US3762844A - Positive displacement rotary heat engine - Google Patents

Abstract

The principle of compression in positive displacement, rotary engines working on basis of sinusoidal, cyclic compression and expansion is described and illustrated. The principle is adapted to a new, continuous flow, four stroke internal combustion engine with all compression-expansion chambers performing complete thermodynamic cycles for every revolution of the rotor, and the internal gas pressures acting on differential vane areas dynamically exposed during rotation. The operation of the engine is described, and estimated engine performances are shown for comparison with representative heat engines.

Description

United States Patent [1 1 Isaksen 1 1 Oct. 2, 1973 1 1 POSITIVE DISPLACEMENT, ROTARY,

HEAT ENGINE [76] Inventor: Kjell Isaksen, 9737 SE. 41 St.,

Mercer Island, Wash.

[22] Filed: May 12, 1970 [21] Appl. No.: 36,605

[52] U.S. C1 418/218, 418/219, 123/835 [51] Int. Cl...... F0lc 1/00, F02b 53/08, F04c 17/00 [58] Field of Search 418/228, 229, 230,

[56] References Cited UNITED STATES PATENTS 2,154,457 4/1939 Knapp; 418/218 969,353 9/1910 Evans... 418/217 X 1,464,408 8/1923 Collier.. 418/219 X 1,686,767 10/1928 Saxon 418/219 X 2,436,285 2/1948 Booth 418/219 2,582,413 1/1952 Clark 418/153 2,724,369 11/1955 Barrett 418/211 X 3,240,189 3/1966 Stumpfig.. 123/809 X 3,472,210 1 10/1969 Savoie 418/191 X FOREIGN PATENTS OR APPLICATIONS 599,609 3/1948 Great Britain 418/218 388,781 6/1965 Switzerland..... 893,197 1/1944 France 27,071 11/1912 Great Britain 123/835 Primary Examiner-Carlton R. Croyle Assistant Examiner-John J. Vrablik [57] ABSTRACT The principle of compression in positive displacement, rotary engines working on basis of sinusoidal, cyclic compression and expansion is described and illustrated. The principle is adapted to a new, continuous flow, four stroke internal combustion engine with all compression-expansion chambers performing complete thermodynamic cycles for every revolution of the rotor, and the internal gas pressures acting on differential vane areas dynamically exposed during rotation. The operation of the engine is described, and estimated engine performances are shown for comparison with representative heat engines.

5 Claims, 24 Drawing Figures PATENTED UB1 2 75 SHEET 2 BF 9 i REFERENCE lead.

trail.

REVOLVING A'NGLE- e b mm wDUmOP INVENTOR M Fig. 6

PAIENTED 013T 2I973 SHEET I [If 9 Fig.21A

IGNITION PATENTEUUET 21975 3.762.844

' SHEETSUF 9 INVENTOR zw wu PATENTEDBBT 2 3,762,844

SHEET 8 OF 9 64 FUEL INJECTlON 64 72 FUEL INJECTION SHEET 7 [IF 9 PATENTEUUCT 2 1973 PATENTEB 2|973 9 1 10 cu.in.

Compress. Ratio: Swept Volume: Leak. Area 0.024 sq.in.

EST. COMPRESSION PERFORM.

SHEET 8 BF 9 ["Le m;

1 mDmwmmm CORR. RPM x 1000 Fig.19

.m U %c 2 2 1 mm V mm e .w S M 1T TI 5 E Rich mixture combust- CORR- RPM X 1000 Fig. 20

POSITIVE DISPLACEMENT, ROTARY, HEAT ENGlNlE CROSS-REFERENCES TO RELATED APPLICA- TlONS This patent application makes no cross-reference to any prior, pending application.

BACKGROUND OF THE lNVENTlON This invention relates to a continuous flow, controlled compression ratio, positive displacement, rotary, heat engine with radial vanes rotating with a centrally shafted, radial, wheel type rotor and forming a plurality of compression-expansion chambers in annuli between the rotor and the rotor housings, while moving in wiping contact with sinusoidal walls of said housings, thereby exchanging torque about the shaft of rotation when internal gas pressures act on differential vane areas dynamically exposed.

The positive displacement, rotary, heat engine as described in this specification is related in art to gas turbines, to reciprocating piston engines and to a variety of rotary engines. The rotary movement and the continuous gas flow concept is inherited from gas turbines, where gas pressures also are acting on rotor blades to produce torque about a central, rotating shaft. The positive displacement concept originated in the reciprocating piston engines, where higher compression pressures were easier attained than in the compressors of the gas turbines.

The double or multi-stage gas turbines are severely limited in utility by the operating characteristic of their compressors with their difficulties in producing high pressures of compression over any appreciable range of rotor speeds, and the difficulties involved in matching teh compressors with the driving turbines. Difficulties experienced in cooling of the turbines have limited the operating temperatures in spite of introduction of costly turbine materials. The gas turbines can therefore not be utilized to the full extent of their thermodynamic posibilities. High costs and poor preformance have limited the general acceptance of the gas turbines.

In engines with reciprocating pistons acting on captive masses of gas, the heavy reciprocating parts severely limit the engine speeds and accelerations. These engines are excessively heavy, requiring expensive parts and regular maintenance. The combustion in these engines is generally very poor due to uneven distribution of fuels, and the exhaust gas emission of unburnt fuels and carbon monoxide is normally excessive. Four stroke operations are obtained through use of valve arrangements, which severely limit the flow and thereby the breathing capabilities of these engines.

Rotary engines generally fall in one of the following groups:

- derivates and improvements of the Wankel engine;

derivatives and improvements of the Cooley engine;

- derivatives and improvements of the two stroke,

eccentric, vane pump; and

other rotary engines.

In a Wankel engine a triangular piston rotates in an oscillatory path inside a housing of elliptical shape. The movements of the rotating piston are controlled by an eccentric shaft geared to the piston. Compressionexpansion chambers are formed betwen the housing and the triangular piston. This engine has until recently been a very poor performer and is still inferior to a good reciprocating piston engine. The lack of performance has frequently been concealed by deceptive information regarding the engine swept volume, which often reflected only one third of the real air breathing capacity. The engine has been plagued with technical problems, some of which are inherent in the basic design. The engine lacks growth potential in its basic design and is therefore often found in compounded versions. The Wankel engine so far has been unable to satisfy the high expectations advocated at its conception, and it has so far made very limited impact on the engine market.

The original Cooley steam engine consisted of a two lobe, oval rotor moving in an oscillatory, rotational path inside a three clover leaf shaped housing. Later models have used four lobes and five leaves. The compression-expansion chambers in this engine do not move with the rotor as in the Wankel engine, and a valve arrangement is essential. This engine is later adapted to the four stroke cycles of an internal combustion engine in a variety of arrangements. This engine has also been a very poor performer and has suffered from similar technical difficulties as the Wankel engine. No such engine is known to have been marketed.

In an eccentric vane pump the circular rotor mounted with radial vanes rotates inside an eccentrically located, circular housing. This engine is restricted to two stroke operations and is dependent on an external source for charging the compression-expansion chambers. The engine appears in a great variety of embodiments, none of which are known to have been marketed.

A great array of geometrical engine designs of rotary types have been presented during the years. Most of these engines have never found any acceptance. Rotary piston engines with cylinders and pistons rotating around a stationary shaft have been in use on early aircrafts, but these were later discontinued due to adverse operating characteristics.

Engines in use at this time are all seriously limited in their application resulting from their operating characteristics or the execution of their embodiments. A valid justification for pursuing a new design under the present conditions must be founded on the anticipation of gaining an advantage in one or more of the following main directions:

improvement in performance characteristics; reduction in weight; reduction in costs; and reduction in harmful air pollutants; over and above known engines representing the state of the art. The invention described in this specification is expected to show significant improvements in all of the stated directions.

SUMMARY The compression concept to be shown in this specification was developed to form the theoretical basis for a new type of heat engine intended to combine the best features of the conventional positive displacement, reciprocating piston engines and the continuous flow gas turbines.

In the engine described and shown, four stroke operating cycles can be completed in a series of compression-expansion chambers during every completed rotation of the chambers in a single stage about the rotor shaft. All requirements for valve arrangements as in the reciprocating piston engines and all requirements for separation of compression and expansion into different stages as in the gas turbines are eliminated. Flow of fluids into the engine through the intake ports and out of the exhaust ports are continuous.

Fluid compression is achieved by confining a volume of fluid within an enclosed chamber, which is then compressed in volume during a rotary movement of the rotor with respect to the rotor housing. The compression is controlled by relieving the fluid in the compressing chamber into a second, expanding chamber, thereby creating a flow of fluids between the two chambers. In the case of an internal combustion engine, combustion will take place during this cross-over flow, injecting burning gases into the expanding chamber. The pressures in the compressing chamber will always exceed those in the expanding chamber in order to maintain fluid flow. The exposed differential work area will, however, soon become positive in value and positive work will result.

The internal fluid pressures act on differentially exposed vane areas during the chamber rotation about the rotor shaft. The exposed areas are negative during compression and positive during expansion.

Fluid induction, exhaust and ignition, as applicable, form integrated functions in the thermodynamic cycles.

The illustrated embodiment of an internal combustion engine is selected from a number of possible embodiments all belonging to the same family and described by the same basic operating principles. The engine size and the compression ratio was selected especially to compare with conventional positive displacement, reciprocating piston engines in common use.

The performance of the shown engine has been analyzed in a conservative manner and is illustrated in the figures for comparative purposes. The shown performance curves are limited to the shown embodiment and the stated engine displacement. Since a variety of embodiments is possible within the operating prinicples, special performance characteristics may easily be developed for special requirements. No statement in this specification shall thus be construed to mean that an optimum engine embodiment is shown.

The invention is limited in scope by the described prinicple(s) of compression and by the stated claims essentially as described in this specification and supported by the attachedillustrations.

The shown embodiment of the internal combustion engine has been designed with the object of securing the following relative advantages:

- a high power/weight ratio;

a high. specific power output;

an extended operating speed range;

an improved operating economy;

a rotor assembly in complete balance;

a compact packaging;

a simplified mechanical operation;

a simplified embodiment;

a simplified accessory requirement;

a full volume, forward flow combustion; a controlled heat transfer rate;

reduced port losses;

complete rotor burst containment capability; reduced tooling requirement; reduced material costs; and

with extensive growth potential all in comparison with the known state of the art in positive displacement, reciprocating piston engine technology as far as comparison is possible.

For the purpose of easing comparison with the reciprocating piston engines in use, the illustrated internal combustion engine embodiment performances will be based on rich mixture combustion with fuel introduced in conventional manner through a carburettor or injected into the intake manifold through a metering systern. This method of combustion will be referred to as full volume combustion as compared to partial volume or partial flow combustions as described later .in this specification.

BRIEF DESCRIPTION OF THE DRAWINGS The following illustrations are offered in support of this specification to describe the basic compression principlc(s) involved and the mechanical embodiment with an analysis of the engine performances during operations:

FIG. 1 illustrates the basic prinicple of compression in the simplest two-stroke concept;

FIG. 2 illustrates the basic principle of compression as applied to the basic four-stroke concept;

FIG. 3 illustrates the relationship between the compression-expansion chamber sector angle and the compression ratio for the cases of FIG. 1 and FIG. 2

FIG. 4 illustrates the boundaries of a compression chamber in elevation and defines the reference line;

FIG. 5 illustrates the volume variations of a compression-expansion chamber as the rotor revolves about the rotor shaft for the case identified in FIG. 3;

FIG. 6 illustrates the differentially exposed vane areas as the compression-expansion chamber of FIG. 5 and the rotor revolves about the rotor shaft;

FIG. 7 illustrates the four-stroke, internal combustion engine embodiment in exterior side elevation;

FIG. 8 illustrates the same engine embodiment in a transverse section along line AA of FIG. 7 and FIG.

FIG. 9 illustrates the engine embodiment in exterior side elevation viewed from the opposite side to FIG. 7;

FIG. 10 illustrates one of the two profile housings in exterior elevation and with the face plate removed;

FIG. 11 illustrates a section of FIG. 10 and FIG. 12-

along the line 8-3;

FIG. 12 illustrates the profile housing in interior elevation;

FIG. 13 illustrates one of the two identical rotor halves as viewed from the oposing rotor half;

FIG. 14 illustrates a section of the same rotor half along line C-C of FIGS. 13 and 15;

FIG. 15 illustrates the above rotor half as viewed from the mating profile housing;

FIG. 16 illustrates details of a rotor vane;

FIG. 17 illustrates the exposed torque areas of two adjacent rotor vanes enclosing a compressionexpansion chamber;

FIG. 18 illustrates the maximum movements of a rotor vane in the rotor and locates the vane center of gravity;

FIG. 19 illustrates the mean pressure of compression of the divided compression-expansion chamber as its centerline passes the top dead center; and

FIG. illustrates the estimated net torque and the brake horse power of the engine as related to rotor speeds.

FIG. 21A shows a schematic of a circumferential view through the compression-expansion chambers while in suction and compression modes of the cyclic thermodynamic four stroke operation.

FIG. 21B shows a schematic of a circumferential view through the compression-expansion chambers while in ignition, expansion and exhaust modes of the cyclic thermodynamic four stroke operation.

FIG. 22 shows a schematic of a circumferential view through the compression-expansion chambers while in compression, ignition and expansion modes of a four stroke thermodynamic cycle operation involving fuel injection into a moving enclosed airstream under compression and/or expansion followed by auto-ignition of the mixture.

FIG. 23 shows a schematic of a circumferential view through the compression-expansion chambers while in compression, ignition and expansion modes of a four stroke thermodynamic cycle operation involving fuel injection into a moving enclosed airstream under compression and/or expansion followed by ignition of the mixture by and external source.

DESCRIPTION OF THE PREFERRED EMBODIMENT The invention and the preferred embodiment is based on the following principle of compression:

In FIG. 1 a sine curve is shown with its axis displaced by a full amplitude or radius to become a tangent to the minimum point on the curve. The area confined by the sine curve, the displaced axis and a defined angular span described as a constant fraction of the circumference or the revolving angle, varies with the location of this angular span around the circumference as shown in two positions identified by shaded boundaries in FIG. 1 and FIG. 2. The ratio of the maximum to the minimum areas confined may be defined as the area compression ratio. The compression ratio is controlled by the magnitude of the fraction of the angular span to the circumference or the revolving angle and is independent of the sine curve amplitude or radius (r) or by the magnitude of the radius of rotation as long as these quantities remain constant around the circumference. Since the area compression ratio is independent of the radius of rotation, the area compression ratio must be identical to the volume compression ratio for all cases also confined between any two radii of rotation.

FIG. 2 shows a two cycle sine curve confining compression areas in the same manner as shown in FIG. 1.

In FIG. 3 the compression ratios of both single and double cycle sine curve operations are shown as functions of the angular span or sector angle. It is seen from this figure that decreasing angular span or sector angle results in increasing compression ratio.

The illustrated principle of compression is different from the principle of compression employed in rotary pumps by the fact that means of controlling the compression ratio is available, while in pumps the compression ratio invariably approaches infinity as the minimum volume approaches zero. The illustrated principle of controlled compression ratio has no meaning in rotary engines, however, unless the divided minimum volumes are interconnected by a flow passage. Overcompression of the fluids in the trailing, compressing part of the chamber is thereby relieved by flow into the leading, expanding part of the chamber, and a compromise compression ratio and thereby a fluid pressure level is reached. When no such passage is available, a compression ratio of infinity is approached as the part of the chamber on the trailing side of the compression peak approaches zero. The presence of such a flow passage, regardless of form, tends to reduce the compression ratios below the theoretical values shown in FIG.

Rotary engines employing single cycle sine curve contours are limited to two-stroke thermodynamic operations, while double sine curve contours will accept both two and four-stroke thermodynamic operational processes. From the magnitudes of the compression ratios for the single and the double cycle sine curves as shown in FIG. 3, it is concluded that the shown principle of compression for the stated curves may be applied in practical thermodynamic process operations in the manner shown in FIGS. 1 and 2. It can be shown that triple and quadra cycle sine curves require increasingly more sector divisions or smaller sector angles to secure practical compression ratios, thus reducing the prospects for adaption in simple embodiments.

From the possible solutions shown in FIG. 3, the special case marked 0 on the double cycle sine curve re lationship between sector angles and compression ratios is selected for embodiment as a four stroke, positive displacement, internal flow combustion, heat engine.

The compression-expansion chamber is illustrated in FIG. 4 as shown in elevation. The reference line is drawn on this figure for the purpose of positive identification of location of the chamber in later illustrations. The single compression-expansion chamber in the fig ure is enclosed by two concentric walls in axial direction at the two radii of revolution R, and R by a wall in the plane of the paper illustrating the displaced axis, by a concentric double cycle sinusoidal plane facing the paper and by two sector walls of variable areas in an axial plane to complete the enclosure of the chamber. By introducing a relative motion between the sinusoidal wall and the remaining walls in a circular direction, the two sector walls will vary in exposed areas in accordance with the respective distances between the sinusoidal plane and the displaced axis plane. Any pressure arising inside the compression-expansion chamber enclosed by the stated walls will create forces on all the enclosing walls. Selecting the sinusoidal wall for a stator and allowing the remaining walls to rotate about the central axis, the internal chamber pressures create torque about this axis in accordance with the exposed areas of the sector walls and their arms with respect to the axis. When the exposed sector wall A is greater than the exposed wall A a positive torque is created about the axis of rotation. When the magnitudes are reversed, the torque about the axis is negative, and when the two areas are equal, the chamber is at a dead center.

FIGS. 5 and 6 illustrate the magnitude of the compression-expansion chamber volume and the differential vane or torque areas in percents as functions of revolving angular locations of any one of the six chambers about the rotor shaft with respect to their reference lines. By cross-relating the two figures and applying the above sign convention to the differentially exposed sector wall areas defined as torque area, (A

A it is easily seen that compression requires work, while expansion gives up work.

The embodiment of the stated internal combustion engine is shown in FIGS. 7 through 18 attached to this specification.

Referring to the figures, FIGS. 7, 8 and 9, the engine power unit comprises a rotor assembly splined and shrunk onto a shaft 31 and free to rotate in two double bearings 32 and 33. The bearings 32 and 33 are housed in two opposing profile houses 34 and 35. The two profile houses 34 and 35 enclose the rotor assembly 30 and are kept in a fixed relationship by means of an oil col lector ring 36 serving as a spacer. The exteriors of the two profile houses 34 and 35 are covered by two face plates 37 and 38 to enclose the coolant passages 39 and 40 and the lubrication annuli 41 and 42 in the profile houses 34 and 35. The face plates 37 and 38 and the profile houses 34 and 35 are penetrated by two inlet ports 43 and 44, two exhaust ports 45 and 46 and two ignitor plug holes 47 and 48. The face plates 37 and 38 are also penetrated by two coolant inlet holes 49 and 50, two coolant outlet holes 51 and 52, two lubrication inlet holes 53 and 54 and two lubrication overflow holes 55 and 56.

The power unit is held together and mounted by fasteners through the holes 57 and 58. The face plates 37 and 38 are also fastened to the profile houses 34 and 35 through the attachment holes 59 and 60 and also 61 and 62. Provisions are made to attach an oil sump at the surface 63.

Six vanes 64 penetrate the rotor discs 65 and 66 through six radial slots 67 as shown in FIGS. 8, 13, 15, 16, 17 and 18. The vanes 64 are supported by pivots 68 at the shaft end of the rotor assembly 30 in the rotor hubs 70 and 71. The rotor vanes 64 move in soft or elastically loaded wiping contact with the sinusoidal walls 72 and 73 of the profile houses 34 and 35 through a set of soft loaded sealing strips 74 at the edges of the rotor vanes 64.

Twelve compression-expansion chambers, six on each side of the rotor assembly 30, are formed between the two opposing profile houses 34 and 35, the rotor discs 65 and 66, the rotor vanes 64, the rotor hubs 70 and 71 and the outer rotor rims 75 and 76. The sinusoidal walls 72 and 73 of the profile houses 34 and 35 are off-set by half a sinusoidal wave length to allow room for the rotor vanes 64.

The inlet ports 43 and 44 and the exhaust ports 45 and 46 of the profile houses 34 and 35 are inclined to reduce the turning angle of the incoming and exhausting gases as seen from FIGS. 7, 9 and 10. Both the inlet ports 43 and 44 and the exhaust ports 45 and 46 are so located that they provide the correct timing for the thermodynamic cycles as shown in FIG. 12. The four engine strokes shown in this figure are referred to the reference line previously identified. Sector I-I indicate the suction stroke, sector IIII the compression stroke, sector III-III the power expansion stroke and sector IV-IV the exhaust stroke. Port overlap is indicated by sector VV.

Passages 77, shown in FIG. 15, are cut in the rotor discs 65 and 66 to provide connection between the two divided parts of the compression-expansion chambers over the compression top in addition to the flow passages created by the running clearances. These passages 77 must be sized to prevent any significant pressure backup in the trailing part of the compressionexpansion chamber and to secure the correct flow ve-,

locities to the compressed gas during the passage. Excessive sizing of the flow passage 77 will reduce the compression ratio needlessly. The igniter plug holes 47 and 48 are positioned near the top dead center to coincide with the flow passages 77. ignition of the flowing gas stream can in this manner be timed well in advance of the compression pressure peak.

Gas seals are provided at the rotor vane edges 74, in recesses '78 in the rotor disc slots 67 and in the groves 79 and 80 at the edges of the sinusoidal walls 72 and 73 of the profile houses 34 and 35. These seals are introduced to improve the low rotor speed operating characteristics. The seals are designed and loaded to provide maximum sealing for minimum friction losses. Details of the pneumatically loaded seals are not shown.

Two separate cooling systems are used due to the differences in cooling requirements and the need for return of the lubricating oils.

The sinusoidal walls 72 and 73 of the profile houses 34 and 35 are cooled by liquid coolant entering the cooling passages 39 and 40, as shown in FIGS. 7, 9 and 10, through holes 49 and 50, counter-flowing the internal gas flow direction in the compression-expansion chambers and leaving through the holes 51 and 52.

Cooling of the rotor assembly 30 is afforded by means of lubricating oils entering the oil annuli 41 and 42 at the holes 53 and 54. The oil then enters the main shaft 31 through the double bearings 32 and 33 and passes into the rotor assembly 30 at the pivot gallery 81, as shown in FIG. 13, for lubrication of the pivot bushings and the pivot pins 69 and 68. From the pivot gallery 81 the oil distributes into radial cooling passages 82 between the two halves of the rotor assembly 30. The cooling passages 82 are throttled at the radial exits 83 to provide better contact between the oil and the rotor disc walls 65' and 66 during heat transfer and to reduce the cooling flow rate at higher rotor speeds. The ejected oil is then collected in the collector ring 36 and is drained back into the oil sump after some splashback against the outer rotor rims 75 and 76. Oil cooling takes place in the oil collector ring 36 and in the oil sump.

MODES or OPERATION The engine operations are shown in FIGS. 21A, 21B, 22 and 23 and are described as follows:

When a leading vane 64 of a compression-expansion chamber passes the compressing peak of the sinusoidal wall 72 or 73 between the exhaust 45 or 46 and the intake ports 43 or 44, a suction is created in the expanding part of the chamber. Air or air/fuel mixture is drawn into the expanding chamber through the intake port at moderate velocities. As soon as the chamber volume has reached its maximum and the trailing vane 64 of the chamber passes the end of the intake port, 43 or 44, the compression stroke commences.

As soon as during continued rotor movement the leading vane 64 of the compression-expansion chamber reaches the compression peak at the igniter plug location in hole 47, the compression-expansion chamber divides into two parts. The leading part of the chamber expands in volume at an initially slow rate, while the trailing part compresses at a higher rate. The pressure differential thereby created between the two parts of the chamber causes a flow of fluid from the trailing and into the leading part of the chamber through the flow passage 77 at a velocity higher than the rotor speed. At a timed position the flow is ignited, and a flame front propagates into the combustible mixtures at increasing velocities relative to the advancing local flow velocity. After a relatively slow start the pressures in the compression-expansion chamber build up in both the divided parts. The flow between the two parts of the chamber continues during the pressure increase, while the flow velocities and the flow densities adjust to the new conditions. The flow combustion continues until either fuel or air supply is exhausted.

During the power-expansion stroke the leading vane 64 increases in exposed area until a maximum is reached. The trailing vane 64 area exposed during this stroke reduces in magnitude until it reches the minimum size as it passes the sinusoidal wall compression peak at the igniter. The differential vane areas are therefore positive. The pressure in the compressionexpansion chamber reaches its maximum value during combustion a short time after the minimum chamber volume has passed and the volume is increasing and then reduces sharply during the chamber expansion as the positive value of the differentially exposed vane or torque area increases.

As soon as the leading vane 64 of the compressionexpansion chamber reaches the exhaust port 45 or 46, the remaining chamber pressure collapses as the exhaust gases expand into atmosphere. The collapse rate and thereby the exhaust noise can to some degree be controlled by the location and shape of this exhaust port.

The scavenging of the compression-expansion chamber is assisted by the chamber compression following the expansion stroke. The port overlap in this engine has a difierent meaning than in a reciprocating piston engine. The port overlap in this engine improves the scavenging, but since the chamber is divided into two parts with only a small passage 77, the inertia effect of the exhaust gases to improve the inflow of new gases through the intake port 43 or 44 is of reduced value. The continuous inflow through the intake ports 43 or 44 will in any case secure the best possible filling of the chambers.

The process cycle then repeats itself.

Twelve identical process cycles are executed for every revolution of the engine rotor, six on each side of the rotor assembly. The shown embodiment will execute a power stroke for every 30 movement of the rotor. The power strokes are thereby overlapping, and the power fluctuations are consequently reduced to a minimum. The moment of inertia created by the rotor itself is therefore adequate to secure an even, continuous rotor rotation, and no requirement for any flywheel exists. A reciprocating piston engine with four stroke operation will require 24 cylinders to execute the same number of power strokes per shaft revolution.

The shown embodiment indicate that fuel/air mixing takes place before the combustible mixtures enter the compression-expansion chambers. In this manner the fuel/air mixtures fill the entire compression-expansion chamber before the compression, and an essentially full volume combustion will take place. Care must be taken in sizing the flow passage '77 to avoid overcompression of the combustible mixtures in the trailing part of the compression-expansion chamber at higher rotor speeds, which otherwise may cause detonations. The

flow type combustion is a considerable improvement over combustion by flame front propagation in a relatively stagnant combustible mixture, as is the case in reciprocating piston engines. Identical distribution of fuel/air mixtures to all compression-expansion chambers makes lean mixture combustion possible with improved fuel economy and reduced amounts of unused fuels and carbon monoxide in the exhaust gases with no apparent loss in power.

A different method of combustion may be utilized when fuel is injected directly into the compressionexpansion chamber during the compression and after a part of the chamber has passed the compression peak and is under expansion on the leading side. The chamber will then receive fuel injection into the trailing part of the chamber only. When this combustible mixture is forced through the flow passage 77 and ignited, a burning jet of fuel/air will be injected into the leading part of the chamber containing air only. Excess fuels injected into the trailing, compressing part of the chamber will be burnt off in the excess air in the expanding part. Extremely low levels of carbon monoxide and unburnt fuels will result from this type of combustion. This method was previously referred to in this specification as the partial volume combustion.

A third method of combustion is also possible in this engine. This method was previously referred to as the partial flow combustion method. By special forming of the the flow passage 77 into different paths it is possible to divide the flow passing over the compression peak. Injecting fuel into the air passing through some of these passages for combustion, while letting the remaining flow paths by-pass the combustion, will make a partial flow combustion possible with a down-stream mixing of combusted and non-combustible gases with a possible burn-off of excess fuels during expansion. This method of combustion has the potential for complete combustion control leaving a minimum of unburnt fuels and carbon monoxide in the emitted exhaust gases. Combustion in excessive air as described in the preceeding combustion methods, will result in greatly improved fuel economy.

As seen from the above description the engine will accept a carburettor system, inlet port fuel injection and, with some modification to the shown embodiment, fuel injection directly into the compression-expansion chamber in at least the two previously described manners. Since the engine operates on continuous flow through the inlet ports, identical charges will reach every chamber on each side of the rotor. The colling system of the engine will also permit higher combustion temperatures than usual in reciprocating piston engines without any decay in the chamber wall metallurgical properties. it is therefore possible to operate this engine much closer to the best power fuel/air ratio than in previous positive displacement engines. A fully lean mixture operation is proposed with fuel enrichment towards best power mixtures at higher rotor speeds and torque requirements, instead of the present rich mixture operation with further fuel enrichment under more demanding operations. Two point fuel injection can be provided for best possible fuel control at moderate costs.

The shown engine embodiment will accept a conventional type circuit breaker in the ignition system geared to give two series of six sparks per rotor revolution. A more advanced electronic system may be required to enable engine utility at higher rotor speeds. No distributor is required for this system. Two ignition coils may instead be used in conjunction with the spark generator and the sparking plugs.

Transmission of torque from the vanes 64 follow the load path through the pivots 68 and bushings 69 to the rotor assembly 30 and the shaft 31. The gas seals recessed in groove 78 in the disc slots 67 are elastically loaded and do not transmit any appreciable load directly to the rotor 30. The torque loads experienced by the pivot pins 68 are thus superimposed on the radial loads caused by rotation. Since the vanes 64 are soquentially loaded on either side, the two sides of the vane pivot pin 68 will be subjected to an alternating load pattern.

The vanes 64 are made with hollow core to maintain a center of gravity close to the pivot 68 as shown in FIG. 18 This close center of gravity position reduces the vane loads on the sinusoidal walls 72 and 73 and improve the rotor speed range.

The working temperatures of the rotor assembly 30 and the sinusoidal walls 72 and 73 must be controlled to levels below their metallurgical limits and also below the critical temperature for ignition of the combustible mixtures prematurely. The wall temperatures are controlled by creating a temperature gradient over the walls by means of heat transfer to a flowing heat transfer fluid or coolant on the outside of the wall. The cyclic heat transfer losses are greatest at lower rotor speed when the combustion temperatures are lower and the time exposures are longer during each cycle. At higher rotor speeds the combustion temperatures stabilize, and the heat transfer rate at shorter time exposure becomes constant in each operating sector reducing the requirement for increasing coolant flow rates. Cooling of the rotor assembly 30 is afforded by returning the lubricating oil through the core of the rotor discs 65 and 66 on its return to the oil sump. This cooling requirement is minor due to the rotor exposure to the cooler inflowing gases during its rotation. The cooling requirements for the sinusoidal walls 72 and 73 vary around the circumference. The greatest cooling requirement exist in the combustion sector of the wall, which is continuously exposed to the high temperatures of combustion. The cooling fluid is therefore pumped in near the exhaust ports 45 and 46 and out near the intake ports 43 and 44 to prevent warping of the structure and reduce the wall temperatures. The rotor vanes 64 are cooled by conducting heat into the adjoining cooled members of the structure.

Fuel requirements for the engine varies with the method adapted for introduction of the fuel and the compression ratio selected. Normal octane value automotive gasoline may be used in conjunction with a carburettor for the engine in shown embodiment. The octane value of the fuels may be substantially reduced when fuel is injected directly into the compressionexpansion chamber in the combustion sector in a manner designed to prevent compression ignition of the combustible mixtures out of sequence. Lower octane fuels emit in many cases less harmful pollutants and are more completely consumed during combustion.

FIG. 19 shows the mean pressure of compression capability of the described engine in a two liter displace ment version and operating under adverse leakage conditions. It is seen that as the rotor speed increases, the effect of leakage on the pressures of compression reduces. The shown pressures are mean values of the pressures existing in the leading and trailing part of the compression-expansion chamber. The pressure differential between the two chambers may not exceed two percent of the differential value for the shown compression ratio.

In FIG. 20 an estimate is shown of the torque and power performance of the described engine in a two liter version as function of rotor speeds. Since the estimate was based on adverse leakage conditions, the torque developed by the engine is relatively low at lower rotor speeds and peaks relatively late in the speed range. Allowances have been made in both curves for operation of essential accessories.

The specific power output of the shown two liter displacement version of the engine is approximately 2.5 B.H.P. per cubic inch displacement as compared with approximately 1.2 for a normal high performance automotive engine of similar displacement and even less for a Wankel engine. The estimated power/weight ratio depends heavily on the choice of materiels for the structure and also on accessories. A power/weight ratio of 2.9 B.H.P. per lb. weight seems reasonable for the power unit alone in the two liter displacement version, and a lower value with all accessories and in running condition. Comparable values for positive displacement engines in running condition do not exceed I.0 B.I-I.P. per lb. weight.

ALTERNATIVES The principle of compression as described in this specification is adaptable to a great variety of applications in compressible fluid machinery. The comparative advantage, however, seems to be connected to machinery using internal compression followed by expansion and preferably with addition of heat or other pressure increasing process occuring in conjunction with the two strokes.

FIG. 3 clearly shows the relationship between the division into compression-expansion chambers and the compression ratio. A departure from the true sine or cosine curve form will change this relationship. The expression sinusoidal wall in this specification is used in the mathematical sense within a band of reasonable tolerances to describe a wall of sinusoidal character regardless of the method used in ariving at this curve. This invention as expressed and shown covers all the indicated compression possibilities along the the curves in FIG. 3 for the type of structure shown in this specification.

The introduction of pivoted vanes is closely connected with requirement for high performance. Varies sliding in slots will perform adequately for many purposes where high performance is not required and the wear rate is not of great importance.

The double acting engine shown in this embodiment improves the packaging and the performance. The two sides operate thermodynamically independent. Alternative methods for securing continuous contact between the vanes and the sinusoidal wall may be used as application requires.

When the compression pressures become adequately high, compression ignition may be adapted instead of electrical ignition. Fuel injection directly into the compression-expansion chamber then becomes imperative.

The engine is growth limited by the relative Mach. no. between the rotor and the entering gas at the intake ports 43 and 44 and the ability to cool the sinusoidal walls of the profile houses 34 and 35 at higher engine speed operations. The shown two liter displacement engine operates so far from these critical limits that considerable growth may be possible before any reduction in rotational speeds becomes essential.

With reference to the preceeding specification supported by illustrations, I hereby respectfully request patent protection for the following claims:

1. A positive displacement, rotary, heat engine, comprising:

a wheel type, radial rotor 30 extending into an expanded rotor rim 75, securely attached to a shaft 31 at the center of the rotor hub 70 and mounted for rotation;

said rotor 30 facing a housing 34 extending radially outwards from said shaft 31 in spaced relation to said rotor 30;

said rotor 30 in conjunction with said housing 34 enclosing an annular chamber;

the inner surface of said housing 34 facing said annular chamber being contoured as a sinusoidal wall 72 and converging towards focal points in the proximity of said rotor shaft 31;

said chamber having means of flow passage 77 over the sinusoidal peaks indented into the face of the rotor disc 65;

said rotor 30 having a plurality of radial slots 67 disposed at equi-angular location, the exact number of said slots 67 being determined by the compression ratio requirement on basis of described mathematical principles; I

-- a rotor vane 64 free to move in axial direction in each slot 67 and having means to maintain edgewise wiping contact with said said sinusoidal wall 72;

said rotor 30, said sinusoidal wall 72 of said housing means 34 and said rotor vanes 64 defining a plurality of sinusoidal expanding and contracting annular sector chambers during rotation of said rotor 30 relative to said housing means 34;

- said rotor vanes 64 moving in wiping contact with said sinusoidal wall 72 being differently and dynamically exposed to fluid pressures in said sector chambers to negotiate torque about said shaft 31 of revolution during rotational movement of said rotor 30 relative to said housing means 34; I

-- said housing means 34 having intake 43 and exhaust 45 ports located for direct periodic communication into said sector chambers and extending circumferentially over predetermined sectors to control the process timing of predefined thermodynamic cycles as said chambers move rotarily relative to said housing means 34;

- said rotor 30 having forced radial fluid flow means 82 for internal cooling; and

said housing means 34 having forced fluid flow means for cooling said sinusoidal wall 72.

2. In an engine as described in 1.:

said rotor vanes 64 being pivotally mounted 68 in the hub 70 of said rotor 30 for transmittal of substantial loads through said pivots 68; and

- having elastically loaded sealing means in said rotor slots 67 and at said rotor vane 64 wiping edges 74.

3. in engines as described in 1.:

- means located in said sinusoidal wall 72 in said housing means 34 for sequential injection of fuels into flowing reacting agents in said sector chambers.

4. ln engines as described in 1.:

igniter means 47 located in said sinusoidal wall 72 of said housing means 34 and operable for sequential ignition of compressed combustible mixtures flowing through passages 77 in said sector chambers.

5. In engines as described in 1.:

means located in said sinusoidal wall 72 in said housing means 34 for sequential injection of fuels into flowing reacting agents in said sector chambers: and

igniter means 47 located in said sinusoidal wall 72 of said housing means 34 and operable for sequential ignition of compressed said fuels and reacting agent mixtures flowing through said passages 77 in said sector chambers.

Claims (5)

1. A positive displacement, rotary, heat engine, comprising: - a wheel type, radial rotor 30 extending into an expanded rotor rim 75, securely attached to a shaft 31 at the center of the rotor hub 70 and mounted for rotation; - said rotor 30 facing a housing 34 extending radially outwards from said shaft 31 in spaced relation to said rotor 30; - said rotor 30 in conjunction with said housing 34 enclosing an annular chamber; - the inner surface of said housing 34 facing said annular chamber being contoured as a sinusoidal wall 72 and converging towards focal points in the proximity of said rotor shaft 31; - said chamber having means of flow passage 77 over the sinusoidal peaks indented into the face of the rotor disc 65; - said rotor 30 having a plurality of radial slots 67 disposed at equi-angular location, the exact number of said slots 67 being determined by the compression ratio requirement on basis of described mathematical principles; - a rotor vane 64 free to move in axial direction in each slot 67 and having means to maintain edgewise wiping contact with said said sinusoidal wall 72; - said rotor 30, said sinusoidal wall 72 of said housing means 34 and said rotor vanes 64 defining a plurality of sinusoidal expanding and contracting annular sector chambers during rotation of said rotor 30 relative to said housing means 34; - said rotor vanes 64 moving in wiping contact with said sinusoidal wall 72 being differently and dynamically exposed to fluid pressures in said sector chambers to negotiate torque about said shaft 31 of revolution during rotational movement of said rotor 30 relative to said housing means 34; - said housing means 34 having intake 43 and exhaust 45 ports located for direct periodic communication into said sector chambers and extending circumferentially over predetermined sectors to control the process timing of predefined thermodynamic cycles as said chambers Move rotarily relative to said housing means 34; - said rotor 30 having forced radial fluid flow means 82 for internal cooling; and - said housing means 34 having forced fluid flow means for cooling said sinusoidal wall 72.

2. In an engine as described in 1.: - said rotor vanes 64 being pivotally mounted 68 in the hub 70 of said rotor 30 for transmittal of substantial loads through said pivots 68; and - having elastically loaded sealing means in said rotor slots 67 and at said rotor vane 64 wiping edges 74.

3. In engines as described in 1.: - means located in said sinusoidal wall 72 in said housing means 34 for sequential injection of fuels into flowing reacting agents in said sector chambers.

4. In engines as described in 1.: - igniter means 47 located in said sinusoidal wall 72 of said housing means 34 and operable for sequential ignition of compressed combustible mixtures flowing through passages 77 in said sector chambers.

5. In engines as described in 1.: - means located in said sinusoidal wall 72 in said housing means 34 for sequential injection of fuels into flowing reacting agents in said sector chambers: and - igniter means 47 located in said sinusoidal wall 72 of said housing means 34 and operable for sequential ignition of compressed said fuels and reacting agent mixtures flowing through said passages 77 in said sector chambers.

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