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Publication numberUS3626911 A
Publication typeGrant
Publication dateDec 14, 1971
Filing dateFeb 2, 1970
Priority dateFeb 4, 1969
Publication numberUS 3626911 A, US 3626911A, US-A-3626911, US3626911 A, US3626911A
InventorsShaw Harry
Original AssigneeTechnology Uk
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Rotary machines
US 3626911 A
Abstract  available in
Images(4)
Previous page
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Claims  available in
Description  (OCR text may contain errors)

United States Patent [72] Inventor Harry Shaw Aldershot, England [21] Appl. No. 7,678 [22] Filed Feb. 2, 1970 [45] Patented Dec. 14, 1971 [73] Assignee Minister of Technology in Her Brltannic Majesty's Government of the United Kingdon of Great Britain and Northern Ireland London, England [32] Priority Feb. 4, 1969 [33] Great Britain [31 5,893/69 [54] ROTARY MACHINES 12 Claims, 6 Drawing Figs.

[52] U.S. Cl. 123/45, 418/195 [51 Int. Cl. F02b 53/00 [50] Field of Search 123/45, 18, 8.09, 8.07, 8.13, 8.47; 418/195, 233

[56] References Cited UNITED STATES PATENTS 2,398,690 4/1946 Bolli 123/ 45 Primary Examiner-Clarence R. Gordon Attorney-Stevens, Davis, Miller & Mosher ABSTRACT: An engine has two rotor drums having their axes set at an included angle of 150 to each other. A similar number of bores is provided in each rotor and are connected by sleeves which are cranked at their midpoints to the same angle as that between the rotor axes. The sleeves are free to slide and rotate in the bores and have their outer ends closed by cylindrical members attached to the respective rotors and extending axially into the sleeves. On rotation of the rotors, the cylindrical members reciprocate relative to the sleeves and may be regarded as opposed pistons. inlet ports are provided in one arm of each sleeve and come into cyclic register with inlet ducts by rotation of the sleeves in the bores. Large exhaust ports in the other arm of each sleeve are uncovered by relative axial movement of the respective cylindrical member. Port arrangements suitable for use in a compressor of generally similar configuration are described and combinations such as engine/compressor, power unit/transmission are envisaged.

mum] mic-145m SHEET 1 0F 4 PATENTEDDEBMBH. 3.626.911

SHEET [1F 4 MAX. VOLUME I POSITION.

IN. vo UME PO ITION.

MAX. VOLUME POSITION.

MIN. VOLUME POSITION.

ROTARY MACHINES This invention relates to rotary machines having rotors in which pistons move in reciprocating fashion relative to working fluid.

Rotary piston machines are, in general, much simpler than conventional reciprocating piston machines in that they have fewer moving parts and dynamic balancing if far less of a problem. Higher rotational speeds (and hence higher outputs) are practicable and the almost total absence of vibration that can be achieved means that mountings can be lighter and less complicated than for reciprocating piston machines.

In many rotary piston machines, however, components which define the working volumes are of irregular shapes which lead to difficulties in maintaining adequate sealing between them when in relative motion, due, for example, to dimensional instability under varying operating and ambient conditions with consequent leakage of working fluid at critical stages of operation. Heat loss can also be a problem due to the incidence of large surface areas relative to working volumes. Either of these factors can result in low efficiencies and the effect is aggravated where they occur in combination. Lowcombustion efiiciencies in internal combustion engines can cause expulsion of unburned or partially burned hydrocarbon products from the exhaust and, apart from being wasteful, can even be illegal.

The ideal working volume in terms of minimum surface area is a sphere but the difficulties in making an expandible chamber of this form which could be used to give a power output in usable manner are almost insoluble.

The most practicable shape is a cylinder with flat ends, the minimum surface area in relation to volume occurring when the length and diameter are equal. Thus, discounting the effects of turbulence and temperature differentials between a gas and its enclosing cylinder walls, the least waste of potential output is likely to be achieved in a reciprocating piston machine during those parts of the operating cycle where a piston moves between its innermost position (i.e. T.D.C.) and a point in its cylinder equal to the diameter of the piston, especially if the total piston movement is of the order of twice this distance. In an internal combustion engine, for instance, this would encompass combustion and at least the initial stages of expansion (i.e. that period at which the gas temperature is highest).

Another factor which helps to reduce heat losses is a lowmean relative velocity between the center of a gas mass and its encompassing chamber, particularly during that part of an operating cycle when the average temperature difference between the gas and the chamber walls is at a maximum and the density of the gas is at its highest (this of course is again during the early stages of expansion in an internal combustion engine). Gas velocity can be reduced by reducing piston movement (stroke) relative to cylinder diameter (bore).

Arising out of the above considerations, it is believed that optimum values are obtainable in reciprocating piston machines having a bore/stroke ratio of the order of l/ 1.5 though this can be modified by other factors such as compression ratios and speed of operation. For example, the quicker the speed of operation, the more heat is likely to be lost in a given time but less per cycle.

It can, incidentally, be shown that no mean relative gas velocity occurs in certain designs of opposed piston machine in which there is no phase difference in the movement of the pistons.

It is largely due to these considerations that reciprocating piston machines, despite drawbacks in other directions, have become widely used particularly as compressors and internal combustion engines. In the course of time they have incidently been developed to a stage where their reliability leaves little to be desired from the practical point of view.

The cylindrical components involved are, incidentally, relatively simple to manufacture and many specialized machine tools for this purpose are readily available. Split spring rings (piston rings, for example) used to contain gas leakage between sliding cylindrical surfaces are among the most effective and simple sealing devices available being compatible with wide ranges of operating temperatures and associated expansions.

The present invention is directed to the introduction of certain advantageous features of reciprocating piston machine construction into rotary machines.

According to the invention a rotary machine comprises a pair of rotors mounted for rotation about axes which are inclined to each other at an angle between the perpendicular and something less than 180, an equal number of cylindrical bores in each rotor disposed regularly about the respective axis, sleeves serving to connect the bores of one rotor with those of the other, the sleeves being cranked at an intermediate point to the angle subtended by the rotor axes with their opposite arms disposed in the bores of separate rotors and being free to slide and rotate relative thereto, cylindrical members fixed relative to the respective rotors serving to close the outer ends of the sleeves and extending axially therein, a fluid inlet passage opening into each bore of one rotor, at least one port formed in one arm of each sleeve and disposed so as to connect the said passage with the interior of the sleeve cyclically in accordance with rotation of the rotor, and a further port or ports formed in the other arm of each sleeve and arranged to connect the interior of the sleeve cyclically with a fluid outlet passage communicating with each bore of the second rotor.

Preferably the sleeve ports are shaped and arranged so that the alignment of those first mentioned with the fluid inlet passage is controlled substantially by the relative rotation of the sleeves and the rotor, while the second mentioned are effectively opened and closed by relative axial movement between the sleeves and the appropriate cylindrical members.

In a preferred form of the invention the included angle between the rotor axes is of the order of Embodiments of the invention will now be described by way of example with reference to the accompanying diagrammatic drawings of which:

FIG. 1 is an elevation of a form of internal combustion engine, partly in section,

FIG. 2 is a similar view to that of FIG. I of an alternative form of internal combustion engine,

FIG. 3 is an isometric view of a sleeve member forming a part of the engine of FIG. 2, and

FIG. 4 illustrates the movement of inlet ports formed in the sleeve member of FIG. 3,

FIGS. 5 and 6 are schematic representations of inlet and outlet porting arrangements in a compressor.

The engine shown in FIG. 1 is depicted in simplified form for the purpose of illustrating the method of operation. The engine proper is supported on a frame comprising a base I and two side members 2, 3 rising from opposite ends of the base. The side members each carry one of a pair of stub shafts 4, 5 which project inwardly towards each other in the same vertical plane and are equally inclined upwardly from the horizontal at an angle of 15. Rotors 6, 7 are mounted on the stub shafts, their adjacent faces being suitably bevelled so as to lie parallel to each other and in close proximity along the line of the respective radii extending to the lowest circumferential point of the rotors. Rings of gear teeth 8, 9 are formed on the said faces of the rotors concentric with the respective axes of rotation, and mesh together at their point of closest proximity X, the rotors being thus constrained to rotate in phase with each other.

Six cylindrical bores, equally spaced round a pitch circle concentric with the rotor axis, pass through each rotor, their axes being parallel to the axis thereof.

Disposed between the rotors and serving as connections between them are six sleeves 10 which are cranked at their midpoints to the same angle as that between the axes of the stub shafts 4, 5 (i.e. 150), the opposite arms of each sleeve extending into one bore in each rotor. The arms of the sleeves fit closely within the bores but with such clearance that relative rotation and axial movement is possible.

Formed on the opposite axially outer faces of each of the rotors concentrically with the said bores are six rings 11, 12 whose radially inner surfaces constitute extensions of the bores. Long cylindrical hollow members l3, 14, each closed at one end and having flanges 15, 16 at the other, pass into the bores being attached to the rings 1 1, 12 by the said flanges and sewing to close off the ends of the sleeves into which their inner ends project. The inner portions of the cylindrical members 13, 14 are of increased diameter to enable them to fit closely within the opposite arms of the sleeves but again with such clearance that relative rotation and axial movement is possible; conventional split spring rings (of which piston rings are a typical example) carried by the cylindrical members bear against the inner surfaces of the sleeves to form sliding gas seals.

The inclination of the rotors results in corresponding pairs of bores moving alternately closer together and then farther apart during rotation of the rotors, the limiting positions being shown respectively at the bottom and the top of FIG. 1. The sleeves are constrained by their angularity to maintain their original alignment and there will thus be relative rotation between the sleeves and the bores equal to one revolution of each sleeve for each complete revolution of the rotors. At the same time, the sleeves will move in and out of the bores. The cylindrical members 13, 14 being attached to the rotors will consequently move axially in and out of their respective sleeves in a manner which can be likened to that of opposed pistons in a cylinder; a working volume of variable capacity is thereby defined by the sleeve walls and the cylindrical mem bers.

The right hand arm (according to FIG. 1) of each sleeve is pierced to give two large diametrically opposed ports 17. Gas manifolds 18 in communication with a gas supply source (not shown) are formed within the rotor 7 and passages 19 extend from them to openings in the bores of said rotor, which openings will come into register with the ports 17 to form gas inlets during such periods as the sleeves are withdrawn from the bores of the rotor to substantially the maximum extent, as shown at the top of FIG; 1. The passages 18 are inclined so that gas may flow smoothly towards the center of the sleeve as indicated by the arrows G.

The shaping of the ports 17 is such that they will be brought into precise register with the openings from the passages 19 by relative rotation of the sleeves, as will be described in greater detail later with respect to the embodiment of FIG. 2.

The left-hand arm (according to FIG. 1) of each sleeve is pierced to form ports 20 which extend peripherally over a substantial part of the sleeve circumference. Gas manifolds 21 fonned within the rotor 6 are connected by passages 22 to openings in the bores of the said rotor arranged to communicate with the ports 20 also during such periods as the sleeves are withdrawn from the bores of the rotor to substantially the maximum extent. Rotation of the rotor during such periods will result in axial movement of the cylindrical members 13 past the ports 20 which will thus be opened and while the ports are open, gas may flow from the interiors of the sleeves to the gas manifolds 20 and thence to exhaust by way of suitable ducting (not shown). This may be likened to the control of exhaust port opening and closing by piston movement inconventional two-stroke reciprocating engines. The closed ends of the cylindrical members 13 are shaped to direct gases freely towards the ports 20 and passages 22 so that rapid discharge of the gases is effected as the ports are opened.

The cylindrical member 14 may be likened to the junk head of a sleeve valve engine, its closed end being provided with a conventional sparking plug 23.

In operation, the opening of the ports 17 as a result of relative rotation between one of the sleeves and the associated bore of the rotor 7 permits gas to flow to the interior of the sleeve. Continued rotation of the rotor results in closing of the ports after which the gas is compressed as the cylindrical members 13, 14 move inwardly relative to the sleeve, the gas being ignited by the sparking plug at such time as the working volume within the sleeve approaches a minimum. The expansion so caused acts on the cylindrical members tending to force them outwardly of the sleeve, a torque thus being set up which is imparted to the rotors to maintain rotation. Relative movement of the cylindrical member 13 within the sleeve opens the ports 20 and the gases pass out to exhaust, the sudden release of pressure occasioned thereby assisting in the rapid purging of the sleeve and possibly also in the induction of further gas from the ports 17 which will have reopened. Further movement results in closing of the ports 20 before compression of the new gas charge.

The engine may in effect be regarded as a two-stroke reciprocating engine operating on the uniflow principle, in which gas admission is controlled by a continuously rotating sleeve valve and exhaust by a piston. Such an engine, in which a similar sequence takes place so far as the operating cycle is concerned, is disclosed in British Pat. Specification No. 1,141,398.

Power output may be taken from either of the rotors, or both, in any convenient manner. For instance, the rotors might be rigidly attached to their respective stub shafts which would be mounted for rotation in the side-members 2, 3 and provided with extensions connected to gearboxes or other transmission means. The engine could also be turned for starting purposes by such an arrangement. Another power takeoff will be described in relation to the embodiment of FIG. 2.

FIG. 2 shows a variation of the embodiment of FIG. I, the same reference numerals being used to indicate corresponding features. The main differences occur in the disposition of the sleeve ports and of the gas inlet and exhaust passages within the rotor. Inlet ports 40 are at the sides of each sleeve 41, rather than at top and bottom as previously, while gas flowing to them passes along a fixed duct 30 formed in the side member 3 and the stub-shaft 5 and thence to bifurcated rotating ducts 31 in the rotor 7 by way of transfer ports. One bifurcated duct is provided for each rotor bore at which the respective arms of the duct terminate as shaped entries 32 located diametrically opposite to each other.

Exhaust ducts 33 extend from exhaust ports 42 through the rotor 6 and by way of transfer ports to a fixed duct 34 formed in the stub shaft 4 and side member 2. A bevel gear 35 serves to transmit drive from the rotor 7 to an output shaft 36 via bevel gear 37.

FIG. 3 shows the sleeve 41 and the relative disposition of the inlet ports 40 and exhaust ports 42 in it. The exhaust ports in this case extend around 60 percent of the periphery of the sleeve.

The position when the cylindrical members of a particular pair of bores are at their maximum distance apart (i.e. when the volume defined by them within the sleeve is at a maximum) corresponds to the bottom dead center position in a conventional reciprocating piston engine and the position after of rotor movement when the cylindrical members are closest together corresponds to top dead center. The difference between these distances can be regarded as the stroke of the engine, which in the embodiment of FIG. 2 is 1.34 times the inside diameter of the sleeve, a ratio which gives a near minimum of metal surface subjected to the scrubbing action of hot gases and whereby heat losses from the gas during expansion will be kept to a low value.

The separation of the inlet and outlet ports means that neither limits the size of the other and adequate cross-sectional areas of port are thus obtainable within short axial dimensions along the sleeves.

The timing of the opening and closing of the exhaust ports is independent of that of the inlet ports.

A typical timing of the cycle of events might be:

Exhaust ports open at 67.5 of rotor rotation before B.D.C.

Inlet ports open at 50 of rotor rotation before B.D.C.

Exhaust ports close at 67.5 of rotor rotation after B.D.C.

Inlet ports close at 70 of rotor rotation after B.D.C.

The resultant axial length of the exhaust ports is 0.31 of the stroke and their maximum area is the same as the cross-sectional area within the sleeve.

The inlet ports open when the rate of exhaust port opening is still substantial and the area of exhaust port open at this point is 42 percent of the above cross-sectional area. In the interval between the opening of the exhaust ports and that of the inlet ports, the flow of exhaust gases out of the sleeves creates a reduction in pressure which is applied to the inlet gas stream helping to draw it into the sleeve.

Volumetric efficiency is determined by the amount of charge which can be induced to flow into the working chamber of a machine and is dependent largely on a time-area relationship of the inlet opening. Since it is important to have the largest possible area available over the period a port is open, the rapid opening and closing of ports obtainable in machines as heretofore described can readily be seen to be advantageous.

High rates of increase of inlet port area are achieved due to the constant angular velocity of the sleeves relative to the bores under equilibrium conditions. Typically, the rate of increase is of the order of percent of the cross-sectional area of a bore per degree of rotation. This enables a substantial rate of flow of gas into a sleeve while the pressure is low, leading to good scavenging and a high-charge density.

In comparison, poppet valves have to be accelerated from rest on opening (and brought to rest on closing) and a considerable time lag ensues before such valves are moving at their maximum velocity. Similarly, the change of direction of motion in an oscillating or oscillatory-rotary sleeve valve involves the acceleration of such a sleeve against inertia.

As can be seen from FIG. 4, each of the inlet ports 40 in a sleeve is substantially trapezoidal in shape with the side edges inclined outwardly at unequal angles and an arcuate top edge (which is towards the outer end of the particular arm of the sleeve); the comers of the port are rounded off. The angles of the side edges differ due to these edges being normal to the direction of relative motion for reasons which will appear later.

Each of the entries 32 in the bores is exactly the same shape as the inlet ports but reversed (i.e. a rotated polar" image) the longer, arcuate edge being at the bottom-or inwardly relative to the engine as a whole.

The inlet port shown in solid lines is in the position of maximum opening relative to the entry in the bore. Alternative positions are indicated in dotted lines, that at A representing the instant of port opening and that at B being the instant of closing. At these two positions the adjacent edges of the inlet port and of the entry are both normal to the direction of relative movement and thus will be in complete alignment, giving a maximum rate of opening or closing. Port movement relative to the entry is indicated by the curved lines 2-2 which represent the positions of the center of gravity of a port during a cycle of events.

For engines as described, the advantages of an opposed unifiow engine are obtainable without the disadvantages associated with reciprocating ports or the need for complicated sealing arrangements and with these are coupled the accepted benefits of sleeve valve utilization. Among the resultant features are:

A compact combustion chamber for high thermal efiicien- Minimum flame travel during combustion.

Separation of the inflow charge from vitiated gases (i.e. a large degree of stratification is possible whereby unbumt fuel mixture can remain in close proximity to the ignition area, also cooling effects can be counteracted).

Hot-exhaust areas are separated from the combustion zone enabling high compression ratios to be used without detonation.

Though there is relative reciprocating motion between component parts, all the working components rotate on circular or elliptical paths making for inherent balance. The absence of true reciprocating components, especially in the port operating arrangements, removes many of the restrictions which impose limits on the maximum speeds of reciprocating engines.

The maximum speed of an engine as herein described is therefore much higher than that of a reciprocating engine of the same swept volume with resultant increase in output, making for compactness in relation to rated power. Mountings, consequently can be light and simple.

Sealing is simple and straightforward; at such times as gas is required to be at high pressure it is contained in chambers from which leakage is normally prevented by sliding spring rings disposed between parts moving very much as in a conventional reciprocating piston engine.

Other seals, for instance between bore openings and manifolds, are required to sustain only comparatively low pressures.

Various modifications to the embodiments disclosed can be envisaged. Thus, the included angle between the rotor axes may be varied between about 170 and a right angle without affecting the working of the engine but an angle of as specified earlier is believed to be about the optimum.

Provision may also be made to move one or both cylindrical members l3, 14, or parts of them, axially relative to their respective sleeve so as to vary compression ratios obtainable to suit starting or changes in charge weight considerations.

Reaction between the sleeves and the rotor bores can be such that the rotors will turn in phase without the strict necessity for other forms of coupling. Thus the gears meshing between them might be omitted in some cases where torques developed between sleeves and rotors are of low values. Where very large torques are developed, these would preferably be transmitted in whole or in part by a gear drive.

Again, arrangements of rotor and stator blades could be provided between fixed inlet and exhaust ducts (e.g. 30 and 34 of FIG. 2) and the matching ducts in the rotors, acting respectively to assist in charging the working volumes and helping drive the rotors.

An exhaust porting arrangement similar to that described relative to the inlet ports can be used wherein the opening and closing of the exhaust is effected by relative rotation of ports in the appropriate part of a sleeve and the exhaust openings in the respective rotor bore, rather than by the relative axial movement of the ports and a cylindrical member. In this way, the exhaust events are not limited to being symmetrically disposed about B.D.C. It is therefore possible to arrange the timing of engines and other machines to suit a wide range of particular operating conditions.

Such an arrangement is of particular utility in a machine generally as those already described but intended for use as a compressor.

FIGS. 5 and 6 show schematically the positions of inlet and outlet openings respectively in the bores of a four-chambered compressor of the same general form as the engine embodiments of FIGS. 1 and 2. The relative positions of sleeve ports, intended to be opened and closed by rotation relative to the openings as mentioned immediately above, are also shown in FIGS. 5 and 6. For convenience, the ports in all sleeves in each Figure are represented as if on a common plane normal to the axis of the respective rotor.

In FIG. 5, sleeves 50 are carried in bores 51 formed in a rotor (not shown). Openings in the peripheral surfaces of the bores directed inwardly towards the rotor axis are arranged to connect with suitable gas inlet passages (not shown). The circumferential extremities of the openings subtend an angle B at the center of each bore, the said openings being equally disposed about diameters of the bores which coincide with a diameter of the rotor. Peripherally-extending ports in the sleeves each subtend an angle A at the center of the respective sleeve, the orientation of the ports in space being the same in all of the sleeves (i.e. the sleeves are all identical and each is maintained in the same alignment relative to the others by virtue of its cranked configuration and the inclination of the rotors in which its opposite arms are carried, as previously described). The bores 51 rotate around the sleeves 50 as the rotor revolves (the direction of rotor rotation being indicated by the arrow P) and the openings in the bores consequently pass over the sleeve ports, the interiors of the sleeves being thus in communication with the gas inlet passages from the instant the trailing edge of the bore opening passes over the leading extremity of a port until the leading edge of the opening passes over the other extremity of the port. This occurs between circumferentially-spaced points (indicated at K and L in FIG. 5) during which time a sleeve has respectively just passed its'axially innermost position relative to its bore and just passed its axially outermost position. Over this phase of rotation, the volume within a sleeve increases from a near minimum to a maximum and begins to decrease. Thus, inlet gas will flow into the sleeve. Further rotation of the rotor will result in a reduction in the volume within the sleeve whereby the gas will be compressed.

FIG. 6 shows the cycle of events occurring at the outlet ports, the sleeves 50 being the same as in FIG. 5 but carried, this time at their opposite ends, in bores 52 formed in another rotor (not shown) which rotates in phase with that of FIG. 5, its axis being of course inclined to that of the latter.

Openings in the bores 52 are arranged to connect with gas outlet passages (not shown), the opening and ports in the sleeves being of less peripheral extent than in the case of those last described, subtending angles of a and respectively. The disposition of the openings in the bores and of the ports relative to the rotor axis are also difi'erent. As a result, the unmasking of the ports, whereby the interiors of the sleeves are connected to the gas outlet passages, is in difi'erent sequence. This occurs between circumferentially-spaced points of rotor movement (indicated at M and N in FIG. 6) during which time the volume within a sleeve is decreased towards a minimum. Gas compressed as described with respect to FIG. will pass out of the cylinder at this stage.

In other variations (not illustrated), rotor and sleeve arrangements generally as in the foregoing embodiments provide self-contained power/compressor or power/transmission combinations. Suitable ports and fluid connections can be provided whereby alternate bores and their associated sleeves act, for instance, as internal combustion cylinders and as compressors. The compressor bores, which may be larger than those of the internal combustion cylinders, can be connected to the said cylinders for air charging. in other combinations, alternate bores form parts of hydraulic motors and compressors respectively, or of internal combustion engines and hydrostatic transmission units.

lclaim:

l. A rotary machine comprising a pair of rotors mounted for rotation about coplanar axes which are inclined relative to each other at an angle between 90 and something less than 180, a plurality of bores formed in each rotor and disposed regularly about the respective axis, sleeves extending between bores in one rotor and bores in the other rotor, each sleeve being cranked intermediately of its ends to the angle between the rotor axes, the opposite ends of the sleeves being disposed in individual bores of the rotors and the sleeves being capable of sliding and rotating relative to the bores, substantially cylindrical members extending axially within the sleeves from the ends thereof and fixed relative to the respective rotors, one bore of each linked pair of bores arranged to communicate with a gas inlet, at least one port formed in each sleeve and disposed so as to connect the gas inlet with the interior of the sleeve cyclically in accordance with rotation of the rotors, and at least one other port formed in each sleeve and arranged to connect the interior of the sleeve with a gas outlet.

2. A rotary machine according to claim 1 in which each bore in one rotor communicates with a gas inlet and each bore in the other rotor communicates with a gas outlet.

3. A rotary machine according to claim 1 having at least one port in each sleeve shaped and arranged to connect the interior of the sleeve with a gas inlet by relative rotary displacement of the sleeve and a rotor bore.

4. A rotary machine according to claim 1 having sleeves formed with ports which are shaped and arranged to connect the interiors of the sleeves with a gas outlet substantially by relative axial displacement between the sleeves and the appropriate cylindrical members.

5. A rotary machine according to claim 1 in which the angle between the rotor axes is of the order of 6. A rotary piston machine according to claim 1 further comprising independent means for constraining the rotors to rotate in phase with each other.

7. A rotary machine according to claim 6 having intermeshing gear teeth formed on the rotors.

8. A rotary machine according to claim 1 further comprising power transmission means operatively connected to at least one of the rotors.

9. A rotary machine according to claim 8 in which the power transmission means comprises toothed gearing.

10. A rotary machine according to claim I in which the rotors are mounted for rotation on fixed shafts and ducts are provided within the shafts to conduct gas to the gas inlets and from the gas outlets.

l l. A rotary machine according to claim 1 further comprising ignition means arranged to initiate combustion of gas within at least some of the sleeves.

12. A rotary machine according to claim 1 comprising gas inlet ports of substantially trapezoidal shape.

Patent Citations
Cited PatentFiling datePublication dateApplicantTitle
US2398690 *Sep 13, 1944Apr 16, 1946Saurer Ag AdolphInternal-combustion engine
US2622567 *May 26, 1951Dec 23, 1952Emile Myard FrancisRotatable piston machine
US3040664 *Apr 13, 1959Jun 26, 1962Flo Motive CorpDual cavity fluid handling device
US3304923 *Sep 2, 1964Feb 21, 1967Parenti Joseph SEngine
US3333577 *Mar 22, 1965Aug 1, 1967Pietro MongitoreRotary engine
Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US5159902 *Dec 31, 1990Nov 3, 1992Grimm C LouisRotary vee engine with through-piston induction
US6662775Oct 2, 2002Dec 16, 2003Thomas Engine Company, LlcIntegral air compressor for boost air in barrel engine
US6698394Oct 30, 2001Mar 2, 2004Thomas Engine CompanyHomogenous charge compression ignition and barrel engines
US8046299Jan 12, 2004Oct 25, 2011American Express Travel Related Services Company, Inc.Systems, methods, and devices for selling transaction accounts
Classifications
U.S. Classification123/43.00A, 91/500, 418/195
International ClassificationF01B3/00
Cooperative ClassificationF01B3/0035, F01B3/0038, F01B3/0052
European ClassificationF01B3/00B4C, F01B3/00B2B, F01B3/00B2