is not concentric with the axis of the shaft  12  or the rotor  14 . For example, the groove  16  may have a rectangular cross section. 
     A pair of complementary inner and outer expansible chambers  18  and  20  (FIGS. 2A-2D) are formed by positioning the groove  16  close to a stator  22  with a protruding annular land ring  24 . The stator  22  has a flat wall  25  that faces and is parallel to the flat wall  17  of the rotor  14 . The annular land ring  24  divides the groove  16  into an inner chamber  18  and an outer chamber  20  (FIGS.  2 A- 2 D). The land ring  24  includes a slot  26  with a slot width and a slot depth are about equal to the radial width w and the axial depth d, respectively, of the groove  16 , and the slot is fitted with an abutment  28 . In one embodiment of the present invention, the abutment  28  has a rectangular cross section and is shaped like a cylindrical disc having a diameter almost equal to the radial width of the chamber  16  and an axial length almost equal to the axial depth of the chamber  16 . The abutment  28  seals the inner and outer expansible chambers against reverse leakage and resembles a vane or piston ring, albeit without the friction created by such prior art structures. 
     The abutment  28  partially depends on the Bernoulli effect to float the abutment between the inner and outer walls of the groove  16 , and this effect reduces frictional contact to a minimum. Because the abutment  28  radially oscillates within the groove  16  and the slot  26 , the abutment is also preferably configured to have a rolling contact inside the slot  26 . See the discussion of FIG. 7 in the following. An inlet port  30  and a discharge or outlet port  32  are positioned in the stator  22  on either side of the abutment  28  and slot  26  and provide for fluid flow (liquid or gas) through the pump  10 . As the shaft  12  and rotor  14  rotate, such fluid is forcibly drawn from the inlet port  30  into both expansible chambers  18  and  20  (FIGS. 2A-2D) in opposite phases and is simultaneously pushed out through the discharge port  32 . 
     In operation, as shown in FIGS. 2A-2D, the expansible chambers  18  and  20  expand and contract in their radial dimensions only. The expansible inner chamber  18  contracts when the expansible outer chamber  20  expands, and conversely. Such phasing tends to reduce pump pressure pulsing at the outlet port  32 . The chambers  18  and  20  are bounded by the abutment  28 , and the pressures developed appear on the surface of the abutment  28 . The slot  26  is preferably formed such that fluid pumping pressure forces acting the abutment  28  are continuously perpendicular to the inside land of the slot  26  and the abutment  28  is equally free to move in either radial direction. 
     The abutment  28  is typically the only component in the pump  10  that accelerates and decelerates during normal operation. Its motion is wholly controlled by the pressure and motion of the fluid that immerses it and seeps by. Such seepage tends to balance between the inside and outside contacts with the groove  16 , due to the Bernoulli effect. Thus the abutment  28  does not ordinarily touch the inside walls or the groove  16  and this feature reduces friction and wear common to prior art pumps. 
     For a single groove  16  divided by the annular land ring  24 , the two chambers  18  and  20  are complementary, but the inner chamber  18  is usually smaller in volume than the outer chamber  20 . This imperfect matching of the chamber volumes can cause an imbalance in the pumping actions and result in a pulsation of the output flow. In alternative embodiments, it is possible to construct a pump or motor with two opposed rotors and grooves that are 180° out of phase with one another in order to eliminate pulsing. Such additional stages may be arranged in axial or radial order, or as two superimposed stepped grooves (one inside the other) requiring two abutments. 
     A stator  22  with opposed rotors allows the making of staged units wherein the output of one stage is fed directly into the input of the next stage. Each rotor  14  may have multiple cycles and also have may multiple grooves in the rotor so that staging may be done either radially or by porting across the stator face to the opposing rotor, or both. 
     FIG. 3 illustrates an alternative embodiment of the present invention, referred to herein as a dual rotary pump  50 . The pump  50  resembles the single rotary pump  10  except that it has a pair of rotors  52  and  54  and a pair of abutments  56  and  58 , compared to the single rotor  14  and abutment  28  of pump  10 . An annular groove  60 , visible in the rotor  54  in FIG. 3, is preferably 180° out of phase in its eccentricity about the axis compared to a similar annular groove (not shown) in the rotor  52 . The annular groove in the rotor  52  receives a concentric land ring  62  that has a matching concentric land ring (not shown) disposed in the groove  60 . An inlet port  64  and an outlet port  66  each open to both sides of a stator  68  and on both the radial inside and outside of the concentric land rings to provide pump flow in and out of the expansible chambers thus formed between the rotors  52  and  54  and the stator  68 . A shaft  69  turns the rotors  52  and  54  when pumping occurs. A pair of end caps  70  and  72  seal to the stator  68 . The operation of pump  50  is otherwise the same as that for the pump  10 . In an alternative embodiment, the inlet and outlet porting between the two pumping halves on either side of the stator  68  can be configured to stage the pumping halves one before the other, e.g., in series. Such would be advantageous in pumping applications where larger pumping pressures are needed and the concomitant lower pumping volumes can be tolerated. 
     As demonstrated by FIG. 3, the present invention allows diverse applications by making multiple abutments, lobes and cavities very easy and straightforward to configure. 
     A single abutment in a single divided chamber will produce radial pulses and axial thrusts on a rotor. Two rotors in opposition, e.g., 180° out of phase, can reduce the net effect of such forces, but not totally eliminate the pulsing on the bearings. Faster pulse times and higher moving rotor and shaft momentum can reduce the adverse effects on the bearings. Such pulsing can be rapid where the number of chambers greatly exceeds the number of abutments. An even number of rotors tends to balance the loads. So only configurations with odd numbers of abutments will have such undesirable radial loads, and these can be minimized by making the number of cycles larger than the number of abutments. 
     FIGS. 4A-4B represents a single rotor  14  single abutment  28  embodiment. It could be duplicated and put 180° out of phase for a double pump joined to produce pulse free flow. To do this, the respective intake and discharge ports are joined to each other. FIG. 4A shows the optimum porting for the unit. It demonstrates how the constant rotor sectors determine the maximum allowed port size for the ports  30  and  32  without communication between intake and discharge ports. The land ring  24  has one or more sectors removed to allow porting to be continuous at all phases of operation. FIG. 4B shows the rotor  14  with the rotor groove  16  having sectors of constant radius and sectors of changing or transition radius. 
     FIGS. 5A-5B illustrate an embodiment of the present invention that uses a three lobed rotor with rotor groove  16 . FIG. 5A shows stator  25  with a protruding land ring sector  24  (and land ring sector with a slot cut for the abutment). By choosing to drop two of the available three cycles in favor of porting, it allows use of larger intake and discharge ports  30  and  32 . The white areas radially inward and radially outward from the land ring  24  are the areas in which pumping action occurs between land ring  24 , stator  25 , and rotor  14 . 
     FIGS. 6A-6C show that the ratio of port area between the inlet port and discharge port may be varied in the pump described by FIGS. 5A-5B by rotating the land ring sector  24  and ports  30 ,  32 . Generally, it is useful to provide a larger port area on the low pressure side to prevent cavitation. 
     In FIG. 6A, the abutment  28  separates intake port  30  from discharge port  32 . The land ring sector  24  divides the groove  16  in the rotor  14 . 
     In FIG. 6B, the land ring sector  24  and the ports  30  and  32  have been rotated to change the relative size of the ports so they are approximately equal. 
     In FIG. 6C, the land ring sector  24  and the ports  30  and  32  have been rotated to reverse the port sizes shown in FIG. 6A, so that the discharge port  32  is now larger than the intake port  30 . 
     Referring now to FIG. 7, radial forces acting on the abutment  28  in one direction or another can be neutralized by one or more specific measures, e.g., to keep the abutment  28  from contacting the chamber walls. For example, the abutment  28  can be made of a solid material that is approximately equal in specific gravity to the fluid filling the chambers  18  and  20 . Since the abutment  28  is typically completely immersed in such fluid, choosing materials such that the specific gravities are about equal would then support the abutment to readily follow the eccentric motion of the fluid around the groove  16 . As shown in FIG. 7, the inside longitudinal limits of the slot  26  may also be curved toward the outlet port  32 , to present a series of tangential point contacts between the circular abutment  28  and the inside longitudinal limits of the slot  26  that are perpendicular for every phase of the rotation cycle. The leakage on either radial contact with the chamber walls will center the abutment  28  in the groove  16  due to the balancing of the Bernoulli effect at both places. The hydrodynamic forces generated keep the abutment  28  from contacting the surface walls. Such forces are magnified with higher fluid viscosity. 
     FIGS. 8A and 8B show a dual chamber pump that is designed for constant flow. The inlet port(s)  30  and the outlet port(s)  32  are joined (not shown) in order to achieve constant torque and non-pulsating flow with a single rotor, two abutments, and single inlet and discharge. 
     FIGS. 9A and 9B illustrate an abutment assembly  28  that consists of a flexible band  90  the axial depth of the chamber and containing a cylindrical spacer  28  to close off the rotor groove  16 . It is made to pivot on a small shaft protruding from the stator  22 . The purposes of the flexible member are: 
     to avoid high rotational bearing loads which could occur on a pivoting member; 
     to provide a wearing surface against the rotor groove walls which is continually changing its points of contact with the walls of the groove in order to minimize wear; 
     to provide better fluid wedge action against the rotor groove walls; and 
     to provide a pivoting mechanism whereby contact with either wall of the rotor groove tends to move it away from that wall rather than toward it. 
     FIGS. 10A-10L show cross sectional views of a three lobe, single abutment pump during a full 120° cycle, in 10° increments. These illustrate means for trapping fluid between the rotor groove  16 , the land ring  24  and the abutment  28 . Fluid is trapped in the volume enclosed by these three members. This trapped fluid is compressed by the rotary action and a force is exerted on the abutment  28  that causes it to move in the desired direction, away from this region of increased force. This concept is most useful in pneumatic applications. 
     FIGS. 11A and 11B show cross sectional views of an embodiment as a compressor, which has two concentric grooves.  16 A and  16 B in the rotor, and corresponding land rings  24  in the stator. The pump is ported such that the discharge port  32  of the outer groove  16 A is constantly ported into the intake port of the inner groove  16 B. This increases the area where compression occurs, in order to prevent hot spots in the rotor  14 . The ports are all closed simultaneously in order to avoid puffback. 
     FIGS. 12A and 12B show cross sectional views of an embodiment as a compressor. It has a slot for an abutment  28  that is shown as a cylindrical abutment but which can also be a sliding vane. An intake port  30  is provided in the outer chamber  20  and a discharge port  32  is provided in the inner chamber  18 . Fluid is drawn in through the intake port  30 , forced around the periphery of the land ring, passes through the slot containing the abutment into the inner chamber. Since the inner chamber is so small, it is almost immediately discharged through the discharge port  32 . Both ports are rotary valved by the rotation of the rotor, providing a simple compressor or expander. 
     FIG. 13A shows an axial cross section through the shaft of an embodiment of this mechanism designed to provide constant flow. It shows the formation of a radially inward chamber  75  and a radially outward chamber  76  where chamber  75  has a greater cross sectional area than chamber  76 . FIG. 13B shows the stator end plate  25  with the protruding land ring  24  and a slot  80  in the land ring to accommodate a sliding abutment  28 . This abutment runs the axial length of the chamber  75 , requiring that the stator be notched to accommodate the longer abutment. This notch is shown by the dotted line in FIG.  13 A. The swept volume of the inward chamber  75  is made equal to the swept volume of the outer chamber  76 , resulting in a pump-motor having constant torque and non pulsating flow with a single rotor groove and a single abutment. 
     FIGS. 14A-14D show a two lobe rator, with a variable width groove  16  and a pivoting abutment  28 . The pivot point is located radially inward from the radial center of the land ring  24 , such that it moves further radially inward from the inner surface of the land ring than it moves radially outward from the outer surface of the land ring. Because of this, the inner chamber is radially wider than the outer chamber and the swept volume of the chambers may be made equal. This results in a pump of constant torque and non-pulsating flow, using a single groove and a single abutment, but where the rotor groove  16  is not of constant width. 
     In FIGS. 15A and 15B, the heat engine  200  includes a shaft  202  that connects to a multistage air compressor  204  having a pair of rotor parts  206  and a stator part  208 . A compressed air pipe  210  connects a flow of compressed air into a regenerator  212  that recovers heat from an air flow brought in by a pipe  214 . The outlet gases are then exhausted to the atmosphere. A pipe  216  carries the compressed and heat regenerated input air to be used in a combustor  218 . The combustor is supplied with fuel from a fuel pump  230  by way of a fuel pipe  232 . The fuel is ignited by a spark plug  234  and burns continuously thereafter. Heated gases from the combustor  218  are then circulated to an expander  220  to heat a stator  222  and a pair of rotors  224 . A pipe  226  inputs the combustion gases to the expander  220  and pipe  214  carries the expanded gases away. The shaft  202  drives both rotors  206  and  224  and is used to output mechanical power. Fan blades  236  on the outer sides of the rotors can be used to maintain isothermal conditions. 
     The input air enters the outermost stage of the multistage compressor  204  and passes inward through each stage sequentially, although the interstage porting is not shown. 
     Heat can be added in the combustor  218  by either internal or external sources. The heated gases connected by pipe  226  are expanded to produce work in the expander  220 . 
     In FIG. 15C, a cross-section through the combustor  218  showing the inlet pipe  216 , the fuel pipe  232 , the spark plug  234 , and the outlet pipe  226 . 
     In various pneumatic applications, the present invention lends itself to uses such as compressors, motors, and engines. The abutments are pressurized by the fluid at the required accelerations to prevent contact with the chamber walls. In such applications, the abutment is preferably constructed of materials that have high hardness and are heat resistant. The density of the abutment material should be as low as possible. Preferably, the chamber has a maximum cavity sector on the radially outward part of the land ring and a minimum inner chamber with a fast transition between outer and inner positions. The inner chamber is ported only to the discharge port. As a compressor, this provides a configuration with a very long stroke. Depending on how efficient the tolerance seals are, a number of stages may be required. A stator with a rotor on either side allows spiraled axially inward staging of the fluid, so that the pressure can be boosted in each stage. 
     The fluid to be exhausted can exit through the axis on one side and the rotor is cantilevered so that the rotor may be attached to a shaft. On the outer periphery of the two rotors, the rotors are joined and the outer radius is in close proximity to the stator abutment for porting a rotary inlet port. An exposed rotor may be equipped with air scoops to supercharge the inlet. The rotor should be of thermal conducting material to be finned for heat rejection. 
     A liquid and gas mixture can be pumped in embodiments of the present invention so that the gases are compressed and deliver heat to the liquid. This can be done by using a multi-lobed rotor, with one of the lobes pumping a liquid and the other(s) pumping a gas, or by pumping a mixture of liquid and gas. The mixture is allowed to separate under pressure and the separated liquid and gas are passed through separate motors to recover the energy of compression. The liquid seals the pump and acts as a heat sink. The gas delivers its heat of compression to the liquid and is expanded through an expander to produce refrigerated gas, such as air, and the liquid is either discarded after passing through the motor, if the liquid is water, or the discharged liquid can be heat exchanged to ambient and recycled. 
     In a particular embodiment, a single-groove double-opposed rotor pump was built of stainless steel. Two abutments with a specific gravity  1 . 14  were constructed of nylon and had a total weight of less than one ounce. Turning the input shaft at 1750 revolutions per minute (rpm) produced twenty gallons-per-minute (gpm) at a peak pressure of 190 pounds-per-square-inch (psi). Such pump has been in intermittent service as a salt water cooling pump on a marine diesel engine on a fishing boat for more than a year. Sand, gravel, seaweed and other organic matter has been observed passing through the pump, and the pump has been run dry as long as fifteen minutes at a time. Subsequent tear down inspections show no discernible wear on the expansible chamber walls nor wear to the nylon abutment discs. This pump, nevertheless, turns so easily that light finger pressure on the input shaft can turn the rotor. A flexible vane pump had previously been in service for the same application and a pipe wrench was ordinarily needed to turn its rotor shaft on the bench. This particular flexible vane pump had always needed an impeller replacement about once a year. 
     Many alternative embodiments of the present invention are possible. A pump-motor can be configured where two or more opposed abutments and associated cycle chambers are included for radial pressure balancing and increased flow and torque. A pump-motor can be configured in which an even number of abutments and an odd number of cycle chambers provide for non-pulsating flow from one rotor. A pump-motor can be configured in which the number of cycle chambers exceeds the number of abutments and the excess cycles have a sector of the land ring removed for increased porting. A pump-motor can be configured in which the position of the land ring sector determines where the pumping action takes place and allows the size of the ports to vary, particularly to allow large intake port versus discharge as a pump and vice-versa as motor. A pump-motor can be configured in which the pumping sector is a small angular part of the whole, and that radial bearing loads are minimized since the length of pumping chamber is small. A pump-motor can be configured in which the radial bearing load is highly oscillatory, which oscillation tends to cancel when angular momentum is considered. A pump-motor can be configured in which the pumping chamber is surrounded by rotor walls except land ring and abutment. A pump-motor can be configured in which the valve action is rotary ported into the stator through the walls. A pump-motor can be configured in which two rotors are provided, joined at their outer diameters and sandwiching a stator plate with land rings, abutments and rotary ports and including check valves, and where each rotor has one or more rotor grooves which stages across the stator into the opposite rotor and back again and the fluid is caused to spiral inward gaining pressure. A pump-motor can be configured in which the compressor has air scoops on its outer diameter (rotor), which rotary valves supercharged air into the compressor and is a conducting material that is cooled by the rotary motion. A pump-motor can be configured in which the unit is used as an expander and where the check valves are omitted and the rotor is of an insulating material. A pump-motor can be configured in which the compressor and expander are linked together to form a heat engine by adding a combustion chamber or a heat exchanger to achieve a refrigeration unit. A pump-motor can be configured with a single rotor cylinder having two grooves in the rotor. The grooves would be of a stepped design (one within the other), with the first groove wider and shallower. The second groove would be narrower, and would start at the bottom of the first groove. Each groove would have its own abutment, and the grooves would be one half-cycle out of phase to minimize pulsation and balance radial loads. 
     A pump-motor can be ported as a pump combination where both a liquid and gas are compressed simultaneously and the compressing gas gives its heat to the liquid, and the liquid also serving to provide better sealing and whereby the gas and liquid under pressure are separated and the pressurized fluids are expanded, gaining back energy of compression and the spent liquid is either discharged or heat exchanged and recycled and the expanding gas provides refrigeration. 
     Although particular embodiments of the present invention have been described and illustrated, such is preferably not intended to limit the present invention. Modifications and changes will no doubt become apparent to those skilled in the art, and it is preferably intended that the present invention only be limited by the scope of the appended claims.