|Publication number||US3853434 A|
|Publication date||Dec 10, 1974|
|Filing date||May 17, 1973|
|Priority date||Apr 7, 1971|
|Publication number||US 3853434 A, US 3853434A, US-A-3853434, US3853434 A, US3853434A|
|Original Assignee||F Parsons|
|Export Citation||BiBTeX, EndNote, RefMan|
|Patent Citations (3), Referenced by (4), Classifications (11)|
|External Links: USPTO, USPTO Assignment, Espacenet|
United States Patent [1 1 Parsons [111 3,853,434 [4 1 Dec. 10,1974
1 1 POSITIVE DISPLACEMENT ROTARY MACHINE  Inventor: Frederick L. Parsons, 759
Morningside Rd., Ridgewood, NJ. 07067 22 Filed: May 17, 1973 21 App1.No.:361,254
Related US. Application Data  Division of Ser. No. 132,140, April 7, 1971, Pat. No.
 US. Cl 418/55, 418/56, 418/60, 418/220  Int. Cl. F01c 1/02  Field of Search 4l8/55, 56, 60, 220
 References Cited UNITED STATES PATENTS 2,764,101 9/1956 Rand 418/56 3,238,882 3/1966 Ryffel 418/55 FOREIGN PATENTS OR APPLICATIONS I 1,181,550 11/1964 Germany 418/55 Primary Examiner-C. J. l-lusar Attorney, Agent, or Firm-Brumbaugh, Graves, Donohue & Raymond  ABSTRACT A rotary' engine for generating mechanical power from energy in a fluid stream includes a positivedisplacement air compressor, a combustion chamber and a multiple-stage positive-displacement power extractor or expander, all arranged about a common drive shaft. Eccentrically mounted inner and outer cylindrical members in the expander form a crescentshaped expansion chamber which is divided by circumferentially spaced rocker-slipper vanes into a plurality of expansion stages. Combustion gases entering the expansion stages are caused to expand upon movement thereof toward the chamber exhaust, and thus to deliver energy to one of the inner or outer members and thence to the drive shaft. A direct connection with the compressor provides high-pressure air flows within the expander for cooling and for sealing the expansion stages against leakage of combustion gases or lubricants, while reducing wear, partly by balancing the pressure forces exerted on the slippers by the combustion gases so as to urge the slippers into light sealing contact with the opposed member. The inner and outer members are arranged either to nutate one to the other or to rotate together at nearly the same speed, yielding significantly reduced relative sliding velocities therebetween to further reduce wear. Similar low-wear conditions are provided in the compressor, where eccentrically mounted double screw vanes,
of a general V-shape in cross section, are caused to nutate within correspondingly shaped cavities in the housing so as to displace air at increasing pressure spirally towards a central outlet from suction inlets at each end. The vanes preferably are composed of flexible, nonmetallic material to simplify sealing and to enhance heat-transfer characteristics. Variable lead vanes may be used to develop optimum compression ratios.
2 Claims, 7 Drawing Figures PAIENI DEC-1 0mm SHEET 10F 5 FUEL BURNER EXPANDER COMPRESSOR FIG.
-FUEL PATENTEUUEEIOIEN 3,853,434
SHEET 2 OF 5 lfllI lHlll POSITIVE DISPLACEMENT ROTARY MACHINE This is a continuation, division. of application Ser. No. l32,l40 filed Apr. 7, 1971, now US. Pat. No. 3,741,694.
BACKGROUND OF THE INVENTION The present invention relates to fuel-burning engines used as power sources for propulsion and like purposes, and is particularly concerned with an improved engine of this type which minimizes emissions to the atmosphere while affording high-efficiency performance and an inexpensive and simplified construction.
In recent years governmental standards respecting permissible levels of emissions to the atmosphere from fuel-burning engines, such as automobile engines and the like, have become increasingly stringent. Such standards are, moreover, likely to be tightened still further in the future. Accordingly, a definite need exists for an engine which will comply with whatever emissions requirements might be established and, at the same time, meet the performance and most requirements of industry. This need is especially acute in the mass market for engines of under approximately 400 hp Currently available power plants, even when modified to improve anti-pollution characteristics, have failed fully to resolve this problem. Typically, such prior art engines are quite inefficient thermodynamically, i.e., they have poor fuel economy, which tends to neutralize the overall effectiveness of whatever emission control systems they incorporate by dumping greater quantities of exhaust gases and waste heat into the atmosphere for a given amount of work. Of these, engines operating on theOtto or Rankine cycles are representative.
In addition, secondary factors, such as elaborate control systems, excessive wear susceptibility, and high noise and vibration levels, associated with the foregoing engines further detract from their acceptability. For example, gasoline reciprocating engines (Otto cyle) require numerous sliding and bearing members, each a source of friction loss, to seal or retain the high pressure combustion gases.
Although mass flow turbine engines, operating as they do on the Brayton cycle, inherently have a low emissions capability, they are basically a constant speed, single throttle setting power source. This feature, coupled with its dependency on critical, hence expensive, materials of construction and bulky regenerators for high efficiency operation, severely restrict the turbines functional utility for many applications, as well as rendering it unattractive from a cost standpoint. Another notable disadvantage of conventional turbines is their high specific fuel consumption.
It is the purpose of the present invention to overcome these and other objectionable aspects of the prior art.
SUMMARY OF THE INVENTION a posttive- 60 Basically, it includes three separate components, a roexpander for cooling purposes and to seal the highpressure expansion stages against leakage. For higher efficiency, a regenerative heat exchanger may be added, suitably at a point in the fluid flow path between the compressor and the combustion chamber.
In a preferred assembly, the compressor and expander are arranged to operate on the same shaft, with part of the power developed by the expander being used to drive the compressor. The burner conveniently may constitute a single combustion chamber of annular form positioned in encircling relation to the shaft. This expander-compressor unit affords a compact, lightweight power plant having inherently balanced rotating parts, which, together with the generally uniform fuel combustion rate, contributes to reduced noise and vibration levels and a relatively simple construction of the main drive shaft and bearings. Being of the Brayton cycle, the engine combines the superiority of mass flow units operating on that cycle in respect of high thermodynamic efficiency and low emissions with a capability, owing to its positive-displacement nature, for high efficiency performance over a wide range of operating speeds and multiple throttle-settings. Moreover, it is operable with a variety of fuel types and qualities, with low specific fuel consumption, and does not require an elaborate control system.
Certain novel features of the compressor and fluid expander afford still further advantages in permitting the use of inexpensive, non-critical materials, significantly reducing wear and thus assuring long part life, and simplifying the sealing of high pressure regions against the leakage of combustion gases or lubricants. These features not only tend to reduce the cost of manufacture, weight and size of the engine as a complete unit, but add to the attractiveness of the expander and compressor for independent application.
More particularly, the fluid expander of the invention has the overall organization of a vane motor in which a plural or multiple stage expansion chamber is formed by an outer ring and an inner member, one of which is mounted eccentrically relative to the other so as to give the chamber a generally crescent-shape in cross section, and a plurality of vanes in the form of rockerslippers carried by one of the members. Valving is provided by a pair of valve plates bounding the sides of the chamber. High-temperature, high-pressure gases from the combustion chamber are admitted to the expansion chamber, by the-inlet valve plate, at a point opposite the region of highest compression. Here the gases are trapped in circumferentially spaced compartments or stages defined between adjacent rocker-slippers that, owing to the eccentric relationship of the inner and outer members, permit expansion of the gases and hence delivery of the energy contained therein to the expander, thereby generating usable mechanical power in the form of rotational velocity energy.
The foregoing components of the expander may be organized in three basic configurations. One, where both the outer ring and the inner member are mounted for rotation relative to the housing; two, in which the outer ring is grounded, i.e., heldstationary, relative to the expander housing and the inner member is supported for nutating motion; and three, where the inner member is held stationary and the outer ring nutates. In all three configurations, a suitable drive connection is made with the main drive shaft to translate the rotating or nutating motion of the members, as the case may be, into rotational motion of the shaft. The valve plates are appropriately driven and located relative to the several expansion stages so as always to be in phase therewith for proper valving action. For convenience of construction, the rocker-slippers are preferably carried by the inner member, but may, if desired, be located on the outer ring. Each configuration results in significantly reduced relative velocities, as compared to the full shaft r.p.m. developed, at all points of contact between members in the expansion area.
It is a feature of the invention that leakage of the high-pressure combustion gases from the expansion compartments is substantially prevented with only very moderate surface-contact forces at the points of sealing engagement between the rocker-slippers and the facing wall of the opposing member. To this end, the rockerslippers are mounted, for rocking movement of their leading ends towards and away from the opposed facing wall, in recesses formed at circumferentially spaced locations in the member carrying them, e.g., the inner member. A fluid passage connects with each recess at a point behind the rocker-slipper, through which highpressure but relatively cool fluid from the compressor outlet is directed. The fluid flow beneath each rockerslipper balances the hot gas pressure in the associated expansion compartment and urges the leading end of the rocker-slipper into light sealing contact with the opposed facing surface, the rocker-slippers being given an asymetrical shape in the direction of their leading ends for this purpose. The compartments are thus sealed between the outer ring and the inner member without the need for mechanical pressure seals and while maintaining the pressure-velocity loading, i.e., conditions productive of wear, at the points of sealing engagement to a minimum. By sweeping around the sides of the rocker-slippers, the cool high-pressure fluid delivered to the recesses also retards leakage of the expanding gases along the valve plates. Loss of lubricating oil from the compartments is similarly opposed.
In addition, the fluid flow to and through the recesses serves as a coolant for the expander parts through which it travels, thus allowing the expander components directly exposed to the hot combustion gases to run at highly elevated temperatures. Upon continued expansion of the combustion gases as they progress to the exhaust, the fluid seeps into the compartments where it aids in any residual burning. Further fluid connections are provided between he compressor outlet and the expander to maintain a fluid bearing between each valve plate and the adjacent housing wall.
As another feature of the invention, a compressor of improved construction has two eccentrically mounted screw vanes, one of a left hand lead and the other of a right hand lead, that are located in corresponding cavities in the compressor housing so as to lead toward a common center discharge from suction intakes at either end of the housing. Rotation of the shaft causes the vanes to nutate in the cavities and, in so doing, to displace the fluid taken in spirally towards the center discharge, providing balanced loading within the compressor with'no resultant axial thrust and a steady, substantially non-pulsating output. The vanes free floating except for rotational restraint, are generally V-shaped in cross section and preferably are composed of flexible, nonmetallic materials. This construction eliminates the need for pressure seals and largely avoids heat distortion problems and the attendant end leakage which these problems frequently cause.
The eccentric mounting for each screw vane includes a generally cylindrical eccentric keyed to the drive shaft and a sleeve rotatably and resiliently supported on the external walls of the eccentrics. Upon rotation of the shaft, the sleeve is caused to roll around the interior of the compressor housing, moving the screw vanes in a nutating manner into and out of the corresponding cavities. The resilient mounting of the sleeve preserves the rolling contact between the sleeve and the housing during operation, and affords the very beneficial result of permitting manufacturing tolerances to be relaxed. As this arrangement involves only three modes of wear in the compression area, one rolling and two sliding movements, and since the relative velocities of the surfaces involved in the two sliding movements are greatly reduced from full shaft speed, a significant advantage is achieved in respect of wear of the materials in the compression area, making practical higher compressor speeds and consequent savings in the size and weight of the unit for equivalent output. Moreover, variable lead screw vanes may be used to provide one or more compression stages productive of any optimum compression ratio. Proper proportioning of the components in the high-pressure regions allows an isothermal compression cycle to be approached, even without the use of a heat exchanger.
BRIEF DESCRIPTION OF THE DRAWINGS For a better understanding of the invention, reference may be had to the accompanying drawings in which:
FIG. 1 illustrates schematically the functional relationship between the basic components of an engine constructed in accordance with the invention;
FIG. 2 is a side elevational view, partly in section, of the compressor taken alone, with certain parts broken away for clarity;
FIG. 3 is a side elevational view, partly in section, showing the compressor, expander and burner in preferred assembled relation;
FIG. 4 is a vertical sectional view taken along the line 4-4 in FIG. 3, looking in the direction of the arrows;
FIG. 5 is a vertical sectional view taken along the line 5-5 in FIG. 3, looking in the direction of the arrows;
FIG. 6 is an enlarged detail view in section taken along the line 6-6 in FIG. 5, looking in the direction of the arrows; and
FIG. 7 is a partial vertical sectional view of another embodiment of the expander.
DESCRIPTION OF AN EXEMPLARY EMBODIMENT The engine of the present invention is designed to operate on a steadily flowing fluid stream to generate mechanical power in rotary motion from energy in the fluid. The basic components of the engine and the functional relationships therebetween are portrayed schematically in FIG. 1. For brevity and convenience of description in connection with-FIG. 1 and the other views of the drawings, the working fluid in its initial form is assumed to be air. It will be understood, however, that any fluid of suitable properties may be used.
In fundamental form, the engine consists of a rotary air compressor 10, a burner or combustion chamber 12, and a rotary expander or power extractor 14. Conveniently the expander and compressor turn on a common shaft, with the compressor being driven by the expander. Atmospheric air is drawn into the compressor from either end (see the arrows in FIG. 1) to be compressed and discharged at high pressure through a common centrally located discharge duct 16 to the burner 12. Fuel is injected into the combustion chamber through one or more fuel nozzles, where it mixes with the primary air entering through the duct 18 and burns. Secondaryv air is admitted to the burner through a branch duct 20 to further the combustion process. A portion of the high pressure air from the compressor outlet is delivered through a connecting duct 22 to the expander for a purpose hereinafter made clear.
The combustion gases exiting from the burner 12, in contrast to the high-pressure but relatively cool air flowing from the compressor, are at both high pressure and high temperature, that is to say, at a high energy state. The combustion products leave the burner 12 through a duct 24, actually a potted nozzle box as is described hereinafter, to enter the expander 14. In the expander, the gases are received in a multiple-stage expansion chamber of increasing volume and expand progressively as they travel from the inlet to the exhaust, falling in both temperature and pressure as they deliver energy to the rotary elements of the expander and thence to the drive shaft. Finally, the gases are exhausted through a duct 25 to waste or to a heat exchanger (not shown) for regenerative purposes. Suitably, a regenerative heat exchanger, if used, is interposed between the compressor 10 and the burner 12 to raise the energy state of the compressed air prior to its entry into the combustion chamber.
Viewing the basic engine depicted in FIG. 1 overall, certain fundamental operational advantages of the en gine over prior art plants become apparent. First, since the engine operates on the Brayton cycle, it has the innate superior characteristics of that cycle respecting high thermodynamic efficiency and low emissions. Similarly, it has the ability to operate on inexpensive fuels of various types and' qualities. Utilizing positive displacement, as contrasted to the mass flow design typical of conventional Brayton cycle units, the engine is not sensitive to variations in operating speed, exhibiting a capability for high torque operation at variable shaft speeds or at multiple throttle settings with low specific fuel consumption. As a further feature, the design lends itself to the addition of a second independent expander,
by virtue of which the drive shaft could be run down to substantially zero r.p.m. with an increase in torque. Also, the second expander could be arranged to run in reverse at nearly full torque.
Turning now to the individual components of the engine and with particular reference to FIG. 2, the compressor 10 is seen to includea housing 26, composed of any suitable metal, formed with peripheral cooling fins 28 spaced along its length. The use of such fins is optional, and other means such as a water jacket or the like may be used instead for cooling purposes. The housing 10 is closed at either end by an end plate 30, and the whole is held in assembled relation by appropriately positioned assembly bolts 32. As the compressor is of the center flow type, wherein air is taken in at both ends and discharged through a common center outlet, the axial halves of the compressor are of substantially identical construction, excepting that the lead direction of the compressing element is reversed.
Therefore only the features of one half of the unit are shown in detail in FIG. 2.
Each head 30 is formed with one or more inlet passages 34 that connect with an annular chamber 36 in the nature of an intake manifold. Inwardly 0f the air inlet passages, the heads contain a counterbore 38 in which is mounted an anti-friction bearing 40. While a roller-type bearing is illustrated in FIG. 2, any suitable anti-friction bearing may be used, including, for example, those of the positive oil-feed variety.
A collar 42 is rotatably supported within each bearing and is keyed, as at 44, to a drive shaft 46 extending axially through the housing 26. The compressor is sealed to the environment by a seal ring 48 of any appropriate design located between the shaft 46 and the surrounding head 30. As previously mentioned, the shaft 46 may be coupled to the expander 14. However, where the compressor 10 is used independently of the complete engine, any suitable source of power may be connected in driving relation to the shaft 46.
Also keyed to the shaft 46, and integral with or otherwise fixed to the interior end of the collar 42, is a generally cylindrical eccentric 50 whose axis is offset relative to that of the shaft. Concentric with the eccentric 50, and rotatably supported on the outer surface thereof by that resiliently carries, through an elastic bushing 56 or the like, a cylindrical sleeve 58 of an axial length substantially coextensive with the interior of the housing 26.
A counterweight 60 is keyed to the shaft 46, as at 62,
internally of the eccentric 50 to neutralize the unbalancing effect of the eccentric.
Resting on the outer surface of the sleeve 58, but otherwise free-floating except for rotational restraint relative to the housing 26, is a helically shaped vane 64 having a generallyV-shaped cross section. A cavity 66 of corresponding configuration is formed in the cylindrical inner wall 68 of the housing 26. Asdepicted in FIG. 2, the vane 64 and cavity 66 have a right hand lead.
Upon rotation of the shaft 46, the movement of the eccentric 50, which travels at full shaft r.p.m., .vithin the spacer 54 causes the sleeve 58 to roll around the internal wall 68 of the housing and, in so doing, causes the vane 64 to nutate in the cavity 66. In this way, air drawn through the inlet passages 34 is displaced spirally along the cavity 66 towards the central discharge outlet 70 at full shaft speed. The vane 64 preferably is composed of nonmetallic material, such as a plastic material for example, of suitable flexibility to make tight sealing engagement with the walls of the cavity 66. In this connection, the V-shape of the vane contributes to a tight seal by directing components of the air pressure forces acting on the vane normal to the seal forming surfaces. Accordingly, the need for rotating pressure seals, required in con'ventional compressors, is eliminated. Moreover, this construction largely avoids heat distortion problems that frequently lead to end leakage in conventional units. It will be apparent also that close manufacturing tolerances between the vane and the cavity and between the sleeve and the housing are not required, in part because the resiliency of the bushing 56 preloads the sleeve into contact with the housing wall 68 and maintains such contact throughout operation.
The intake phase of the compressing action consists of the first vane section, i.e., the first turn of the helix, rotating through one revolution. This revolution traps the volume of air that later is to be compressed. Preferably, a constant lead is used at this stage for maximum intake capacity. Thereafter, the air is compressed in one or more stages (turns of the helix) that compress the air to match the outlet pressure. If desired, variable lead vanes may be used to produce any desired optimum compression ratio.
Since compression proceeds equally from both suction ends toward the center, axial loads within the compressor are balanced so that no resultant axial thrust is produced. In addition to the vanes 64 having reverse leads in the separate halves of the compressor 10, they are indexed 180 from each other and the eccentrics are similarly offset to produce a continuous, rather than pulsed, flow at the outlet 70, notwithstanding that the flow in each half is sinesoidal. This system, therefore, produces a highly efficient unit, having no pneumatic line pulsations, which requires less horsepower than conventional compressors for a given throughput.
A significant feature of the compressor design resides in the low-pressure velocity loading achieved at points of wear, thus assuring long part life for traditional materials and making possible the use of materials having other beneficial properties but hitherto thought not suitable because unduly susceptible to wear. As an example, materials having high effective heat transfer characteristics, but relatively low wear resistance, might be used to enhance the dissipation of compression heat, a result desired in many applications.
Wear occurs in the compressing areas in principally three ways, these being the rolling, relatively frictionfree contact of the sleeve 58 with the internal wall 68 of the housing, the sliding engagement of the vanes 64 with the walls of the cavities 66 during nutating movement therebetween, and the sliding contact between the inner surface of the screw vanes and the outer surface of the sleeve. By its very nature, little wear results from the rolling contact between the sleeve 58 and the housing wall 68. Moreover, because the sleeve 58 nu.- tates relative to the eccentric 50, the velocity of its motion is on the order of from one-eighth to one-tenth of the shaft speed. This significant reduction in the rotational speed of the sleeve 58 and therefore of the numtional motion of the vanes 64, results in a corresponding decrease in the sliding velocity between the sleeve and the vanes and the vanes and the housing wall. Accordingly, wear at these points of part contact is greatly reduced. With certain working fluids lubricants can be dispersed with altogether; alternatively, low friction materials can be used in fabricating the sleeve 58 and vanes 64.
In addition to the aforementioned advantages in respect of longer part life and new material use, reduced wear also makes practicable higher shaft speeds, with consequent savings in compressor size and weight.
Another advantage of the compressor geometry of the present invention is its ability to provide very large cooling surfaces in the high pressure regions. Indeed, a cooling surface of sufficient size to approach isothermal compression may be designed into the unit. Variable lead vanes may also be used to increase thermal efficiency and to further reduce size. Heat transfer from the compressor is further aided by the velocity effect of the volumes of air trapped by the vanes 64, since they traverse from the inlet to the outlet at full shaft speed. In many instances, therefore, a heat exchanger is not needed to remove compressor heat.
As previously mentioned, the compressor 10, burner 12, and the expander 14 are preferably coaxially aligned, with the shaft 46 serving as a common drive shaft for both the compressor and the expander (see FIG. 3). To this end, the expander housing 72, which is also cylindrical in cross-section, is secured by bolts 74, or any other suitable means, to an end plate of the compressor. The burner housing 76 is similarly at tached to the other end of the expander housing. The shaft 46 may be extended through the burner housing 76, as indicated in phantom in FIG. 3, for power takeoff purposes. This arrangement of the components provides a compact, light weight unit that allows the use of a simplified main drive shaft and bearing construction and whose principal rotating parts are inherently in balance. It further simplifies the functional interconnection of the separate components.
Thus, the high-pressure air discharged from the compressor flows only a short distance, suffering little pressure drop,'through the primary air duct 18 and the secondary air duct 20 to the burner 12, which conveniently may constitute a single combustion chamber in annular form lying in surrounding relation to the shaft 46 (see FIGS. 3 and 4). Fuel is sprayed into the combustion chamber, as, for example, through the line 78 and nozzle 79 (see FIG. 4) where, in accordance with conventional practice, it intermixes with the primary and secondary air and is burned, thus increasing the energy state of the air stream. Desirably, combustion goes on continuously and at a generally uniform rate. One advantage of such combustion conditions is that noise and vibration levels generated by the engine are kept down. This advantage is reinforced by the aforementioned inherent balance of the engines rotating elements.
After flowing through the combustion chamber (see the arrows in FIG. 4), the air stream, now in the form of combustion gases, is discharged through a potted nozzle box 80, having a connecting passage 82, to enter a torus 84 formed in the adjacent end cover 86 of the expander 14 (see FIG. 3), from where the gases are admitted to the interior of the expander through an annular slot 88.
The expander includes a second end cover 90 spaced from the end cover 86 by a cylindrical member 92 lining the interior of the housing 72. The end covers 86 and 90, which along with the cylindrical member 92 are fixed relative to the expander housing, each have a central bore 94 for receiving cylindrical members 96 and 98, respectively, that are formed integrally with or otherwise supported by the burner housing 76 and the expander housing 72, respectively. Anti-friction bearings 100 of any suitable design are provided on the internal walls of the member 96 and 98 for rotatably supporting the main drive shaft 46. A positive oil-feed bearing is illustrated, but other bearing types may be used.
Intermediate to the facing ends of the cylindrical members 96 andx98, the shaft 46 includes an adapter 102 that is peripherally splined, as at 104, to a cylindrical-rotor 106 (see also FIG. 5). The adapter 102 and rotor 106 are arranged to be concentric with the shaft i 46 and, of course, to be rotatable therewith. Positioned in a surrounding relation to the rotor 106 is an outer ring 108 held between two radially extending annular valve plates 110 and 112. The valve plates are supported at their inner ends by collar-like members 114 and 116 that are rotatably mounted, through suitable anti-friction bearings 118 and 120, on the external surfaces of the cylindrical members 96 and 98, respectively. While the collars 114 and 116 are coaxial, their axis is offset from the axis of shaft 46 so as to be eccentric thereto. Hence, as illustrated in FIG. 5, the outer ring 108 is eccentric to the rotor 106. Appropriate spring loaded gaskets 122 are provided between the ends of the collars 114 and 116 and the adjacent housing walls. It will be apparent, therefore, that both the inner rotor 106 and the outer ring 108 are mounted for rotation within the expander, though about offset axes.
Because of the eccentric relationship between the rotor 106 and the outer ring 108, a chamber 124 having a generally crescent shape in cross-section (see FIG. is defined between the outer wall 126 of the rotor 106 and the inner wall 128 of the outer ring 108; that is to say, the chamber 124 has a region of minimum volume,
indicated generally at 130, and a diametrically opposed region of maximum volume, indicated generally at 132, separated in the direction of rotation by a region of progressively increasing volume, indicated generally at 134, and a region of progressively decreasing volume, indicated generally at 136. The chamber 124 is closed at either axial end by the valve plates 110 and 112 (see FIG. 3). Accordingly, gases admitted at high pressure to the chamber 124 at the region of minimum volume, i.e., the region of highest compression, will be caused to expand as they progress through region 134 in the direction of the region, at- 132, of lowest compression, i.e.,maximum volume. So as to reduce the load imposed by the expanding gases over any particular portion of the rotor 106 and the outer ring 108, it
is desirable to carry out this expansion in stages. To this end, a plurality of vanes 138 (see FIG. 5) in the form of rocker-slippers are provided on the rotor 106 at spaced locations about its periphery.
Each rocker-slipper 138 is located in a recess 140 formed in the outer wall 126 of the rotor, and is supported therein, through a pin 142 received by a post 144, for rocking movement of its leading end 146 towards and away from the inner wall 128 of the outer ring 108. The leading end 1460f each rocker-slipper is urged into engagement with the wall 128 partly by centrifugal force and partly by a flow of high pressure air, taken from the compressor outlet, admitted tothe recess 140 through a passage 148 (see FIG. 6) formed in the rotor 106 and entering the recess at a point behind the rocker-slipper. It is for this purpose that the direct connection through duct 22 (see FIG. 1) is made between the compressor outlet and the expander. Due to pressure losses in the working air stream between the compressor and the expander, the air admitted to the recesses 144 is at a slightly higher pressure than the combustion gases exhausted from the turner 12. The high-pressure air acting on the underside of the rockerslippers therefore balances the forces'exerted by the high-pressure combustion gases on the rocker-slippers so as to maintain a light sealing contact between the leading end 146 of each slipper and the wall 128. To facilitate the establishment of this sealing force, the
rocker-slippers 138 are given an asymetrical shape in the direction of rotation, i.e., a greater surface area is provided on the leading side of the pin 142 than on the trailing side. In this way not only is the need for sliding pressure seals of conventional construction eliminated, but a low pressure-velocity loading is achieved at the point of contact of the leading ends 146 of the rocker slippers with the opposed wall 128 due to the substantially balanced relationship between the forces exerted by the expanding combustion gases within the chamber 124 and the high-pressure air flowing through the recesses 140.
The air flow to the recesses 140 also seeps around the sides of the rocker slippers (see FIG. 6) to flow through the clearance space provided between the slippers and the adjacent valve plates 110 and 112, preventing the leakage of the combustion gases through those spaces. Seal rings 149 are interposed between the rotor 106 and the valve plates to retard air escape internally of the rotor.
By virtue of the foregoing construction, a series of separate expansion compartments, or stages, each of which is enveloped, in effect, by a positive, or at least neutral, pressure difi'erential directed against leakage of the combustion gases therefrom, is provided within the expansion chamber 124.
Moreover, since the air from the compressor outlet, though at high pressure, is relatively cool compared to the hot combustion gases, it functions to cool the rotor 106, rocker-slippers 138 and other parts with which it comes in contact, and thereby allows the combustion gases to be elevated to operating temperatures higher than would be practicable from the standpoint of part deterioration without such cooling.
Entry of the combustion gases to the chamber 124 is controlled by the inlet valve plate 110 formed, for this purpose, with a narrow inlet slot 150 (see FIG. 5) of relatively short circumferential extent. In a similar vein, the outlet valve plate 112 contains an exhaust port 152 for venting the expanded gases from the chamber 124. An exhaust torus (not shown) is provided in the end cover to receive the gases and communicate them to the outlet exhaust duct 25. As depicted in FIG. 5, the exhaust port 152 is of considerable width and circumferential extent. Since both the rotor. 106 and the outer ring 108 rotate during operation, a series of pins (not shown) are provided in the rotor 106 to engage the valve plates 110 and 112 in a cycloidal manner to locate them relative to the rotor and so as always to maintain the inlet port 150 andv the exhaust port 152 in phase with the appropriate regions of the expansion chamber 124. That is to say, the inlet port 150 is positioned opposite the region of maximum compression or minimum volume, while the exhaust port 152 is positioned to begin shortly after the region 132 of maximum volume and to extend over nearly the full length of the region 136 of decreasing volume. v
As each rocker -slipper 138 moves past the inlet port 150, a charge of high-temperature, high-pressure combustion gases is delivered to the expansion compartment or stage defined between that slipper and the following slipper. Thereafter, these gases become trapped, by virtue of the sealing engagement of the leading ends 146 of the slippers with the outer ring 108,
within the compartment upon movement of the following slipper beyond the inlet port. An identical charging and entrapping process takes-place in connection with the other compartments. Thus, as the compartments progress through the regions 134 and 132 towards the exhaust port 152 the gases within them are caused to expand, falling both in temperature and pressure as they deliver energy to the rotor 106. This energy is transmitted through the spine 104 to the shaft 46 in the form of rotational velocity, thereby generating usable mechanical power. It will be understood, of course, that this procedure goes forward cyclically with each revolution of the main shaft.
It will further be appreciated that since power is generated on the basis of high compression ratios, high torque outputs can be developed at variable shaft speeds or at multiple throttle settings. Moreover, bulky regenerative heat exchangers of the type required with mass flow engines are not needed to attain high efficiency in view of the capability of the design for using materials of extreme resistance to heat, such as low expansion graphite or ceramics, in the expansion area, i.e., the end covers 86 and 90, cylindrical member 92, rocker-slippers 138, posts 144, rotor 106 and outer ring 108. This makes practicable operating temperatures considerably higher than those within the reach of conventional engines. These factors, high compression ratios and elevated operating temperatures, afford superior specific fuel consumption performance.
Another beneficial aspect is that as the expansion compartments move away from registry with the inlet port 150, with a consequent pressure drop with increase in volume, a certain portion of the cool air flowing through the associated recesses 140 seeps into the compartments where it aids in burning any residual combustibles. This further increases the energy delivered to the rotor 106.
While the rotor 106 turns at full shaft rpm, the relative velocity between the rotor and the outer ring 108 is significantly slower. This aspect of the invention, combined with the light pressure loading established between the rocker-slippers and the outer ring, results in very favorable wear conditions at the point of engagement of the leading ends 146 of the rocker-slippers with the inner wall 128 of the outer ring. These components, therefore, may be fabricated from materials selected primarily for heat-resistance rather than for wear-resistance considerations. As mentioned, high efficiency operation may thereby be achieved without sacrificing engine lifetime due to wear.
In that regard, wear producing contact between the valve plates 110 and 112 and the adjacent end covers 86 and 90, respectively, is prevented through the establishment of an air bearing in the clearance spaces between those members. For this purpose, appropriate connections (not shown) are made withthe compressor outlet to deliver a flow of high-pressure air to the space between the outer ring 108 and the cylindrical member 92, this air then flowing downward between the valve plates and the end covers (see FIG. 6) to set up the air bearings.
Although in the embodiment of the expander 14 described above, the rotor 106 and the outer ring 108 are both rotatably mounted, this need not be the case. For example, a construction could be used where the outer ring 108 is grounded to the expander housing 72, i.e., held stationary, and the inner ring mounted for nutating movement within the outer ring. In an alternative arrangement, the rotor 106 could be held stationary while the outer ring nutates around it. As a further modification, the rocker-slippers 138 need not be located on the rotor 106, but instead could be carried by the outer ring 108. In describing these possible alternative constructions, the inner member has been referred to as the rotor and the outer member as the outer ring; but this has been done merely for convenience of illustration and is not intended to assign specific functions to these members. Depending on the arrangement used, a suitable driving connection would be made between one of the inner or outer members to translate the motion imparted to it by the expanding gases to the drive shaft 46. Likewise, an appropriate connection would be established with the valve plates 1 10 and 112 to provide the proper valving action.
FIG. 7 portrays an embodiment of the expander 14 incorporating the stationary outer ring-nutating inner member variation. In the FIG. 7 embodiment, the outer ring 154 is fixed directly to the expander housing 72 between the end covers 156 and 158. An annular cavity 159 may be provided in the outer wall of the ring 154 to reduce its weight and for cooling purposes. The construction of the end covers 156 and 158 is essentially unchanged from the previous embodiment, except that recessed portions 160 are provided to receive the valve plates 162 and 164, which are now freefloating relative to the end covers and the outer ring. The rocker-slippers 168 are of identical construction to those previously described. The inner member 170 carrying the slippers 168, however, is splined, as at 172, to
a cylindrical adapter 174 rather than directly to the shaft. The adapter 174, in turn, is supported on an eccentric anti-friction bearing 176 whose axis is offset from that of the shaft46. At the exterior end of the adapter 174 is provided a collar 178 that is splined through a crown spline 180 on its internal wall to a cylindrical member 182 splined, as at 184, at the other end to the burner housing 186, thereby grounding the adapter 174 and inner member 170 against rotation but allowing nutating movement thereof relative to the outer ring 154 upon rotation of the drive shaft 46.
In order that the valve plates 162 and 164 are kept in proper phase for charging and exhausting the expansion chamber, a drive connection is established between each plate and the shaft 46. Accordingly, the exhaust valveplate 164 is pinned, at 188, to the eccentric bearing 176, and the intake valve plate 162 is connected in a one-to-one driven relationship to the shaft 46 through a ring gear l90-that is rotatably supported on anti-friction bearings 192 in concentric relation to the drive shaft.
With entry and expansion of the combustion gases within the compartments of the expander depicted in FIG. 7, the inner member 170 is caused to move in a nutating fashion within the outer ring 154. Whereas this motion is relatively slow, thus giving a good pressure-velocity loading factor in respect of wear of the rocker-slippers 168 and the outer ring 154, it is translated by the eccentric bearing 176 to high speed (on the order of ten times faster) rotational motion of the main shaft. As before, theexpanded gases leave the expansion chamber through the exhaust valve plate 164 to be received in the exhaust torus 194 in the end cover 158.
The valve plates 162 and 164 are prevented from contacting the end covers 156 and 158 or the outerring 154 by air bearings established in the running clearances between these members. Again, suitable seal rings 196 are located between the valve plates and the inner member 170.
As has been suggested above, many modifications and variations in the specific embodiments of the invention described herein may be made by one skilled in the art without departing from the inventive concepts disclosed. Accordingly, all such modifications and variations are intended to be included within the spirit and scope of the appended claims.
1. A compressor comprising:
a housing having a generally cylindrical internal wall,
greater than the internal diameter of the internal housing wall but less than the external diameter of the cavity and of a width in radial dimension at least as great as the cavity is deep,
means restraining rotational movement of the vane relative to the cavity while permitting nutating movement therebetween,
a generally cylindrical sleeve positioned within the vane for engagement with the inner surface thereof,
a drive shaft generally concentric with the housing internal wall, and
an eccentric mounted on the drive shaft for rotation therewith, the sleeve beingsupported on the external wall of the eccentric in concentric relation therewith, whereby upon rotation of the shaft the sleeve is caused to impart nutating motion to the vane within the cavity so as to displace fluid spirally from the inlet to the outlet.
2. A compressor according to claim 1 wherein the vane is composed of flexible, nonmetallic material.
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|US3238882 *||Feb 19, 1964||Mar 8, 1966||G Datwyler Dr||Helical pump|
|DE1181550B *||Mar 29, 1957||Nov 12, 1964||Jean Delsuc||Drehkolbenpumpe|
|Citing Patent||Filing date||Publication date||Applicant||Title|
|US6074184 *||Aug 13, 1997||Jun 13, 2000||Imai; Atsushi||Pump utilizing helical seal|
|US6425744||Apr 30, 2001||Jul 30, 2002||Kabushiki Kaisha Toshiba||Helical blade type compressor having a helical blade in a stationary cylinder|
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|US20130207401 *||Feb 10, 2012||Aug 15, 2013||Saade Youssef MAKHLOUF||High efficiency radioisotope thermodynamic electrical generator|
|U.S. Classification||418/55.1, 418/60, 418/220, 418/56|
|International Classification||F01C11/00, F01C1/44|
|Cooperative Classification||F02G2250/03, F01C1/44, F01C11/008|
|European Classification||F01C11/00C2, F01C1/44|