WO2007036955A1

WO2007036955A1 - Split cycle rotary internal combustion engine

Info

Publication number: WO2007036955A1
Application number: PCT/IN2006/000392
Authority: WO
Inventors: Jiban Jyoti Mistry
Original assignee: Seth, Chandan, Kumar; RAY, Unman
Priority date: 2005-09-30
Filing date: 2006-10-03
Publication date: 2007-04-05
Also published as: WO2007036955B1

Abstract

A split cycle rotary internal combustion engine having a first rotary configuration (C1) for executing the expansion and exhaust phases of a four phase engine cycle and a second rotary configuration (C2) for executing the intake and compression phases of a four phase engine cycle. Each of the first and second rotary configurations (C1, C2) has individual center shaft (44, 54) integrated with an eccentric lobe (45), individual outer body having a trochoidal inner chamber and an inner rotor body (40) for rotation within said inner chamber. Each rotor body has three apex portions including swivel apex seal arrangements (66) for keeping the seal elements (78, 79, 80) in perpendicular sealing relation with said trochoidal inner chamber peripheral wall (31). Air passages (99) for interconnecting said first and second rotary configurations to accomplish two successive compressed air deliveries in a single rotation of the center shaft (44)and the first rotary configuration execute two consecutive power impulses in the same rotation of said center shaft (44). Seal arrangements (81) for periodically dividing their adjacent working chambers into a leading and a trailing working chambers during a respective working chamber passing through its so called 'Top Center' position. In the first rotary configuration (C1), when combustion event is initiated within the forwarding portion the trailing portion is then near the conclusion of the exhaust phase of previous combustion. The engine having dual power transmission mode, of which the initial power transmission mode effectively starts before the respective rotor working face reaches to its top center position and continue during entire active combustion phase. Each of the first and the second rotary configuration (C1 , C2) includes a pair of internal gears (51a, 51b) fitted on both sides of respective rotor body, a pair of stationary external gears (52a, 52b) confined on two facing sidewalls (32, 33) of a outer body and being in mesh with said pair of internal gears (51a, 51b) for maintaining the relative position therebetween and also for transmitting driving force to center-shaft during said initial power transmission mode. Lubrication for all the internal seal elements is provided by means of internal lubrication oil.

Description

Split cycle rotary internal combustion engine

Technical field of the invention

The present invention relates to an internal combustion engine; more specifically to a rotary internal combustion engine. The present invention particularly relates to a split cycle rotary internal combustion engine.

Background Art

The basic concept of IC engine operation remains unchanged for more than a century. The well-known two-stroke and four-stroke reciprocating piston engines, both having two types, viz., spark ignition (SI) and compression ignition (CI). The four stroke engines are recognized as the prime mover in the ground transportation and many other industries. Modern piston engines adopted several modifications such as MV (Multiple valves), DOHC (Double overhead cams), MPFI (Multi-port fuel injection), DFI (Direct fuel injection), VCR (variable compression ratio) and many others to attain maximum possible thermodynamic efficiency and lowest emission rate. In spite of these efforts, the overall thermodynamic efficiency of the reciprocating engines remains limited within one third of the foel energy.

The following problems still persist in known reciprocating piston engines:

1. Hot intake chamber and valves at intake passage both reduce volumetric efficiency.

2. In order to make a successful burning of fuel and to optimally utilize the pressure that produced by combustion, the ignition has to be initiated a few degrees before top dead center(BTDC) of a compression stroke, and this point onward the remaining compression is done against the initial pressure-rise in the compression chamber, resulting in efficiency loss.

3. The piston-crank combine movement does not match well with the expanding fluid pressure and this shortens the power stroke from both ends. In the beginning of a power stroke, peak cylinder pressure is attained at about 20 degrees past top dead center (TDC), and at the end, the exhaust valve has to be opened at about 50 degree before bottom dead center to avoid conflicting stress of expanding gases on piston head. Thus, a significant amount of fuel energy is wasted.

4. Though a hemispherical combustion chamber is the best known chamber shape for IC engine configuration, various designing constraints make such an arrangement impracticable, such as, difficulty in arranging suitable valves and valve operating cam mechanisms, limiting the scope to provide squish region. In spark ignition engine, sparkplug cannot be placed centrally.

5. Power loss due to linear to rotational force translation.

6. Vibration produced by repetitious acceleration and deceleration of piston connecting rod assembly.

The world of IC engine is dominated by reciprocating piston type systems. As a result, commonly available texts and literatures on this subject are based on the said reciprocating piston engine systems.

It has long been a target to design an efficient rotary engine to avoid the reciprocating engine's linear to rotational power transmission mode which suffers from various problems such as vibration due to piston-connecting rod inertia, various complicated and power consuming moving parts, extremely short power stroke, low volumetric and thermal efficiency, low power to weight ratio etc. Various engines were designed to reach those targets but unfortunately, none of them could outperform the reciprocating engines.

The most notable rotary engine known as the "Wankel engine" appeared along with a bunch of promises in this field in the middle of last century. It was assumed that numerous difficulties of reciprocating IC engine could be solved with this rotary. But unfortunately the engine geometry and its working method did not match to make the goal. The engine combustion chamber shape, its high surface- to- volume ratio and high burning gas flow are the major reasons for a large amount of heat loss to the surrounding during combustion, thereby a little portion of energy is left to work.

In a reciprocating piston engine, the cylinder wall, cylinder head and the piston top define a combustion chamber, also known as the clearance volume, where a part of exhaust gas remains trapped at the end of each four-stroke cycle. The next cycle begins with the intake stroke whereby fresh charge is inducted into the enlarging cylinder volume. A swirling motion in the intake charge is created by design of the intake port or a shrouded intake valve to speed up the mixing of air and fuel droplets. Though this swirl is favorable for mixing of air and fuel, at the same time a part of atomized fuel gets centrifuged towards the cylinder wall due to the swirling motion and forms a fuel rich region near the said cylinder wall. In order to minimize heat transfer from the exhaust valve to the fresh intake charge (to achieve maximum volumetric efficiency), it is a compulsion to keep the intake valve as far as possible from the said exhaust valve, which happens to be the hottest part of the combustion chamber. To start ignition, in an SI engine, the sparkplug fires a considerable degrees before TDC, considering a preparation time to form an appropriate early flame kernel. This preparation time is known as ignition delay or ignition lag. hi a CI engine, ignition occurs by raising the pressure and temperature of the charge to a point where the fuel ignites spontaneously. Ignition delay is a common phenomenon to both CI and SI engines, having both physical and chemical reasons depending upon the nature of the fuel, upon both the temperature and pressure, the proportion of the exhaust gas and also upon the temperature coefficient of the fuel, that is, the relationship between temperature and rate of acceleration of oxidation or burning. The induction swirl generally is too weak to mix-up all the burned and unburned charge efficiently, thereby the sparkplug region suffers from a weak mixture, which is unsuitable to form an appropriate early flame kernel. It is known that appropriate turbulence speeds-up combustion speed and in reciprocating engine it is available only near at the TDC. This turbulence is generated by the squish effect between piston top and cylinder head squish pads, from where the fuel rich charge rushes out radially inward, produces turbulence and mixes up with fuel lean part and thus form a turbulent mixture that helps to speed up the flame propagation, thus starts the second phase of combustion or the actual combustion phase. The said ignition delay is responsible for a considerable amount of time loss.

Various geometric constraints such as, requirement of suitable space for placing intake and exhaust valves, necessity of squish pad etc. limit the scope to provide a hemisphere like compact combustion chamber and operatively centrally located sparkplug. These limitations substantially affect the combustion quality, speed and resultant thermal efficiency.

Valves in both intake and exhaust ports of a reciprocating engine decrease its volumetric efficiency as well as increase pumping loss.

All the working phases within a single cylinder, results in interference to each other.

In both conventional reciprocating and rotary IC engines, the intake chamber volume generally defines the expansion chamber volume, whereas the hot expansion gas has the ability to effectively expand itself to more than the intake volume to work optimally.

A comparatively newer concept of split four stroke engine is disclosed in U.S. Pat. No. 7,017,536 to Scuderi. In this engine two separate cylinder-piston arrangements are employed to mutually execute different phases of a four stroke engine cycle within a single crankshaft rotation. In this arrangement, a first cylinder in which a first piston reciprocates between an intake stroke and a compression stroke within a single rotation of respective crank shaft and a second cylinder in which a second piston reciprocates between an expansion stroke and an exhaust stroke within the same rotation of the said crankshaft. A gas passage interconnects said first cylinder and said second cylinder to deliver the compressed charge from said first cylinder to the said second cylinder where the said charge gets combusted and expands against respective piston to impart driving force to the said crankshaft. Both the first and the second piston are connected to a common crankshaft but with different crank throws, through individual connecting rods in a phase-shifted relation to each other. Inlet check valve and poppet type crossover outlet valve are disposed on either distal end of said gas passage to control the gas delivery. Mechanical cam arrangement is employed to drive said crossover outlet valve, intake valve and exhaust valve in a predetermined order. On completion of an exhaust stroke of the said second piston crossover outlet valve opens to deliver compressed charge to said second cylinder where said second piston starts descending and at the same time the first piston of said first cylinder still ascending near top center to complete the compression stroke. As the second piston descends to a position relating to about 30 degree angle of respective crank, an electric spark starts the ignition and thus starts the power stroke.

The above stated ideas do not address well-known demerits of an IC engine. Moreover, some new problems arise with these split-cycle arrangements, such as:

1. Delivery of compressed charge starts after TDC. Ignition initiates almost at the end of charge delivery. Therefore, considering the time required for combustion and its preparation, a fewer crank degrees, as compared to conventional piston engines, are then left for imparting the driving force to the crankshaft.

2. Ignition is initiated at about 30 degree crank angle past top dead center, the chamber volume expansion rate being considerably higher from this point onward, the engine cannot gain sufficient operating speed.

3. Duration of crossover delivery is short, so peak compression cylinder pressure becomes very high, thereby consumes a considerably large amount of driving force.

4. Firing once in every two strokes may lead to engine over-heating that may result in auto ignition, piston deformation, etc.

5. Crossover outlet valve is highly problematic. To retain a poppet type valve (as mentioned in U.S. Pat. No. 7,017,536) closed against 'a predetermined firing condition gas pressure' of compressed charge a strong valve retaining force becomes necessary that might destroy the valve seat.

The most reputed rotary engine known as the "Wankel engine". In this engine a equilateral triangular rotor rotates around a shaft eccentric within a trochoidal housing. The rotor has three apex portions and generally oval working faces, each of the working faces is extended between two consecutive apices. Each rotor working face and adjacent housing peripheral wall, being enclosed by housing sidewalls, defines a working chamber there between. During rotation of said rotor, volume of working chambers varies between a maximum and a minimum working chamber volume. Therefore, three volume-changing chambers, adjacent to three respective working faces of respective rotor, simultaneously execute three different phases of a four-phase engine cycle. In rotary IC engine category, the Wankel type engines proved their lead due to their constructional simplicity, rotary smoothness, high power to weight ratio etc. At the same time, Wankel type engines suffer from a number of demerits, such as;

1. Higher break specific fuel consumption in comparison to reciprocating piston engine power output.

2. Continuous angle changing relation between apex seals and adjacent peripheral wall results in poor sealing therebetween and high wear rate of sealing elements. Formation of chatter marks on peripheral wall surface is another problem often resulted in this kind of sealing system.

3. Lubrication is generally provided by mixing oil with intake charge. This results in poor lubrication, inefficient combustion and carbon formation in working chamber compartments.

4. High exhaust emission rate.

5. Rate of heat loss is high.

6. Inability of ignition advance.

The limitations of the Wankel engine are discussed herein as a referential context of the present invention.

In a Wankel rotary, during power delivery phase, the rotor exerts radial inward force on the shaft-eccentric, this working mode is similar to the working relation of piston-crank mechanism of a reciprocating piston engine, hence, cannot extract useful work near TDC. The narrow elongated combustion chamber shape of the wankel engine, having a high surface to volume ratio, is found to be responsible for its poor combustion quality. During combustion, as a portion of burning mass comes in contact with comparatively cooler surrounding surfaces, gets extinguished by releasing heat to said surfaces and gets deposited as unburned hydrocarbon that finally is expelled to the exhaust.

In Wankel, as the compression chamber approaches TDC, a volume changing relation is formed between its recess connected leading and the trailing parts where the leading part expands while the trailing part contracts with the rotor movement. This volume changing relation, considering the operative speed of the engine, produces a huge pressure differential therebetween, and so charge continues to flow from the trailing to the leading chamber at high velocity and said charge flow persists as long as the trailing chamber exists. When passing through TDC, the respective working face of rotor moves almost tangentially and therefore the trailing chamber pressure, which is higher than that of leading chamber, acts against rotor's forwarding motion. The said gas flow produces strong turbulence in the leading chamber working fluid that accelerates combustion. But, at the same time, the high surface to volume ratio of the combustion chamber and presence of strong turbulence in burning fluid, leads to considerably large amount of heat loss to the surrounding surfaces. Moreover, during combustion, the pressure rise in the trailing chamber has to work against the contraction of said trailing chamber volume and this conflict suffers loss of a considerable amount of heat energy. This also results in housing injury near sparkplug vicinity, shaft bend, engine overheating etc.

The apex sealing arrangement of the Wankel engine remains problematic due to its variable contact angle with the chamber peripheral wall on which the sealing elements skate. Due to this variable contact angle the connecting tip substantially is of oval cross-sectional profile. This requires heavier load on the seals to keep them in sealing contact and this in turn results in high friction and high rate of seal wear.

The lubrication of apex seal is generally done by mixing oil with intake air-fuel charge which results in poor lubrication, as well as burning of lubricant oil that causes carbon deposition in engine compartments and poor combustion quality. The carbon particles tend to restrict the movement of various sealing elements, which results in compression loss, combustion leakage, lubricant contamination, engine overheating, etc.

An alternative attempt, to overcome the said lubrication demerits, is disclosed in U.S. Pat. No. 5,305,721 to Burtis. Burtis discloses an apex sealing arrangement in which dual cooperative apex sealing elements are employed at each apex of a tringular rotor. Each seal element fabricated with spring elements to provide diametric and lateral loading of the seal against the housing. The spring elements that provide lateral loading of a seal element, bear against a protruded soulder of the concomitant seal element and vice versa. The lubrication of the apex sealing elements is provided by feeding oil from the internal lubrication system of rotor. An oil passageway interconnects the apex seal groove and said internal lubrication system of the rotor. An efficient lubrication is anticipated by relative movement of the dual seal, whereas, both the seal elements substantially have an end gap between its lateral loading spring side and the protruded shoulder of the concomitant sealing element to allow lateral movement there between. This end gap connects the bottom side gap of sealing elements where the diametric loading spring is employed. These interconnecting gaps provide direct link to adjacent working chambers and apex seal lubricating oil passageway. Therefore the pressurized lubricating oil accomplished by centrifugal force results in oil-flood in working chambers and at the same time high pressure combustion gas may enter into rotor coolant passages through same gaps and contaminate oil.

Some of the variations of Wankel engine include a different type of sealing arrangement to provide periodic sealing cooperation between trochoidal housing peripheral wall and rotor working faces to carry-out different purposes. Such a sealing arrangement is disclosed in U.S. Pat. No. 3,391,677 to Hejj, where a Wankel type rotary engine having the said sealing arrangements which are disposed at the minimum radius portions of the epitrochoidal housing. Being radially adjustable, said seal decrease the gap between the housing and the rotor that approaches TDC and thus divides the respective working chamber into two volume changing portions, which are the leading expanding portion and trailing volume contracting portion. An 'antechamber' is located near TDC, which is connected to both the leading and trailing portion of the working chambers through connecting bores. The compressed air of trailing chamber is then forced through said bores to enter into said antechamber, in which fuel-air mixing and combustion take place. The combusted charge then expands through said leading chamber portion through the bores interconnecting the antechamber and said leading working chamber.

U.S. Pat. No. 3,994,266 to Jones discloses an arrangement of housing-rotor sealing cooperation for a Wankel type rotary diesel engine. In this engine, seal means are disposed near the lobe junction area of a two-lobed epitrochoidal housing chamber. Said seal means, being operated by an operating member, periodically expands to provide a sealing relation between the inner housing peripheral wall and the respective rotor working face near TDC. In order to achieve a higher compression ratio sufficient for diesel operation, seal means are provided so that during the compression phase each working chamber of the engine is divided into leading and trailing portions with compression of its intake charge being substantially confined to its trailing portion. Compression thereafter continues in the trailing portion of the chamber whereupon the minimum volume of this trailing portion at the top dead center position of the working chamber is substantially less than what it would be in the absence of said seal means. In this way, compression ratio, sufficient for diesel operation is attainable in the trailing portion of each working chamber and therefore upon fuel injection into this trailing portion, the fuel self ignites to provide diesel type combustion.

In both Hejj and Jones patents, axially parallel deep grooves or recesses are provided on inner housing peripheral wall to accept sealing elements into said grooves. During rotation of a rotor, each apex seal passes over each groove once on every complete turn of said rotor. As the apex seals are so spring biased as to produce diametric loading of said seal on said housing peripheral wall, detrimental foul will occur between said apex seals and said housing to rotor seal grooves. It should be considered that the tip of a conventional apex seal element is substantially of oval cross sectional profile because of its continuous angle changing relation with adjacent peripheral wall. Therefore, said tip profile of said seal element limits the scope of overcoming the said problem by increasing the seal thickness.

In the said patent (U.S. Pat. No. 3,994,266) Jones discloses a disel combustion method, where, immediately after the initiation of combustion, the combusting fluid of trailing combustion chamber suddenly allowed to expand through the entire leading chamber. Thereby the peak cylinder pressure drastically drops without doing any work. Therefore, useful thermal efficiency cannot be achieved in this configuration.

U.S. Pat. No. 3,777,720 to Williams discloses a rotary engine arrangement in which only combustion and exhaust phases of an engine cycle occur. Therefore, in the engine arrangement, two separete combustion chambers are so provided as to achieve six succesive power impulses within a complete rotation of the rotor. The engine comprises a two-lobed epitrochoidal housing chamber in which a triangular rotor rotates to execute the said combustion and exhaust phases of an engine cycle. Each apex portion of said triangular rotor comprises apex seal means which rotate-ably keep sealing relation to epitrochoidal peripheral wall. Each pair of apex seals defines a working chamber there between. Auxiliary seal means are mounted in a central location in each of the convex faces of the triangular rotor. Near the top center position, these auxiliary seal means comes into engagement with housing peripheral wall near its minimum radius area. Thus, within the leading portion of the rotor faces, forms a combustion chamber defined by said auxiliary seal means and the respective leading apex seals. A delivery slot on a rotary sleeve valve rotatively connects to corresponding combustion chamber inlet port to deliver precompressed charge from a pre-compressed charge storage tank where a predetermined pressure being maintained throughout. The sleeve valve is driven in timed relation with the engine shaft.

In this engine six consecutive combustions occur within a single rotation of rotor. It is known to those well informed in the related art, that the shape of the Wankel engine combustion chamber is one of the major reason for engine overheating, whereas conventional Wankel engine fires thrice within a single rotation of rotor. Therefore the Jones patent is not practicable without appropriate measure to cope with the problem of engine overheating. Auxilliary seal means are always kept exposed to hot combustion product and so lubrication of said seals are very difficult whereas lubrication is very much necessery for this kind of seals that reciprocate to maintain contact with corresponding surfaces. A rotary sleeve valve in a close proximity of combustion chamber remains problematic in its application with different engines. Instead of this, with increase of engine speed the charge delivery time in this valve system decreases. As charge delivery pressure is constant, shorter delivery time results in fuel defflciency in combustion chamber. The auxiliary seal means maintain sealing contact during a critical portion of each active combustion. During this period, power transmits to the drive-shaft via rotor positioning gears. Thereby, the single sided gear arrangement would generate a tilting force on the rotor, causing wear of said gear-set, rotor bearing and shaft journal as well as the malfunction of apex seal means.

It remains quite complicated to expand a Wankel type engine with more than two rotor configurations because such an arrangement with a common shaft restricts the installation of stationary relation gears for intermediate rotor configurations due to the presence of eccentric lobes on said shaft. To mitigate the above drawbacks of the prior art, it has been the applicant's endeavor to invent an IC engine having split cycle and added advantages as described hereinafter in this specification.

Summary of the invention:

The previously stated ignition delay can be minimized by improving scavenging efficiency, breathing capability, early mixing of air and fuel, heating up the mixture more quickly, creating a strong turbulence earlier and providing compact combustion chamber with centrally located sparkplug.

This and other advantageous aspects achieved in the present invention will be clearly understood from the following description: a) Scavenging efficiency: According to present invention, crossover valve initiates compressed air delivery at a moment when respective working face of rotor is approaching top center position but the leading portion of said working chamber is still connected to the trailing exhausting portion through a cavity formed on the leading portion of each working face of respective rotor. The initially delivered compressed air expands in the leading portion by expelling the trapped mass of exhaust gas from previous combustion. The trapped exhaust gas then rushes out from the forwarding portion to the trailing portion through said cavity, and later is rejected through the exhaust passage. Soon after the ejection of the exhaust gas, a housing-to-rotor sealing arrangement becomes activated to separate the said leading portion from the trailing exhaust portion. b) Breathing efficiency: In the present embodiment, a separate arrangement is used for intake and compression phases of a four-phase engine cycle, where the valve-less intake passage offers nominal obstruction to intake airflow. After compression of the air, it is delivered to the expansion-exhaust configuration Cl. Being spatially apart from said expansion-exhaust configuration Cl, the intake-compression configuration C2 remains colder than expansion-exhaust configuration that results in lower heat transfer from chamber wall to intake air, thereby higher charge density is obtained. Moreover, the intake time is 1.5 times longer than the conventional reciprocating engine. All these facts accomplish to achieve much higher volumetric efficiency. c) Air-fuel mixing: The compressed air is delivered to the combustion chamber in the form of high-pressure air jet. The fuel injector, being fitted angularly to the compressed air inlet port, injects fuel aiming to this air jet stream. During injection, the high-speed air jet penetrates the fuel droplets; break the droplets into finer particles. Being accompanied by hot turbulent air, the atomized fuel then vaporizes and mixes very quickly with air. The strong turbulence accomplishes quick heat transfer from hot combustion chamber surfaces to the air-fuel mixture, resulting in earlier temperature and pressure rise in combustible mixture that substantially decrease the ignition delay. d) Combustion chamber shape: The shape of the effective combustion chamber, that consumes the major volume of the total combustion chamber, is nearly hemispherical. The said combustion chamber shape results in lesser heat loss to the surrounding, fast burning rate, etc. Due to the air inlet fashion, the volume consumed within the thinner portion of the said combustion chamber, contains fuel lean mixture, which produce less heat energy during combustion and thus heat loss to the walls of the said thin volume is minimized. e) Sparkplug location: The sparkplug fires at the center of the moving combustion chamber in a synchronized manner. The combustion chamber shape being nearly hemispherical, equidistant flame propagation in all directions within the hemisphere becomes possible, thereby quick and more complete burning of fuel is achieved. f) Expansion efficiency: The expansion chamber of expansion-exhaust configuration expands to a volume that is substantially higher (about 28%) than the intake chamber maximum volume of intake-compression configuration. Therefore, in contrast with the conventional engines, a substantially higher amount of heat energy has the scope to be converted into useful work. Said higher expansion-ability results in lower exhaust gas pressure, which allow late opening of the exhaust port and longer expansion phase. g) Power transmission efficiency: The present invention offers advantages and alternatives over the prior art by providing a split four-phased rotary engine having a Dual Power Transmission Mode. Of which, the 'initial power transmission mode' become effective during at least the active combustion event when combusting charge exerts pressure only on the leading portion of the working face of respective rotor, thereby delivers a rotational force to respective rotor rather than "piston-crank like radial force" of the conventional rotaries and thereby can efficiently extract useful work when the working face of respective rotor moves through 'top center' and even a few degrees before said top center. On completion of combustion and during early phase of expansion of the combustion gas, said pressure area on the leading portion of said working face of said rotor is then allowed to expand through the entire respective working chamber. Thereafter the said power transmission mode gets converted to piston- crankshaft like radial power transmission mode which, therefore, may be defined as the 'final power transmission mode'. Thereby, higher amount of heat energy gets converted to work. The said initial power transmission mode exerts a cranking load on the rotor-housing relation gears. Considering this fact, pair of internal and external helical gears is provided on either side of rotor and housing respectively. h) Compressed-air delivering efficiency: Split cycle IC engines generally use different arrangements for compression and expansion, which are critically dependent on the operating competency of the crossover valve arrangement. Long crossover delivery time (about 60 degree shaft rotation time), Precise delivery of compressed air, small amount of operating force for valve operation, is achieved by a novel crossover valve system described later. i) Designing flexibility: Unlike the known internal combustion engines, the present invention has an engine arrangement wherein each of the working phases of a four phase engine cycle are non-overlapping. Moreover the split cycle rotary IC engine, according to the present invention, can easily be configured and converted for different types of fuel. Flexibility to expand the engine configurations in different arrangement is quite high. Modular arrangement of the engine makes it convenient to assemble, disassemble and even reconfigure to different arrangements.

Brief description of drawings:

FIG. 1 shows the sectional view of the expansion-exhaust configuration of a preferred embodiment of the present invention with the center-shaft.

FIG. 2 shows the open side axial view of expansion-exhaust configuration. FIG. 3 shows the open side axial view of intake-compression configuration.

FIG. 4 schematically illustrates the shaft relation between expansion-exhaust configuration and intake-compression configuration and the locations of balancing counterweights.

FIG, 5 shows the phase relation between the intake-compression configuration and the expansion-exhaust configuration.

FIG. 5 A and 5B shows relative rotor positions of expansion-exhaust configuration

Cl and intake-compression configuration C2 during ignition.

FIG. 6 illustrates the relationship between the seal holder and the seal holder bearing.

FIG. 6A shows seal holder bearing. FIG. 7 shows the exploded view of the seal holder. FIG. 7 A shows a seal element.

FIG. 8 shows the relationship between the apex seal element and the peripheral wall.

FIG. 9 shows the housing seal and its operating mechanism.

FIG. 10 shows cross-sectional view of seal arrangement shown along the line 10-10 of FIG. 9.

FIG. 11 shows the geometrical relationship of the intersecting arcs for forming contour of rotor working face.

FIG. 12 shows the two parts of the bipartite housing peripheral wall.

FIG. 13 shows crossover valve arrangement.

FIG. 14, 14A, 14B and 14C illustrate sealing arrangement of demobilizing disc.

FIG. 15 shows the relation between leading and trailing chamber during scavenging.

FIG. 16 shows the combustion chamber during injection of air and fuel. FIG. 16A schematically shows the relation between inlet air swirl and fuel jet.

FIG. 17 shows the relation between the combustion chamber, spark plug and air inlet passage.

FIG. 17A shows air flow pattern during initiation of ignition.

FIG. 18 shows lubrication arrangement for swivel seal holder and seal elements when the needle valve is open for the discharge of oil.

FIG. 18A shows lubrication arrangement for swivel seal holder and seal elements when the needle valve is closed to stop oil discharge.

FIG. 19 shows an alternative arrangement for diesel application. FIG. 20 shows a plan view of an alternative four rotor arrangement. Best mode of carrying out the invention:

A rotary IC engine arrangement, comprising at least two supplementary configurations of which the first one is provided for executing only expansion and exhaust phases of a four phase engine cycle and the second one is provided for only the intake and compression phases of the said engine cycle. Crossover passages are provided for delivering compressed air from the intake-compression configuration to the expansion-exhaust configuration

[A rule of representation: Some of the callout number used in this document represents a group of operatively similar components and individual callout numbers for these similar components are formed by attaching different small letters to the original callout number representing the group, (e.g., swivel seal holders 66 represents a group of operatively similar seal holders, and individual seal holders under this group are numbered as 66a, 66b, 66c, etc.)]

With reference to the accompanying drawings FIGS. 1 and 2 illustrate the expansion-exhaust configuration Cl comprising an epitrochoidal inner chamber including epitrochoid shaped peripheral wall 31 enclosed by two opposing sidewalls 32, 33. Inlet passages 34a and 34b are provided at symmetrically opposite position near the indentations 35a, 35b on epitrochoidal peripheral wall 31. Near the inlet passages 34a, 34b fuel injectors 36a, 36b and sparkplugs 37a, 37b are also located accordingly to contribute in the preparation of combustion, therefore the unity of inlet passage 34, fuel injector 36 and sparkplug 37 will hereinafter be referred as combustion means. Two exhaust ports 33a, 33b are provided on sidewall 33 in such a position, where expansion chambers connect said port in turn at the conclusion of their respective expansion phases.

The sidewalls 32, 33, having centrally bored through-holes 38, 39 respectively, said holes are coaxial with respective housing. The rotor 40 rotates within said housing about the shaft lobe 45 eccentrically integrated with a center-shaft 44, which is coaxial with the housing axis. Said shaft extends through the said holes 38; 39 on the sidewalls 32, 33. The ratio of the extended radius to the eccentricity of the housing epitrochoid (usually known as K-factor) has preferably a value of about 7.

Referring to FIGS. 1 and 2, the rotor 40 of expansion-exhaust configuration Cl has an equilateral triangular shape with modified working faces extended between each pair of its three apices. Hemisphere shaped cavities 47 (47a, 47b, 47c) are provided on the forwarding portion of each working faces 46 (46a, 46b, 46c) of the rotor 40.

Coaxial bore 48 is formed into the rotor 40 to support the outer raceway 49a of the roller bearing 49, which is press fitted into said bore 48 and secured in place by suitable fastening agents. The center-shaft eccentric 45, about which said rotor 40 rotate, acts as the inner raceway of said roller bearing 49. Said center-shaft 44 is coaxial with respective housing and said center-shaft eccentric 45 is coaxial with respective rotor 40.

Internal ring gears 51 (51a, 51b) are fitted coaxially into matching bores 53 (53a, 53b) formed on either sides of said rotor 40. Theses bores are substantially larger in diameter than the bore 48 provided to support roller bearing 49.

The stationary external ring gears 52a, 52b are rigidly fitted on bores 38, 39 of respective sidewalls 32, 33. The internal gears 51a, 51b on the rotor 40 are connected in mesh with the stationary external gears 52a, 52b respectively. Both the internal gears 51a, 51b and the stationary external gears 52a, 52b are preferably helical gears. The internal ring gears 51a and 51b are substantially of opposite hand helical gears and thereby the corresponding external gears 52a, 52b are of matching gear hand to said internal ring gears 51a, 51b accordingly. The gear ratio of internal ring gears and the stationary external ring gears is 3:2, so that for a single revolution of the rotor about its own axis, the respective shaft rotates thrice in the same direction.

Both the configurations Cl and C2 comprise individual center-shafts 44 and 54 respectively. This allows the insertion of stationary ring gears 52, 62 through their both ends. During assembly, the external gears insert-ably mesh with corresponding internal gears. Because of their helical profile, the external gears require a twist during insertion to mesh with corresponding internal gears. Considering this, the external gears would be fitted finally on said sidewalls, preferably after assembling each elementary configuration. Splined shaft end 55 of shaft 44 of expansion-exhaust configuration Cl , joins coaxially to the corresponding splined shaft end 56 of shaft 54 of intake-compression* configuration C2.

Referring to FIG. 3, the intake-compression configuration C2 is geometrically similar to the above stated expansion-exhaust configuration Cl with a proportionally reduced dimension e.g., in an exemplary embodiment of the invention, the intake-compression configuration C2 has a volume about 22 % smaller than that of the expansion-exhaust configuration Cl. The intake passages 42a, 42b are located on the sidewall 42, whereas the compressed air delivery passages 61a, 61b are located on the housing peripheral wall 41. Therefore, two complete intake-compression phases take place during one complete revolution of center-shaft. The major difference lies on the rotor configuration. The rotor 50 of the intake-compression configuration C2 has no recess on its working faces.

Referring to FIG. 4, both the expansion-exhaust configuration Cl and intake- compression configuration C2 are individually balanced by suitable counterweights 64 and 65 respectively. Said counterweights are provided preferably on both sides of their extended shafts. Both the configurations are, therefore, phase independently balanced by themselves. The expansion-exhaust configuration Cl and the intake-compression configuration C2 coaxially join together, leaving a space there between. This space prevents thermal transfer from the expansion-exhaust configuration Cl to intake-compression configuration C2, as well as provides assembling ease.

FIG.^' 5 shows the phase relation between the intake-compression configuration C2 and the expansion-exhaust configuration Cl in a preferred embodiment of the invention. The phase difference between the intake-compression configuration C2 and the expansion-exhaust configuration Cl is about 20° crank angle. According to the present invention the intake-compression configuration C2 is always phase advanced to the expansion-exhaust configuration Cl in accordance with their rotational direction. FIG. 5 also shows working chamber status during the initiation of crossover delivery of compressed air from C2 to Cl. Arrow 99a schematically shows the crossover connection between C2 and Cl . Crossover outlet valve arrangements 104 (104a, 104b) and crossover inlet check valve arrangements 103 (103a, 103b) are described below.

FIGS. 5A and 5B showing the relative rotor positions of expansion-exhaust configuration Cl and intake-compression configuration C2 during ignition. Relation of air delivery passages are shown schematically by the arrows with broken line 99a and 99b.

Referring to FIG. 6, swivel seal-holders 66 are rotate-ably fitted into matching bores 76 respectively on apex regions of the rotor 40. The bores 76 (76a, 76b, 76c) provide bearing support for the seal holders 66, and therefore, will be referred hereinafter as seal holder bearing. Said bearings 76 are longitudinally one side open bores extended throughout the thickness of the rotor. The cross-axial opening of said bores 76 (see FIG. 6A) is substantially smaller than the bore diameter.

In addition, suitable seal means 60 (60a, 60b, 60c) are provided (as in FIGS. 1 and 6) between the rotor 40 and housing sidewalls (32, 33) to complete a seal grid around each of the working chambers 57, 58, 59.

Referring to FIG. 7, each seal holder 66 is an assembly of two corresponding pails 67 and 68 are of equal diameter. Axially projected cylindrical portion 68b of the part 68 slide-ably fits into the corresponding hollow portion 67a of the part 67. Pair of longitudinally extended guide pin 69 of part 67 slide-ably fits into a pair of corresponding notch 70 of part 68 to ensure simultaneous rotation their between. Spring element 71 is provided between the seal holder parts 67 and 68 to produce lateral pressure to secure corner sealing. The seal-holders 66 having shape of a cylinder longitudinally cut off to form a plane surface on it. Said plane surface will hereinafter be referred as 'seal holder face' 72. Each seal-holder is slide-ably coaxially inserted along the seal holder bearing 76 (see FIG. 6A) and the seal holder face 72 is kept exposed outwardly through the longitudinally open portion of the respective bores. Said seal holders 66 are sized to rotate within the respective bearings 16, about their own axis in either direction.

Referring to both the FIGS. 7 and 7A₅ on each seal holder face 72 three parallel grooves 73, 75, 74 are extended throughout the length, into which respective seal elements 78, 80, 79 fit slide-ably. The grooves 73, 75, 74 are substantially perpendicular to the respective seal holder face 72. One of the apex seal elements 78, shown here as an example, is a two-piece diagonally-split seal strip, loaded by spring elements 77a, 77b which provide diametric and lateral loading of the seal against the peripheral wall and adjacent sidewalls respectively.

Referring to FIG. 8, each apex seal holder 66 contains three seal elements 78, 80, 79 of which the intermediate element 80 is an oil scraper that operatively scrape excess oil off from the respective housing peripheral walls. During operation, the face 72 of seal holder 66 maintain a uniform clearance with the facing housing peripheral wall 41, the spring biased seal elements 78, 80, 79, being extended from said holder, always maintain a perpendicular sealing relation with both the said facing peripheral wall 41 and adjacent sidewalls 32, 33. Similar apex sealing arrangements as described above are employed to both the configurations Cl and C2.

Referring to FIG. 9, the seal element 81 is a rectangular strip, comprising a narrow elongated blank 82 to receive an arc shaped leaf spring 83. A small rectangular groove 82b is extended from the blank 82 to allow the self-engaging retainer pin 84.

As shown in FIGS. 9 and 10, the seal elements 81 slide-ably fit into the corresponding housing seal slots 85 near the indented portions 35 of epitrochoidal peripheral wall. Said seal elements 81 sequentially close the gap between said indented portions and corresponding worldng faces of rotor, thereby divide the respective chamber volumes into two separate chambers on either side of the seal element 81. The length of the housing seal slot 85 is substantially smaller than the respective rotor thickness. A few millimeters thick flanges 86 are left on either end of housing seal slot 85. The flanges 86 provide the guide tracks for apex seal elements 78, 80, 79 to accomplish their passing over the seal slot 85 without fouling. The self-engaging retainer pin 84 is a spring biased rectangular pin slide- ably disposed into the matching groove 87 on housing seal slot 85. The retainer pin 84 has a chamfer on its face to facilitate easy entrance of the seal element 81 during install. Each housing seal slot 85 having an open end 88 towards inner housing and a close end 89 towards the outer periphery of the housing. The self- engaging retainer pin provides support for housing seal leaf spring 83 to bias the seal element 81 towards the close end 89 of housing seal slot 85. The seal element 81 is arranged to be retracted periodically from the respective rotor working face by a cam arrangement 90, located on the outer periphery of both the expansion- exhaust configuration Cl and intake-compression configuration C2. Each cam arrangement 90 includes a camshaft 92, on which cam 102 is arranged to be driven at the same rotational speed as the engine center shaft by means of timing relation. A cam follower 91, comprising two arms 93 and 94, extending radially from a pivot 95. The arm 93 is biased upwardly (as shown in FIG. 10) by a strong spring 96 to follow the cam profile 102. The arm 94 has a bifurcated extension 94a, 94b from pivot 95. The bifurcated arm 94a, 94b connects two spatially arranged linkages 97 (97a, 97b) movably extended through corresponding holes 98 which extending through the close end of respective housing seal slot 85. As the cam operatively releases cam follower 91, spring 96 exerts pressure to the linkages 97a, 97b via arm 94. The linkages 97 thereby extend towards seal slot 85 to press the seal element 81 against respective rotor working face. The relatively weak housing seal spring 83 is engaged for retracting the seal element 81 when permitted by the cam 102. Similar sealing elements 81, as described above, are employed to both the configurations Cl and C2.

Each rotor working face is contoured so as to maintain a minimum possible clearance with the indented lobe junction of respective inner peripheral wall or the housing minor axis. While passing the minor axis region, the modified contour of rotor face enhances the performance of housing seal by reducing the gap there between.

Referring to FIG.l 1, The modified contour of rotor working face can be derived by intersecting two circular sectors Sl and S2 of different radii Rl and R2 accordingly whereas the centers of the said circular sectors Pl and P2 lie on the line L extended from rotor center X through the opposite apex Y.

The most preferable rotor face contour can be derived by computer aided design generation method, i.e., continuously intersecting a moving circle (the circle consuming rotor diameter) by a suitable stationary epitrochoid where the moving circle follows the rotor's movement. Considering the operating housing to rotor clearance, the rotor profile would be generated by CNC tooling.

It will be apparent from FIG. 12, that in a preferred embodiment of the present invention, the epitrochoidal peripheral wall is constructed by joining two separate lobes 3 Ia, 3 Ib together where both the opposing sidewalls are single-piece member. The joining faces of said two-piece lobes coincide with the extended minor axis of respective inner housing. Dowel pins and suitable fasteners secure the joining of separate lobes and sidewalls respectively. The said method of construction provides machining convenience for housing seal slot 85 and other related components. Both the peripheral walls of Cl and C2 would preferably be constructed using above described method. The conventional method of one piece casting of epitrochoidal peripheral wall can also be appreciated for this engine.

Referring to FIG. 13, crossover air passage 99, interconnects the crossover inlet check valve arrangement 103 and the crossover outlet valve arrangement 104. Being connected by the crossover passage 99 the crossover inlet valve arrangement 103 and the crossover outlet valve arrangement 104 define a pressure chamber 105 therebetween. The crossover outlet valve arrangement 104 comprises an outlet valve disk 106 and a spaced apart demobilizing disk 107 on a common stem 108. An axially apart switch valve 109 comprises a pressure release vent 111 in its stem 110. A sub-chamber 112 extends between the demobilizing disk 107 and the switch valve disk 109. The stems of both the crossover valve and the switch valve extend outside of the pressure chamber 105 and sub chamber 112, where a compression spring 113 is engaged with the crossover valve stem 108 to sustain a pressure on the crossover valve disk towards its closed position and the compression spring 114 is employed to sustain a pressure on the switch valve 109 towards its closed position. The check valve 103 permits the one way flow of compressed air from the compression chamber of intake-compression configuration C2 to the pressure chamber 105. The crossover valve 104 permits the one way flow of compressed air from the pressure chamber 105 to the expansion chamber of the expansion-exhaust configuration Cl . Outlet valve disk 106 connects the bifurcated combustion chamber inlet passage 34. During operation, previously trapped compressed air of pressure chamber 105 exerts pressure on both the outlet valve disk 106 and the demobilizing disk 107. . Demobilizing disk 107 is arranged to receive an equal and opposite directional force to that of the outlet valve-disk 81, thereby makes the crossover valve inactive. Thus the pressure of compression spring 113 remains as the determining force, which keeps the crossover valve 106 shut. In order to open the crossover valve, the switch valve 109 is opened by a cam arrangement 116 to allow compressed air into sub-chamber 112 in order to equalizing pressure of both pressure chamber 105 and sub-chamber 112, and thus achieved pressure equilibrium on both the sides of the demobilizing disk 107 for making demobilizing disk inactive. Thus the outlet valve disk 106 is facilitated to open by the air pressure of pressure chamber 105, which is much higher than the opposing compression spring 113. On opening of outlet valve disk 106, pressure chamber 105 becomes connected to working chamber 57, which results in pressure drop in pressure chamber 105 below the then developed respective compression chamber pressure of intake-compression configuration C2. This pressure differential facilitates the check valve disk 115 to open. A stream of compressed air then persists from compression chamber of C2 to leading working chamber of Cl until the crossover outlet valve 106 closes. During the final phase of delivery of compressed air, ignition is initiated in combustion chamber 57a (see FIG. 2) which results in pressure rise in said combustion chamber 57a. Thereby, soon after the ignition, pressure differential between pressure chamber 105 and combustion chamber 57a becomes too low to keep the crossover outlet valve 106 open against compression spring 113. At the same time, switch valve become closed following cam 101 and synchronizingly, the pressure release vent 111 opens to atmosphere, results in pressure drop in sub-chamber 112, which also facilitates the closing of crossover outlet valve 106.

For purpose herein, in the exemplary embodiment of the invention, the mechanical cam 101 on cam-shaft 92 rotate-ably provides the driving force to the switch valve 109. The camshaft 92 is a common camshaft provided to serve driving force to housing seal operating mechanism described earlier.

In the present embodiment of the invention, the performance precision of crossover valve arrangement critically depends on the sealing efficiency of the demobilizing disk 107. As shown in FIG. 14, pair of ring seal 117, 118 is disposed into the seal seat of demobilizing disk 107 to produce diametric sealing on surrounding surface to prevent compressed air leakage from pressure chamber 105 to sub-chamber 112. The ring seal 117 is a typical split end ring seal. The ring seal 118 comprises matching radial step-cut (as shown in FIGS.14 A, 14B) on its split ends 118a, 118b so as to provide radial overlap between the said ends. This radial overlapping closes the radial gap on the seal 118. Gaps that exist along seal thickness of both the seals get mutually closed by covering each other. The minute bar 119 (as shown in FIGS.14C) is protruded perpendicularly from a side of ring seal 118. Being placed within the end gap of ring seal 117 the said minute bar 119 of ring seal 118 prevents the relative rotation between them.

Referring to FIG. 15, the crossover valve is devised to initiates the compressed air delivery to the respective leading portion of working chamber 57 of the expansion-exhaust configuration Cl at a time when said leading portion of said chamber 57 is still open to the trailing portion by means of hemispherical cavity 47. The initially delivered compressed air (indicated by arrows with broken lines) expands in the leading chamber and occupies the space, thereby the trapped mass of exhaust gas (indicated by arrows with solid lines) of previous combustion gets expelled from the leading portion of said working chamber to the trailing portion through the said opening, and later is rejected to the exhaust port 33b. Soon after the ejection of the exhaust gas, the leading portion of working chamber 57 get sealed off from the trailing portion of the same working chamber by the housing seal 81 (as shown in FIG.16) and thus forms the combustion chamber 57a.

Referring to FIG. 16, the combustion chamber inlet passage 34 is so bifurcated as to produce desired swirl and turbulence in inlet air (schematically shown by arrows 120, FIG. 16A), which accomplishes the early mixing of fuel and air. The swirling inlet air collects the heat quickly from the surrounding hot combustion chamber wall and being accompanied by turbulence, distributes the heat to all the chamber contents, resulting a quick vaporization of fuel. Due to high pressure differential between delivery side and delivered side, the compressed air enters combustion chamber through inlet passage 34 in a very high speed during the early phase of induction. The high-speed air-jet (shown by arrows) penetrates the injected fuel droplets, breaking them into finer particles to form a fuel mist. The high turbulence in early phase of air-fuel mixing substantially reduces the preparation time. The fuel injector is devised to spray as long as the moving hemispherical cavity is present over it, whereas the compressed air delivery continues until the delivery side pressure remain considerably higher than the delivered side. In a spark ignited engine, as shown in FIGS. 17 and 17A₅ during the initiation of ignition, air inlet ports remains outside the hemisphere to complete the final phase of delivery. The then delivered compressed air, in accordance to inlet passage design, flows more or less along the comparatively thinner part of the chamber (as shown in FIG. 17A). This makes the thinner chamber area fuel-lean, resulting less heat loss to the chamber walls during combustion.

The burning speed of fuel in this kind of combustion chambers is very high. This result in very quick pressure rise in combustion chamber. The volume of combustion chamber inlet passage 34, which is a fuel less region, reciprocally contracts with the rising pressure of respective combustion chamber and compensate the shock.

Referring to FIG. 1, the main oil passageway 121 extends axially through the center shaft 44 for lubrication and cooling all the internal components of the engine housing. Pressurized oil is delivered through this passageway by an oil pump (not shown). Oil passageway 123a and 123b extend from the main oil passageway 121 to lubricate the roller bearings 130a and 130b respectively. Bearings 130a, 130b provide support for the center-shaft 44. Oil passageway 124 extends from main oil passageway 121 to feed oil to roller bearing 49 supporting the rotor 40.

Though use of the central oil passageway and an oil pump is well known in the rotary engine art, an enhanced arrangement is necessary for the present embodiment, where combustion occurs in every 180 degrees rotation of shaft, twice than a conventional Wankel type rotary engine.

Oil passageway 122 extends from said main oil passageway 121 to the farthest end of center-shaft eccentric 45 to provide oil for cooling the rotor 40. An extra mass to the thickness of the rotor is added to provide space for oil passageway 122. Plurality of oil inlet passageways 126 arranged on circular collar 131 on the rotor 40 to connect facing rotor coolant inlet passageways 126. The passageway 122 on shaft eccentric 45 meets all of the facing oil inlet passageway 126 in turn, to provide oil for cooling the rotor. Fresh oil enters into rotor through passageway 126 by expelling equivalent amount of existing hot oil from rotor coolant chamber 125 through plurality of outlet oil passageway 127. Plurality of open passages 132 is provided on the engine housing. Said open passages 132 are connected to oil drainage passages 133 to drain back hot oil to a sump for redistributing after cooling through an oil-cooling radiator. Oil passageway 128 interconnects adjacent rotor coolant chamber 125 and bearings 76 for supporting apex seal holders 66 for lubrication of bearing surface of the bearing 76 as well as for lubrication of apex sealing elements 78, 79, 80.

As shown in FIG.18, Needle valve 135 and valve spring 137 are disposed into an enlarged bore 138 on the said oil passageway 128 to control oil discharge. Cam groove 140 is formed on the cylindrical surface of seal holders 66 to be engaged with the cam following end 136 of needle valve 135. Tapered end of needle valve 135 is operatively drawn outwardly from oil passageway 128 so as to permit oil discharge. Each seal holder operatively rotates about 50 degree in either direction about its own axis 139. Thereby the needle valve 135, following the cam groove 140, reciprocates between an open position (as shown in FIG.18) to allow oil discharge and a close position (FIG.18A) to stop oil escape from oil passageway 128. Spring 137 facilitates the needle valve to reciprocate. Being connected with said cam groove 140, channel 141 extends lengthwise on seal holder journal for distributing oil to entire bearing surface. Oil passage 142 interconnects channel 141 and groove 74 to lubricate oil scraper 80. Oil passage 143 and 144 connect between cylindrical seal holder surface and apex seal groove 73 and 75 respectively. During operation, small amount of oil from lubricated bearing surface enter into said oil passageway 143 and 144 and reach to groove 73 and 75 by means of centrifugal force. Groove 73 and 75 holds sealing elements 78 and 79, which are compression seals. Heat resistant wiping mat 145 is preferably provided to prevent oil seepage.

After lubricating the apex seals, excess oil deposits on the chamber peripheral wall 31 due to centrifugal force and said oil facilitates to skate the apex seals on said wall. During operation, the oil scraper 80 scrapes the excess oil off from the adjacent housing walls. As apex seals passing over the housing seal slot 85, a small amount of oil, carried by oil scraper outer edge, gets deposited into said housing seal slot 85 and on the tip of seal 81 due to centrifugal force. As seal 81 operatively extends against the working face of rotor, the deposited oil on the tip of said seal 81 provides lubrication to the sealing contact therebetween.

The above described configuration can be read with a split cycle rotary IC engine for diesel type fuel. Considerable reduction in the volume of said cavity formed on said working faces of said rotor of expansion-exhaust configuration Cl and a diesel type fuel injector, a diesel configuration would be appeared.

Referring to FIG. 19, another preferable embodiment of the present invention showing a compression ignition type alternative. A reasonable reduction in volume of the cavity on said rotor working face of expansion-exhaust configuration increasing the compression ratio to a value that is sufficient for diesel like combustion. Therefore, ignoring the presence of the sparkplugs the above described specification illustrates a diesel combustion engine.

Referring to FIG. 20, one of the alternative embodiments of the present invention illustrates a preferred multi-rotor modular arrangement 150 including two expansion-exhaust configurations CIa, CIb and two intake-compression configurations C2a and C2b. A rigid cuboid frame 151 having four perpendicular sides 151a, 151b, 151c and 15 Id. The expansion-exhaust configurations CIa and CIb are fitted on two opposing sidewalls 151a and 151c of frame 151 respectively and the intake-compression configurations C2a and C2b are fitted on two opposing sidewalls 151b and 15 Id of said frame 151 accordingly. The center shafts 44a and 44b of expansionrexhaust configurations CIa and CIb extends coaxially inwardly and the center shafts 54a and 54b of two intake-compression configurations C2a and C2b extends coaxially inwardly towards said cuboid frame center. A support 152 for supporting a closed assembly of four miter gears 154, 156, 158 and 160 is located at the center of said cuboid frame 151. Said miter gears 154, 156, 158 and 160 comprising respective axial shaft 155, 157, 159 and 161 extending outwardly from said support 152 for joining with center-shafts 44a, 54a, 44b and 54b respectively. Crossover passage 99a interconnects a first compressed air delivery passage of intake-compression configuration C2a and a first air inlet passage of expansion-exhaust configurations CIa and crossover passage 99b interconnects a second compressed air delivery passage of intake- compression configuration C2a and a second air inlet passage of expansion- exhaust configurations CIb. Crossover passages 99c and 99d interconnect intake- compression configuration C2b and expansion-exhaust configurations CIb and CIa accordingly. Each intake-compression configuration serves one alternate delivery to each adjacent crossover passages in a single rotation of center-shaft. Following the sequence thereunder, two simultaneous combustion events take place at symmetrically opposing combustion chambers of CIa and CIb near respective crossover passages 99a and 99c or at other pair of combustion chambers near crossover passages 99b and 99d accordingly. The center-shaft 54b of intake-compression configuration C2b is extended outwardly to serve as driveshaft.

As will be understood by those skilled in the applicable arts, various modifications and changes can be made in the invention and its particular form and construction without departing from the spirit and scope thereof. The embodiments disclosed herein are merely exemplary of the various modifications that the invention can take and the preferred practice thereof. It is not, however, desired to confine the invention to the exact construction and features shown and described herein, but it is desired to include all such as are properly within the scope and spirit of the invention disclosed and claimed.

Claims

ClaimsWhat is claimed is:

1. A split cycle rotary internal combustion engine comprising at least a pair of rotary configurations, said rotary IC engine including: a) at least one first rotary configuration being adapted to carry out expansion and exhaust phases of a four phase engine cycle, said first rotary configuration being provided with a coaxial center shaft, the first rotary configuration including: i. at least a first outer body having an inner chamber defined by trochoidal peripheral wall of multi-lobe profile enclosed by two facing sidewalls, means for initiating combustion, compressed-air inlet passages confined at predetermined positions near lobe junctions of said trochoidal peripheral wall for executing two successive combustion events in a single rotation of said center shaft, said first outer body having at least two exhaust gas outlet passages, said passages allow alternatingly the exit of combustion product, said first outer body further comprises a pair of external ring gears each of which coaxially confined on both of said facing sidewalls; ii. at least a first rotor body of polygonal profile having three apex portions and two sides, said first rotor body being mounted for relative rotation within said inner chamber of said first outer body, said first rotor body being rotatable around a lobe eccentrically integrated with said center shaft, said first rotor body having working faces extended between adjacent apex portions, said apex portions having apex seal arrangements, each containing at least a seal element to provide sealing cooperation with said trochoidal peripheral wall surface of said first outer body to define a plurality of working chambers between said rotor working faces and said first outer body, a cavity being formed on leading portion of each working face of said first rotor body for combustion, a pair of internal ring gears each of which being confined coaxially on both sides of said first rotor body for meshing with said external ring gears confined on the said facing sidewalls of first outer body for maintaining relative rotation between said first rotor body and said first outer body and also for transmitting driving force periodically to said center shaft; b) at least one second rotary configuration for carrying out intake and compression phases of a four phase engine cycle, said second rotary configuration being provided with a coaxial centre shaft, the second rotarjr configuration comprising: i. a second outer body with an inner chamber defined by trochoidal peripheral wall of multi-lobe profile enclosed by two facing side walls, at least two air intake passages each of which is provided for one intake in a single rotation of said centre shaft, at least two compressed air delivery passages being confined on said trochoidal peripheral wall of said second outer body, said second outer body further comprises a pair of external ring gears each of which coaxially confined on both sidewalls; ii. a second rotor body of polygonal profile having three apex portions and two sides, said second rotor body being mounted for relative rotation within said inner chamber of said second outer body, said second rotor body being rotatable around a lobe eccentrically integrated with said centre shaft, said second rotor body having working faces extending between adjacent apex portions, said apex portions having apex seal arrangements, each containing at least a seal element to provide sealing cooperation with said trochoidal peripheral wall surface of said second outer body to define a plurality of working chambers between said rotor working faces and said second outer body, said working chambers being variable in volume due to relative rotation of second rotor body and second outer body, a pair of internal ring gears each of which being confined coaxially on both sides of said second rotor body for meshing with said external ring gears on facing side walls of said second outer body for maintaining relative rotation between said second rotor body and said second outer body; c) at least a pair of air passages each of which being provided for interconnecting respective air inlet passage of said first outer body and respective compressed air delivery passage of said second outer body, each of said air passages including an inlet check valve and an outlet valve arrangement defining a pressure chamber there between, each of said inlet check valve connecting respective compressed air delivery passage of said second rotary configuration to permit one way flow of compressed air from said second rotary configuration to said pressure chamber and said outlet valve connecting respective air inlet passage of said first rotary configuration to permit one way flow of compressed air from said pressure chamber to said air inlet passage of said first rotary configuration, both the said air passages being provided for delivering compressed air once in each single rotation of centre shaft; d) a seal arrangement being provided adjacent to each lobe junction of inner chamber peripheral wall of both the said first outer body and the said second outer body for providing periodical sealing cooperation between a portion of each lobe junction and the adjacent working face of rotor so as to effectively divide said respective working chamber into leading and trailing working chambers, said seal arrangement includes at least a self retractable seal element and a matching groove to provide support for said seal element, length of said matching groove being substantially smaller than the breadth of respective peripheral wall defined by the distance of facing sidewalls, so as to provide a narrow track on both ends of said groove, said narrow tracks are provided for continuing the peripheral wall profile along the groove thickness and thereby facilitate passing of said apex seal element over said seal groove without fouling; e) means for engaging said seal elements before stalling fuel injection and for retracting said seal elements after completion of active combustion; and f) means for lubrication and cooling of internal moving components of both the said first and second rotary configurations.

2. A split cycle rotary internal combustion engine, as claimed in claim 1, wherein said means for combustion comprises: a) arrangements for injecting fuel for preparation of combustible charge; and b) means for initiating ignition in synchronization with the position of said cavity on said leading portion of respective working face of said first rotor body.

3. A split cycle rotary internal combustion engine, as claimed in claim 1, wherein said means for combustion comprises: means for injecting fuel for executing combustion.

4. A split cycle rotary internal combustion engine, as claimed in claim 1, wherein said compressed air inlet passages being further provided with multi-aperture profile to produce predetermined air flow pattern.

5. A split cycle rotary internal combustion engine, as claimed in claim 2 wherein said means for injecting fuel is adapted to start injection substantially on or after activation of said sealing engagement between said peripheral wall near lobe junctions of first outer body and said respective working faces of said first rotor.

6. A split cycle rotary internal combustion engine, as claimed in claim 2 wherein said means for initiating ignition is adapted to ignite in response to relatively center position with said cavity on the working face of respective rotor.

7. A split cycle rotary internal combustion engine, as claimed in claim 1 and claim 2 wherein said cavities on working faces of said first rotor body being preferably hemispherical in shape.

8. A split cycle rotary internal combustion engine, as claimed in claim 1 and claim 3 wherein said fuel injector completes the injection of fuel before the respective working face of rotor body reaches at its top center position.

9. A split cycle rotary internal combustion engine, as claimed in claim 1 wherein said meshing internal gears and external gears of each rotary configuration further comprises helical gears.

10. A split cycle rotary internal combustion engine, as claimed in claim 9 wherein a pair of said meshing internal and external helical gears are substantially of opposite hand gears to the other pair of meshing internal and external helical gears of the same rotary configuration.

11. A split cycle rotary internal combustion engine, as claimed in claim 1 wherein said apex seal arrangements further include swivel seal holders and bearing support for said swivel seal holders, each swivel seal holder having longitudinally extended holder face with plurality of parallel grooves to provide support for plurality of sealing elements, said grooves being substantially perpendicular with said holder face and longitudinally extended through the length of respective holder, said bearing support is provided by means of longitudinally opened matching bore formed on each apex portion of both the first and second rotor bodies, being journaled on said matching bores said seal holders are rotatable within the required angle in either direction about its own axis so as to keep the sealing elements perpendicular to adjacent peripheral wall surface of respective outer body.

12. A split cycle rotary internal combustion engine, as claimed in claim 11 wherein each swivel seal holder is a combination of at least two separate parts, fit axially slide-ably together being intermediated by a spring element to produce lateral loading of seal holder ends to provide corner sealing with the respective inner chamber sidewalls.

13. A split cycle rotary internal combustion engine, as claimed in claim 11 wherein said plurality of longitudinally extended parallel grooves on each seal holder is provided with preferably three parallel grooves.

14. A split cycle rotary internal combustion engine, as claimed in claim 13 wherein the end seal grooves of said three parallel grooves preferably contain compression seal elements.

15. A split cycle rotary internal combustion engine, as claimed in claim 13 wherein the intermediate groove of said three grooves supports an oil scraper element for scraping the excess oil off from said peripheral wall of respective inner chamber of respective outer body.

16. A split cycle rotary internal combustion engine, as claimed in claim 1, wherein said outlet valve arrangement of said air passage comprises an outlet valve disk and a spatially apart demobilizing disk on a common stem, both the outlet valve disk and demobilizing disk connecting said pressure chamber with their facing valve sides so as to cancel out the force exerted on each other by compressed air, said demobilizing disk further includes a seal means to prevent compression leakage, a switch valve providing a separation between said pressure chamber and a sub-chamber, said sub- chamber being extended to the opposite side of said demobilizing disk, the valve stem ends of both the outlet valve and the said switch valve being extend to outside through a pair of holes formed on said sub-chamber wall, where separate compression springs are employed with both the valve stems to keep said valves close, said switch valve opens periodically by a cam arrangement to allow compressed air from said pressure chamber to enter into said sub-chamber in order to make the demobilizing disk neutral.

17. A split cycle rotary internal combustion engine, as claimed in claim 16 wherein said seal means further comprise a pair of split end ring seal in a common seal groove, the first ring seal includes a common split-end type ring seal and the second ring seal providing radial step cut seal ends with overlapping steps to close radial opening, a minute bar is protruded on a side of the same ring at a position diametrically opposite to said step cut end, said minute bar fit loosely within said first ring end gap to prevent relative rotation there between, said first ring seal is facing the compression side.

18. A split cycle rotary internal combustion engine, as claimed in claim 16 wherein said switch valve further comprises pressure release vent through a portion of its stem to connect said sub-chamber with atmosphere when said switch valve is at its closed position.

19. A split cycle rotary internal combustion engine, as claimed in claim 1, wherein said self retractable seal elements further comprise a narrow elongated blank in it to hold spring means within said blank, being provided with a retainer means said spring means exert continuous pressure on said seal element to retract it.

20. A split cycle rotary internal combustion engine, as claimed in claim 1 wherein said means for engaging the self retractable seal elements comprises mechanical cam arrangement, spring biased cam followers, said spring being substantially stronger than the spring provided to retract said self retractable seal elements, said means for engaging the seal elements are disposed on outer body of both the first and second rotary configurations, said cam follower comprises at least an extended arm, mechanical linkages interconnects said seal elements and said extended arm of said cam follower for periodically extend said seal element against respective working face of respective rotor for dividing adjacent working chamber into two separate working chambers.

21. A split cycle rotary internal combustion engine, as claimed in claim 1 wherein said means for lubrication and cooling comprises plurality of oil passageways to provide lubrication and cooling for internal moving components, a main oil passageway extending axially through the center shaft, auxiliary oil passageways extends radially from said central oil passageway to provide pressurized oil to bearings, gears and internal cooling chambers of each rotor, an apex seal lubrication means comprises apex lubrication passage interconnecting said rotor cooling chamber and apex seal holder bearing on apex portions of each rotor body to provide lubrication to said bearing and seal holder, distribution passages on seal holder is provided for lubrication of entire bearing surface and for forcing oil by means of centrifugal force to apex seal grooves and inner chamber peripheral walls, valve mean is provided into said apex lubrication passage for controlling oil discharge.

22. A split cycle rotary internal combustion engine, as claimed in claim 21 wherein said valve means further including spring biased needle valve means, said needle valve means reciprocate within said apex lubrication passage between a close position to stop oil discharge and an open position to permit oil discharge momentarily by means of a cam profile formed on said swivel seal holder.

23. A split cycle rotary internal combustion engine, as claimed in any of the preceding claims wherein said intermediate oil scraper elements are capable of scraping excess oil off from adjacent peripheral wall and depositing a portion of said scraped off oil to said seal grooves for supporting said self retractable seal elements carried by said outer body, said deposited oil being carried thereafter by said self retractable seal tip lubricate the sealing contact with adjacent working face of respective rotor body.

24.A split cycle rotary internal combustion engine, as claimed in any of the preceding claims wherein said first rotary configuration further comprising a pair of similar rotary configurations being fitted on two opposing perpendicular sides of a rigid cuboid frame, said second rotary configuration further comprising a pair of similar rotary configurations being fitted on other two opposing perpendicular sides of the same cuboid frame, each of the first and second rotary configurations having individual . center shafts being extended towards the center of said cuboid frame, a closed assembly of four miter gears having individual axial shafts being extended outwardly to provide operative relation between all the rotary configurations, a support means for supporting said miter gears assembly is located at the center position of said cuboid frame, end portions of said extended center shafts of said first and second rotary configurations being joined with the corresponding ends of said axial shafts of said miter gears, each rotary configurations having a phase difference of 180 degrees with the facing similar configurations, said pair of air passages further comprise two pairs of air passages, each of said second rotary configurations being connected by a pair of air passages alternatingly delivers compressed air to one of the inlet passages of both of the first rotary configurations in a single rotation of center shafts,

25. A split cycle rotary internal combustion engine, as claimed in claim 24 wherein any one of the center shafts of the said second rotary configurations can be taken as the output shaft.