INTERNAL COMBUSTION ENGINE AND OPERATING CYCLE THEREFOR

Abstract: An internal combustion engine and method therefor includes a plurality of cylinders (CYL1...n) arranged parallel to and equally spaced from a longitudinally extending engine axis (LA) with a pair of opposed pistons (P1, P2) mounted in each cylinder for controlled reciprocation therein. Respective rotatably mounted axial-faced cams (C1, C2) and connected followers control the motion of the pistons through the operating cycle to permit at least one and preferably more power strokes from each cylinder during each rotation of the engine. A controllable divider valve (DV) is located between the opposed faces of the pistons to define, on the one side, a first air/fuel receiving compression chamber and, on the other side, a second relatively high-pressure ignition air receiving chamber. Before the cam-controlled compression stroke, the divider valve (DV) is closed so that the air/fuel mixture on the one side of the closed valve is compressed to a first pressure and the ignition air on the other side of the divider valve (DV) is compressed to a relatively higher pressure to cause the temperature of this more highly compressed air to increase to a temperature above the ignition temperature of the compressed air/fuel mixture. Ignition is caused by opening the divider valve (DV) to permit the high-pressure, high-temperature ignition air in the second chamber to enter the compressed air/fuel mixture in the first chamber and to cause burning thereof.

Description: INTERNAL COMBUSTION ENGINE AND OPERATING CYCLE THEREFOR BACKGROUND OF THE INVENTION Field of the Invention The present invention relates to internal com¬ bustion engines and, more particularly, to internal combustion engines of compact design having high thermal efficiency, very high power-to-weight ratio, high fuel economy, low pollutant emissions, constant horsepower capability, variable compression ratios and variable compression braking. Prior Art The predominant internal combustion engine config¬ uration presently used is the crankshaft/cylinder ar¬ rangement with a working piston reciprocates within a cylinder to drive a rotatable crankshaft. Variations of this configuration have included the in-line, V, radial, and horizontally opposed cylinder alignments. Regardless of the configuration, the piston reciprocates within its cylinder in accordance with one of two pre¬ dominant operating cycles, namely, the two- or four- stroke spark-ignition Otto cycle or the two- or four- stroke compression-ignition Diesel cycle. With the spark-ignition cycle, a homogeneous mixture of air and fuel at a preferred air/fuel ratio is compressed with ignition caused by an electrical spark or the equi- valent. In the compression-ignition cycle, fuel is injected into air that has been compressed to cause an adiabatic increase in its temperature to a temperature above the auto- or self-ignition temperature of the fuel. Both types of operating cycles and the various physical engine configurations that have been developed have proved satisfactory although each has attendant drawbacks. In the spark-ignition engine, the fuel must be pre- mixed with air to provide a desirably homogeneous mix- OMPI ι . WIPO ture with the ratio of the air to the fuel controlled so as to fall within a preferred ratio range, e.g., between 11:1 and 17:1. Air/fuel ratios greater than 17:1 result in mixtures which may or may not combust and air/fuel ratios below 11:1 result in mixtures which are ineffic¬ ient from the standpoint of fuel consumption and un¬ acceptable with regard to pollution. Additionally, the compression ratio of the spark-ignition engine is lim¬ ited to some maximum so as not to cause unintentional pre-ignition of the homogeneous air/fuel mixture during the compression stroke. The compression ratio limit also disadvantageously limits the thermal efficiency of spark-ignition engines. In contrast to the spark-ignition engine, the com- * pression-ignition engine utilizes air that has been heated during the compression stroke to a temperature greater than the auto-ignition temperature of the fuel so that fuel can be injected in a heterogeneous manner into the so-heated air to cause burning. * Thus, the fuel injected in a compression-ignition engine can be burned in considerable excess air to provide a rather large mass of heated air for the expansion stroke, but does not use all the available air in the cylinder for com¬ bustion. Accordingly, the compression-ignition engine provides a substantial increase in thermal efficiency compared to the spark-ignition engine. Unfortunately, compression-ignition engines require a rather sophis¬ ticated and expensive fuel delivery and injection system that mitigates against the increase in thermal effic- iency. In the traditional crankshaft/cylinder arrangement, regardless of the particular operating cycle imple¬ mented, the piston is usually at or near its uppermost position (TDC) during that portion of the cycle in which the fuel is ignited. In this configuration, the crank- OMPI shaft throw position presents a moment arm against which the piston acts that is the smallest of the operating cycle, and, accordingly, the mechanical conversion effi¬ ciency of the combustion products expanding against the piston during that portion of the cycle is poor. Additionally, the mechanical constraints of the traditional crankshaft/connecting rod arrangement is such that it is very difficult to vary the compression ratio of the engine, since the crankshaft throw distance and the connecting rod length are fixed and not easily changed. Because of this, the traditional crankshaft/- connecting rod arrangement is not amenable to variable compression schemes. As can be appreciated, a variable compression engine would permit the use of various types and grades of fuels and is particularly suitable in situations where the cost of various alternative engine fuels varies or where fuel supplies may be uncertain, such as in a military environment. In addition to the mechanical constraints of the crankshaft/connecting rod arrangement, the traditional cam operated valve train necessitates a fixed valve timing that is optimum for only a portion of the engine power curve. As can be appreciated, variable valve timing would permit opti¬ mization over the entire engine operating range. SU MARY OF THE INVENTION In view of the above, it is the primary object of the present invention, among others, to provide an en¬ tirely new and very much improved cycle for internal combustion engines in which air and various types and grades of fuel in a homogeneous mixture is ignited by compression ignition techniques at compression ratios much higher than those previously attained. It is also an object of the present invention to provide a new type of internal combustion engine in which cam-operated pistons produce high mechanical ef¬ ficiency and provide one, two, or more power strokes per revolution, which results in a very compact design and a high power-to-weight ratio. It is also an object of the present invention to provide an engine having a new and very much improved * operating cycle that provide ' s a heretofore unattainable thermal efficiency in which the air/fuel mixture's phys¬ ical and chemical delay periods are reduced, in which combustion occurs under constant volume and/or constant pressure conditions, and in which very high compression and expansion ratios are utilized. It is also an object of the present invention to provide a design which allows, contrary to prior de- signs, for the shortening of a piston stroke in such a way to cause a considerable increase in efficiency and thereby provide for substantial fuel economy. It is also an object of the present invention to provide an engine in which an additional increase in thermal and mechanical efficiency can be obtained by recovering the pressure and heat energy in the exhaust gas flow and the heat energy in the cooling system by compounding to provide a considerable increase in the engine power output. It is also an object of the present invention to provide an engine having means for varying the compres¬ sion ratios during engine operation and combine them with new operating cycles, an engine having a true mul- tifuel capability, and an engine with a low pollutant exhaust. It is also an object of the present invention to provide an engine by which valve timing can be varied during engine operation, to provide increased engine control and variable compression braking. It is also an object of the present invention to provide an engine that will eliminate expensive fuel delivery systems by mixing air and fuel in the manifold and, when combined with the new and unique engine cyc- les, will provide an engine that will operate equally well with a very rich air/fuel ratio (9:1) for maximum power or a very lean air/fuel ratio (150:1) to provide a large quantity of heated air for expansion. The result is an engine with a constant horsepower capability that will eliminate or reduce transmission requirements in many applications. The addition of compounded turbo- charging provides additional engine output power. In accordance with these objects, and others, the present invention provides for an internal combustion engine having at least one cylinder with first and sec¬ ond pistons located in the cylinder for controlled re¬ ciprocating motion by which the pistons advance toward each other to respective top dead center (TDC) positions and retract away from one another to respective bottom dead center (BDC) positions. A divider valve is located in the cylinder between the pistons to divide the por¬ tion of the cylinder between the piston faces into a first compression chamber defined between the first piston and the divider valve and a second compression chamber defined between the second piston and the div¬ ider valve. The motion of the pistons and the opening and the closing of the divider valve is controlled so that inlet air passes through the first chamber, through the open divider valve and into the second compression chamber and remains therein when the divider valve is closed and a mixture of air and fuel enters the first compression chamber. During the compression stroke the divider valve remains closed, and the motion of the pistons controlled so that the first piston compresses its air/fuel mixture in the first compression chamber defined between the first piston and the closed divider valve to a selected pressure and the second piston com¬ presses the air in the second compression chamber de- fined between the second piston and the closed divider valve to a pressure sufficient to cause an adiabatic increase in temperature thereof to a temperature greater than the self-ignition temperature of the compressed air/fuel mixture in the adjacent, first compression chamber. When the compression stroke of each piston is completed, the divider valve is opened to permit the high temperature compressed air in the second chamber to pass through the opened divider valve into the first compression chamber containing the air/fuel mixture to cause ignition thereof. The subsequently formed com¬ bustion products from the so-ignited air/fuel mixture in the first chamber then pass through the opened divider valve into the second chamber to further heat the air in the second chamber. The heated and expanded gases in both chambers cause the pistons to forcibly retract during the power stroke toward their respective bottom dead center (BDC) positions to provide an effective 'hyper' expansion of the combustion products as ex¬ plained below. • ^ JRE4 OMPI In one preferred embodiment of the present inven¬ tion, a plurality of working cylinders are aligned par¬ allel to and equi-spaced from a longitudinally extending engine axis with each cylinder containing a pair of working pistons and a divider valve to divide the re¬ spective cylinder into first and second compression chambers. Respective rotatibly mounted axial-faced cams are located on opposite ends of the cylinders with a cam follower/connecting mechanism coupling the pistons to their respective motion controlling axial-faced cams. The selected cam profiles cause the pistons to move through a compression stroke followed by ignition and a subsequent power stroke with an intermediate exhaust/- scavenge/inlet phase at the end of each power stroke, just prior to the following compression stroke. The exhaust gases produced during the combustion process can be provided to one or more turbine units for the purpose of extracting usable power from the exhaust gases. The turbine shaft can be connected to or otherwise drives the compressor units that pressurize the inlet air and/- or supply compressed air for pneumatic control valves. Additionally, the turbine shaft can be connected to the engine output shaft to provide engine compounding or a compounded turbocharger configuration so that useable power extracted from the exhaust gas flow above that needed to drive the inlet air -compressor units augments the engine power output. Fuel is preferably provided through electronically or mechanically controlled fuel injectors, although conventional carburetors of the shut-off type are suitable. In one form of the invention, the axial-face cam controlling those pistons that compresses the air/fuel mixture is selectively displaceable along the longi¬ tudinal axis of the engine to permit the compression ratio of the air/fuel mixture to be varied to optimize combustion parameters for a particular grade or type of fuel. Additionally, the axial-face cam profile is con¬ toured to advantageously allow for efficiency enhancing constant volume and/or constant pressure combustion during the initial portion of the power stroke. The divider valve that divides each cylinder into the air/fuel compression chamber and the ignition air compression chamber can take several forms. The valve, in one form, can include first and second mating valve disks, each having a plurality of sector-shaped through openings with one of the disks mounted for oscillating movement so that the openings of both disks are alter¬ nately in registration with one another to permit com¬ munication from one side of the valve to the other and ou t of registration with one another to effectively close the valve. In another form, an oscillating inter¬ mediate valve disk having plural sector-shaped openings can be positioned between two stationary valving disks having like sector-shaped openings with opening and closing of the valve controlled by movement of the os- cillatable intermediate valving disk. In another form, the divider valve can be configured as a sliding block valve. Because of the temperatures and pressures involved, the divider valve can be fabricated from those materials conventionally utilized for exhaust valve type appli¬ cations. In addition, ceramics are well suited for fabricating the divider valve. If preferred, any leak¬ age that occurs between the mating surface-to-surface interface of the valving surfaces can be directed through bypass passageways to the engine exhaust to thereby prevent leakage between chambers and potential pre-ignition. The oveable portion of the divider valve can be controlled in various manners including mechanically - y - through cams and cam followers, pneumatically utilizing pneumatic actuators, electrically, using electro-mag¬ netic solenoids or hydraulically using hydraulic actu¬ ators and by other control means. The present invention advantageously provides a new cycle for the compression-ignition of a homogeneous, rather than heterogeneous, mixture of air and fuel to thereby dispense with the expensive high pressure fuel injection arrangements previously utilized on compres- sion-ignition engines and direct-injected spark-ignition engines. In addition, the engine provides for the rapid heating of a relatively large mass of air and for a comparatively large expansion ratio that affords an opportunity to extract the maximum working energy from the gaseous combustion products and thereby optimize overall thermal efficiency. The use of engine com¬ pounding or a compounded tύrbocharger arrangement per¬ mits additional extraction of energy from the exhaust gas flow for both the purpose of pressurizing the inlet air and providing additional energy to the engine output shaft. BRIEF DESCRIPTION OF THE FIGURES The description of the figures, as well as the objects, features, and advantages of the present inven¬ tion will be more fully appreciated by reference to the following detailed description of a presently preferred, but none the less illustrative, embodiment in accordance with the accompanying drawing wherein: Fig. 1 is a schematic flat development of a two- stroke by two-stroke (2/2 stroke) operating cycle in accordance with the present invention; Fig. 2 is a schematic flat development of a four- stroke by two-stroke (4/2 stroke) operating cycle that includes a variable compression braking feature in ac¬ cordance with the present invention, Fig. 3 is a schematic flat development of a four- stroke cycle which includes a vacuum stroke by two- stroke (4V/2 stroke) operating cycle that includes a variable compression braking feature in accordance with the present invention; Fig. 4 is a partial, cross-section, isometric view of a preferred embodiment of an engine in accordance with the present invention with the selected parts thereof broken away, shown in schematic form, or omitted for reasons of clarity, and with selected transducers and actuators of a variable valve timing control system shown; Fig. 5 is a cross-sectional view of a preferred embodiment of an engine in accordance with the present invention; Fig. 6 is a detailed view of the inlet and divider valve and their actuators of the engine illustrated in Fig. 4; Fig. 7A is an isometric, detailed view of the upper valve cartridge assembly of the Fig. 5 engine; Fig. 7B is an exploded isometric, detailed view of the lower valve cartridge assembly of the engine shown in Fig. 5; ■ Fig. 8A is a cross-sectional view of a first alter- nate cam follower embodiment; Fig. 8B is a partial cross-sectional view taken along line 8B-8B of Fig. 8A illustrating the respective rolling contact points for the cam follower rolling member and pin follower arrangement; Fig. 8C is a cross-sectional view of a second al¬ ternative cam follower embodiment; Fig. 8D is a partial cross-sectional view of a third alternate cam follower embodiment; Fig. 8E is a cross-sectional view of a fourth al- ternate cam follower embodiment in accordance with the present invention; and Fig. 9 is a comparative table illustrating per¬ formance attributes of various engine configurations. fEAtT OMPI DESCRIPTION OF THE PREFERRED EMBODIMENT Figs. 1, 2, and 3 illustrate example operating cycles for the various engine embodiments described gen¬ erally below in relation to Figs. 4 and 5. The oper- ating cycles shown in Figs. 1, 2 and 3 are merely exem¬ plary of a wider group of possible operating cycles and variations thereof. Fig. 1 illustrates a 180° two- stroke by two-stroke (2/2 stroke) operating- cycle in which two complete operating cycles occur within one 360-degree revolution of the engine and which includes an efficiency-enhancing constant-pressure combustion feature. Fig, 2 illustrates a four-stroke by two-stroke (4/2 stroke) operating cycle in which two full operating cycles are also completed within one 360-degree revolu— tion of the engine and which includes a desirable con¬ stant-volume combustion feature and which can be mod¬ ified to include variable compression braking by modi- fing the valve timing. Fig. 3 illustrates a four-stroke cycle including a vacuum stroke by two-stroke (4 V/2 stroke) operating cycle varient that includes a constant volume/constant pressure combustion feature and which can be modified to include a variable compression brak¬ ing feature by varying valve timing and which is suit¬ able for transportion engine applications. In Fig. 1, circular axial-faced cams, Cl and C2, are represented in flat development with several views of an exemplary cylinder CYL illustrated between the opposed cam profiles of the cams. The cylinder CYL includes a divider valve DV in either an open or closed position; a piston PI located between the divider valve and the cam Cl profile that defines the air/fuel mixture chamber; and a piston P2 between the divider valve and the cam C2 profile that defines the ignition-air cham¬ ber. The piston PI is shown having a frusto-conical crown while the piston P2 is shown having a comple¬ mentary indented crown. The divider valve DV is sloped to conform to the clearance envelop defined between the two pistons when they are at their respective top dead center positions. Each of the pistons PI and P2 are connected to their respective cams Cl and C2 by cam follower mechanisms CFl and CF2 that are schematically shown as spherical rolling members positioned between the pistons and the respective cam surfaces. Both of the cams Cl and C2 are mechanically coupled together by a direct shaft connection for rotation in the same dir- ection. In Figs. 1, 2, and 3, the shafting and/or other mechanical coupling is not shown for reasons of clarity. The several views , a , .... , e l of Fig. 1 of the exemplary cylinder show the relative positions of the pistons Pi and P2 for various portions of the operating cycle. The intake for each cylinder, for both ignition air and the air/fuel mixture, is through circumferentially positioned inlet ports IP located adjacent the bottom dead center position of the piston Pi. The inlet ports IP are covered and uncovered by the motion of the pis¬ ton PI with the closed ports represented by the solid darkened fill. Cylinder exhaust is provided through circumferentially positioned exhaust ports EP that are located adjacent the bottom dead center position of the piston P2. The exhaust ports EP are covered and uncov¬ ered by the motion of the piston P2 with closed ports represented by the solid darkened fill. The two-stroke by two-stroke (2/2 stroke) operating cycle of Fig. 1 can be best appreciated by first consid¬ ering the portion of the operating cycle shown in posi¬ tion 'a' in which the pistons PI and P2 are at or near their respective top dead-center positions, the divider valve DV is open to allow communication across the valve, and the profiles of the cams Cl and C2 are at a raised crest portion to constrain the pistons Pi and P2 near their-top dead center positions to define a com¬ bustion chamber between the pistons. Just prior to that portion of the operating cycle shown in position 'a', and as described more fully be¬ low, the divider valve DV was opened to allow high- pressure air compressed by the piston P2 (for example, by a compression ratio of 24:1) to enter the air/fuel mixture compressed between the face of the piston Pi and the divider valve DV to cause ignition of the air/fuel mixture. The end-stroke position of the piston Pi is con¬ trolled by the profile of its respective cam Cl to com- press the air/fuel mixture by a selected compression ratio, such as 12:1. In general, the selected compres¬ sion ratio provided by the piston Pi must be sufficiently high to maximize the heat output of the particular fuel though not high enough to risk unwanted pre-ignition or detonation. The air compressed by the piston P2 is com¬ pressed to a pressure sufficient to cause an adiabatic rise in its temperature to a temperature that is greater than the ignition temperature of the compressed air/fuel mixture on the opposite side of the divider valve DV; a compression ratio of 24:1 being generally sufficient. Since the compression ratio increases dramatically as the distance or head space between the faces of the pistons PI and P2 and the divider valve DV decreases, the com¬ pression ratio can be accurately controlled by control- ling the end—stroke position of the pistons by control¬ ling the rise or 'lift' of the respective profiles of the cams Cl and C2. Additionally, as described more fully below in connection with Figs. 4 and 5, the entire cam Cl can be advanced or retracted along its axis of rotation to vary the compression ratio of the air/fuel cylinder to thereby provide a very desirable variable compression feature. As can be readily appreciated by those skilled in the art, a variable and controllable compression ratio permits the use of different types and grades of fuel. For example, different grades of diesel or gasoline fuels, as well as other gaseous, liquid or particulate fuels (including high-energy fuels), can be burned by merely varying the compression ratio for optimum effic¬ iency. When the divider valve DV is open, the high pres¬ sure, ignition-temperature air compressed between the divider valve DV and the piston P2 enters the compressed air/fuel mixture compressed between the face of the pis¬ ton PI and the divider valve DV and causes the ignition thereof as indicated by the sinuous lines in position a of Fig. 1. Since the divider valve DV has several cir¬ cumferentially and equi-spaced ports, .several ignition points or foci are started simultaneously causing a wave front that burns from the divider valve toward and to the face of the piston Pi. The wavefront initially advances under constant pressure conditions because of the elas¬ tic, unburned air/fuel mixture between the advancing wavefront and face of the piston PI. Since the air/fuel mixture is localized during burning between the divider valve DV and the face of the piston PI, the combustion process can be considered as a "layered" or "stratified" process. Because of the relatively large volume of air compressed by the piston P2, a substantial excess of oxygen is available to support the burning air/fuel mix- ture regardless of the air/fuel ratio selected or the compression ratio by which the air/fuel ratio was com¬ pressed. As the air/fuel mixture burns, the pressure in the combustion chamber rises with the hot combustion products then passing through the opened ports of the divider valve DV to heat and mix with the comparatively larger volume of air in the chamber between the open divider valve DV and the face of the piston P2. As can be appreciated by those skilled in the art, complete high pressure combustion takes place with a large volume of excess oxygen available to support combustion of the fuel and the subsequent oxidation of any un-burned or partly burned hydrocarbon molecules in the exhaust products. The excess oxygen, of course, is available to support any post-combustion processing of the exhaust gas including catalyst-prompted reactions requiring available oxygen. The above-described ignition sequence by which the divider valve opens to introduce ignition-temperature air into the air/fuel mixture results in a considerably re¬ duced ignition delay and an improved overall ignition sequence efficiency that is superior to both the conven¬ tional compression ignition and spark ignition engines. In the.conventional compression ignition engine, the fuel is injected under considerable pressure into the cylinder as discrete droplets of varying sizes and traveling in diverse directions throughout the combustion chamber. Because of this, a physical delay period occurs during which the various droplets undergo partial vaporization to be surrounded by a zone of fuel vapor, the temperature of which then rises to that of the ignition temperature combustion air. When the initial fuel vapors begin to burn, a chemical delay takes place, during which delay, the fuel vapor is consumed and during which the heat of combustion causes further vaporization of the remaining fuel droplets. As can be appreciated, some of the fuel begins burning while other of the fuel droplets still undergo fuel vaporization. As a result, the physical and chemical delay of conventional compression ignition en¬ gines can be quite extensive and not at all predictable. Conversely, in the spark ignition engine, the physical and chemical delay is less than that of the conventional compression ignition engines since- the air/fuel mixture is in a homogenous state just prior to ignition. How¬ ever, ignition occurs at a point source and spreads through the combustion chamber as an expanding wave front. In contrast to both the compression ignition and spark ignition engines, the ignition delay of the present invention is much reduced since the air/fuel mixture is in a homogeneous state just prior to ignition and ignition occurs, not as a single point as in spark ignition engines, but as multiple points in a common plane adjacent to the open divider valve to cause a burning wave front that advances from the open divider valve across the air/fuel mixture to the face of the piston Pi to cause the above-described layered or strati¬ fied combustion. Because of the comparatively larger volume of excess air provided by the ignition air cylinder, the heat of combustion per unit weight of fuel in the air/fuel cham¬ ber is required to raise the temperature of -a relatively larger mass of air provided from the ignition air cyl- inder so that the temperature differences between the beginning and the end of the expansion stroke will be relatively larger than in conventional compression igni¬ tion engines. As a result, an extension in efficiency is obtained over the efficiency of conventional compression ignition engines. This extended temperature difference obtained during the expansion stroke is facilitated, in part, by the shortened ignition delay described above. As the cams Cl and C2 rotate, as shown, to allow the pistons Pi and P2 to retract in response to the substan- tial gas pressures developed during combustion towards their respective bottom dead-center positions. At least a portion of one or both of the declining profiles of the cams Cl and C2 during the expansion stroke of the pistons PI and P2 may be configured to provide a highly desirable constant pressure combustion feature. Of course, the cam profiles are selected to minimize jump or change-over during the transition to the return profiles. As the pistons PI and P2 are forced downward toward their res¬ pective bottom dead-center positions during the power stroke, the piston forces are translated by the returning cam profiles into the desired rotary motion. As can be well appreciated, the moment arm presented to both of the working pistons Pi and P2 during their power stroke is constant to provide a substantial mechanical efficiency advantage over conventional reciprocating crankshaft/- connecting rod type engines in which the moment arm pres¬ ented to the working piston during the power stroke is smallest when the gas pressure is highest (that is, at or just after top dead-center) and, conversely, the moment arm presented to the piston is greatest when the piston is moving toward bottom dead-center and the gas pressure has diminished by virtue of the volumetric expansion. In the various engines described herein, the torque output can be readily increased or decreased by increasing or decreasing the radii of the axial-face cams Cl and C2. The pistons PI and P2 continue their movement toward their bottom dead-center positions to enter the exhaust portion of the cycle, shown in position 'b' of Fig. 1, during which the piston P2 uncovers the exhaust ports EP to vent a substantial portion of the combustion products therethrough as indicated by the long-dashed arrows. When the exhaust ports are initially uncovered by the piston P2, the piston PI is controlled by a short dwell portion of the cam Cl profile to permit the bulk of the combustion products to be exhausted and the internal cylinder pressure to drop to a level low enough to permit cylinder scavenging. At the end of this short dwell period, the cam Cl profile returns to cause the piston PI to uncover the inlet ports IP to begin cylinder scaveng- ing as shown in position 'c' of Fig. 1 and as indicated - ORE OMPI by the short-dashed arrow swirls. By the time the cyl¬ inder inlet ports IP are uncovered, the internal cylinder pressure has dropped to a pressure less than that of the pressurized inlet air which enters the cylinder through the now-uncovered inlet ports IP to sweep remaining ex¬ haust gas products along the cylinder CYL, through the opened ports of the divider valve DV and then through the open exhaust ports EP to effectively remove the residual exhaust gas products and recharge the entire cylinder CYL with fresh inlet air, that is, the volume present between the divider valve and the piston Pi and the divider valve and the piston P2. As the inlet air sweeps through the divider valve, desirable divider valve cooling takes place with the heat returned to the cylinder CYL2. Upon completion of the scavenging portion of the operating cycle, as shown between positions 'c' and 'd' of Fig. 1, the divider valve DV closes ' and, substantially simultaneously therewith, the profile of the cam C2 rises to cause the piston P2 to advance, cover, and thereby close the exhaust ports EP so that the chamber between the now-closed divider valve and the face of the piston P2 contains a fixed charge of fresh air for subsequent compression as piston P2 continues to advance as des¬ cribed below. The position of the piston Pi is con- trolled so that the inlet ports IP remain open a short time after the divider valve DV is closed. During this time, fuel is injected into the inlet air by an appro¬ priately controlled fuel injector through the still- opened inlet ports in the chamber defined between the face of the piston PI and the closed divider valve DV. The quantity of fuel injected is sufficient to provide a desired air/fuel ratio. The air/fuel ratio selected is not limited to a narrow range as in the case of spark- ignition engines but can vary over a very wide ratio range, for example, from a low of 9:1 to an extremely high air/fuel ratio. The upper limit of the higher air/- fuel ratio is determined by the power required to over¬ come mechanical friction and to compress the combustion and ignition air. Just subsequent to the fuel injection, the profile of the cam Cl changes over to an inclined rise to cause the piston PI to advance to cover and thereby close the inlet ports IP as shown to the right of position 'd' of Fig. 1. At this time in the operating cycle, the divider valve DV remains closed and the cham- ber defined between the closed divider valve and the face of the piston P2 contains a charge of fresh air with minimum exhaust products, if any, and the chamber defined between the closed divider valve and the face of the piston Pi contains an air/fuel mixture of selected ratio. The profile of the cam Cl continues to advance the piston Pi at position 'e' of Fig. 1 to compress its air/- fuel mixture by a selected compression ratio, such as 12:1, this compression ratio being sufficient for use with most gasolines to effect maximum heat extraction without causing auto-ignition. The compression ratio could be lowered for lower octane gasolines and increased for certain diesel or kerosene type fuels. Simultan¬ eously, the profile of the cam C2 continues to advance piston P2 toward its top dead-center position to compress its air charge by a selected compression ratio, such as 24:1; this compression ratio being suitable to cause an adiabatic rise in the temperature of the so-compressed air to a temperature greater than the ignition temper¬ ature of the compressed air/fuel mixture located on the opposite side of the closed divider valve DV. At the completion of the compression stroke, both of the pistons Pi and P2 are at their respective top dead- center postions. At a moment shortly before or after the pistons PI and P2 attain their top dead-center positions, the divider valve DV is opened to allow the hot, igni- tion-temperature and higher pressure air compressed bet¬ ween the divider valve DV and the face of the piston P2 to pass through the now-opened divider valve and enter and mix with the compressed air/fuel mixture to cause ignition thereof and repeat the cycle as described above. In Fig. 1, the aforedescribed operating cycle occurs in 180° of cam rotation so that two full operating cycles occur within one 360° rotation of the engine. As can be readily appreciated, the cam profiles of the cams Cl and C2 can be readily changed so that more or less operating cycles can take place during each rotation of the engine. Fig. 2 illustrates, in a manner similar to the flat development layout of Fig. 1, the operating cycle of a four-stroke by two-stroke (4/2 stroke) cycle in accord- ance with the present invention. Fig. 2 differs, in part, from Fig. 1 in using flat-topped pistons and in that the inlet ports IP.of the Fig. 2 cylinder are lo¬ cated adjacent to the divider valve DV with inlet valving controlled by a rotatably mounted inlet valve IV mounted on the exterior of the cylinder CYL. The inlet valve IV, as described in more detail below, is rotatable between two positions and includes holes that alternately reg¬ ister with the inlet ports IP to effectively open the inlet ports and also offset from the inlet ports to ef- fectively close them. The operating cycle of Fig. 2 can be best understood by first considering position 'a' of Fig. 2 in which both pistons PI and P2 are positioned at or adjacent to their top dead-center positions with the divider valve DV open. Just prior to the components having attained this pos¬ ition, the divider valve DV had been opened to allow compressed ignition temperature air to pass from the chamber defined between the divider valve DV and face the piston P2 through the open ports thereof into the com- pressed air/fuel mixture in the chamber defined between the open divider valve DV and the face of the piston Pi with ignition in position 'a' of Fig. 2 indicated by the sinuous lines. The pistons PI and P2 are held in the position shown in position 'a' of Fig. 2 by dwell portions of their re¬ spective cams Cl and C2 to provide the desirable and efficiency enhancing attributes of constant volume com¬ bustion. The advantages of constant volume combustion are outlined, for example, in Degler, H.E., "Internal Combustion Engines", John Wiley & Sons, 1938, pp. 77 to 98, and Obert, E.F. "Internal Combustion Engines and Air Pollution", Harper & Row, 1973, pp. 166 to 175. Since the pistons Pi and P2 are held in their respective posi¬ tions adjacent to or near their respective top dead- center positions to define a constant volume combustion chamber, the air/fuel mixture burns in a desirable and an advantageous constant volume conditions, in contrast to other types of reciprocating engines in which the piston is forced towards its bottom dead-center position during the combustion process to cause a moderation in the de¬ velopment of the maximum cylinder pressure and, often¬ times, a quenching of the combustion process that leads to less than optimum operating efficiency and unburned combustion products in the exhaust gases. As the air/- fuel mixture burns, the pressure in the now-constant volume combustion chamber rises with the hot combustion products then passing through the opened ports of the divider valve DV to heat and mix with the comparatively larger volume of air in the chamber between the. open divider valve and the face of the piston P2. As the combustion process takes place, both of the cams Cl and C2 rotate until their respective profiles decline to allow the pistons PI and P2 to begin retrac¬ tion toward their respective bottom dead-center positions in response to the increase in the gas pressure developed during the combustion phase. As the pistons PI and P2 retract, the piston forces are imparted to the returning profiles so that cams Cl and C2 are caused, to rotate. The pistons PI and P2 continue their power stroke toward their respective bottom dead-center positions at which point the piston P2 uncovers the exhaust ports EP in position 'b' to cause the expanded relatively lower pressure combustion products to exit the cylinder. Con¬ currently with the uncovering of the exhaust ports EP, the profile of the cam Cl rises to advance the piston Pi toward its top dead-center position, as shown in position 'b* of Fig. 2, to provide a piston-assisted exhaust ram or 'push-out' movement. The amount of the piston Pi advance is, of course, controlled by the rise profile of the cam Cl but generally the piston Pi should be at least advanced to or adjacent to the inlet ports IP as shown in position 'c' of Fig. 2. When the piston Pi reaches the maximum extent of its exhaust removal assist, the cam profile for the cam Cl changes to a short dwell, and a short scavenge phase takes place. During the scavenge phase, which takes place with the components positioned as shown in position *c' of Fig. 2, the inlet ports IP are opened by appropriate control of the rotatable inlet valve IV to allow fresh pressurized inlet air into the chamber defined between the face of the now- dvanced piston Pi to sweep through the open ports of the divider valve DV through the chamber defined between the divider valve DV and the face of the piston P2 and out through the still uncovered exhaust ports EP. The air flow through the cylinder during the scavenge phase of the cycle assists in cooling the divider valve and returning heat to the cylinder CYL2 and cooling the faces of the pistons Pi and P2. Upon completion of this piston-assis¬ ted exhaust/scavenge phase of the cycle, the profile of the cam C2 changes to a rising profile to cause the pis- ton P2 to advance and thereby cover the exhaust ports EP OMPI and the divider valve DV is closed as shown between posi¬ tions 'c' and 'd* of Fig. 2. In this way, the chamber defined between the now-closed divider valve DV and the now-covered and closed exhaust ports EP is charged with fresh air for subsequent compression to a temperature above the ignition temperature of the air/fuel mixture to be combusted. Concurrently with the compression of the fresh air charge in the chamber between the divider valve DV and the face of the piston P2, the profile of the cam Cl declines and the piston Pi retracts to effect a volu¬ metric expansion and pressure reduction to aspirate an air and a fuel charge through the opened inlet ports IP as indicated by the solid swirling lines in position 'd' . The air/fuel mixture can be introduced into the cylinder by normal intake stroke aspiration, or preferably, under pressure as supplied through a super- or a turbo-charger. In addition, the fuel can be premixed outside the cyl¬ inder, for example, by carburation or by injecting fuel from a fuel injector into the inlet air flow prior to the inlet air passing through the opened inlet ports in the combustion chamber. The inlet air flow channels, the inlet ports, and the combustion chamber are configured to provide a swirl pattern to swirl the air/fuel mixture in the combustion chamber and to provide a turbulence in the flow which creates an increase in the air/fuel mixing efficiency, and decreases the physical delay period. The decrease in the physical delay period assists in de¬ creasing the chemical delay period and thereby shortens the overall ignition delay period. In addition to mixing the fuel outside the combustion chamber, fuel can be injected directly through the opened inlet ports into the combustion air/fuel chamber itself where the turbulent swirl pattern effects the air/fuel mixing. Once the air and fuel mixture is introduced into the cylinder, the inlet valve IV is closed between positions 'd' and 'e' of Fig. 2. The cam profiles of the cams Cl and C2 in position 'e' respectively rise to cause the piston Pi to compress the air/fuel mixture between it and the closed divider valve DV by a selected compression ratio, such as 12:1. The compression ratio, as discussed above, is selected to effect efficient utilization of the heat energy in the fuel while not high enough to risk undesirable pre-igni- tion. Concurrently, the piston P2 compresses the fresh air charge in the chamber defined between the piston P2 and the closed divider valve DV by a selected compression ratio, such as 24:1; this compression ratio being suf- ficient to cause an adiabatic rise in the temperature of the so-compressed air to a temperature greater than igni¬ tion temperature of the less highly compressed air/fuel mixture on the opposite side of the closed divider valve DV. Since small changes in the head space defined by the respective top dead-center positions of the pistons Pi and P2, in light of their relative strokes, can cause meaningful changes in the compression ratio, the lifting or rising profiles of each of the cams Cl and C2 can be easily controlled to control the respective compression ratios. Also, and as described more fully below in re¬ lationship to the embodiments of Figs. 4 and 5, the en¬ tire cam Cl can be advanced or retracted along its axis of rotation to change the compression ratio to accom- modate different grades and types of fuels. When both the pistons Pi and P2 achieve their ulti¬ mate top dead-center positions, the divider valve DV is opened to allow the relatively high pressure compressed air in the chamber between the divider valve and the piston P2 to pass through the open divider valve DV ports into the less highly compressed air/fuel mixture and cause ignition thereof. Ignition foci develop adjacent to the opened divider valve DV to cause a wave front that burns from a zone adjacent to the opened divider valve DV toward and to the face of the piston PI to effect the aforedescribed "layered" or "stratified" combustion, this ignition being represented in position 'a 1 of Fig. 2 by the sinuous lines. The combustion process takes place under desirable constant volume conditions as established by the cam profiles of the cams Cl and C2 to thereby allow the pres¬ sure within the cylinder to rise. Thereafter, the pro¬ files of the cams Cl and C2 both decline to allow ex¬ pansion during the power stroke as described above. The operating cycle shown in Fig. 2 is completed in 180° of engine shaft rotation with two full cycles com¬ pleted during one full revolution of the engine; as can be appreciated, the cam profiles can be modified to per¬ mit a greater or lesser number of cycles per engine revo- lution. The operating cycle of Fig. 2 includes a desirable compression braking mode that is achieved by controlling the divider and inlet valve timing. As shown in position 'a' of Fig. 2, the compression braking cycle begins with the pistons Pi and P2 at their respective top dead center positions and the divider valve and inlet valve in their opened positions. As the cams Cl and C2 rotate, both the pistons PI and P2 retract toward and to their respective bottom dead center positions and aspirate a fresh air charge through the opened inlet valve ports. The profile of the cam Cl is such that the piston Pi advances towards its top dead center position while the piston P2 remains adjacent its bottom dead center position as indicated at positions 'b' and 'c* of Fig. 2. When the piston PI reaches its top dead center position, the divider valve is closed and the piston P2 begins its compression stroke to compress its air charge between it and the now-closed divider valve. The profile of the cam Cl declines caus¬ ing the piston PI to retract toward and to its bottom dead center position to re-aspirate an air charge. When the piston PI attains its bottom dead center position, the inlet valve is closed and thereafter, the rising profiles of the cams Cl and C2 cause the pistons PI and P2 to advance toward and to their respective air charges. When the pistons Pi and P2 attain their respective top dead center positions, as illustrated in position 'e' of Fig. 2, the divider valve and inlet valve are opened to release the now-compressed air. A more detailed des¬ cription of the variable compression features is pre- sented below with regard to the embodiment of Fig. 4. As can be appreciated, the compression braking mode is variable; that is, the amount of compression braking can be controlled by controlling the opening and closing timing of the divider valves and inlet valves. Thus, the compression braking effect can be modified by opening the divider valve and inlet valve at some time prior to the pistons Pi and P2 attaining their top dead center pos¬ itions as illustrated in position 'e' of Fig. 2 and also by closing the divider valve and/or the inlet valves prior to a full air charge entering the cylinder for subsequent compression by the pistons PI and P2. Fig. 3 illustrates, in a manner similar to the flat development layouts of Figs. 1 and 2, the operating cycle of a four-stroke cycle which includes a vacuum stroke by two-stroke (4V/2 stroke) operating cycle in accordance with the present invention. The operating cycle and components shown in Fig. 3 differ from those shown in Fig. 1 in that a variable compression braking feature is provided, which braking feature is desirable for engines used in transporatation applications, pistons having frus- to-conical crowns are used, and the divider valve is de¬ fined by a sliding block type valve having a single aper¬ ture, hole, or slot. The operating cycle of Fig. 3 can be best understood by first considering position 'a* of Fig. 3 in which both pistons PI and P2 are positioned at or adjacent to their respective top dead center positions with the divider valve DV open. Just prior to the components having at¬ tained the position illustrated in position 'a' of Fig. 3, the divider valve DV had been opened to allow com¬ pressed ignition-temperature air to pass from the chamber defined between the divider valve DV and the face of the piston P2 through the open port of the divider valve into the compressed air/fuel mixture in the chamber defined between the open divider valve DV and the face of the piston PI to cause ignition as indicated by the sinuous lines. The pistons PI and P2 are held in the positions shown in position 'a' of Fig. 3 by dwell portions of their respective cams Cl and C2 to provide the desirable and efficiency enhancing attributes of constant volume and constant pressure combustion. Thereafter, and as the combustion process takes place, both of the cams Cl and C2 advance through a short constant volume combustion section until their respective profiles begin to decline through a constant pressure combustion section to allow the pistons PI and P2 to retract toward their respective bottom dead center positions in response to the increase in gas pressure developed during combustion. As the pistons Pi and P2 retract, the piston forces are imparted to the respective returning or declining profiles to cause the cams Cl and C2 to rotate. The pistons PI and P2 continue their power stroke toward their respective bottom dead center positions at which point, as shown in position 'b' of Fig. 3, the piston P2 uncovers the ex- haust ports EP to cause the expanded, relatively lower pressure combustion products to exit the cylinder CYL. A scavenging fan SF or the equivalent is provided to assist in removing and partially evacuating the exhaust gases from the chamber while the inlet ports IP remain closed by the piston Pi as controlled by a short dwell period on the cam Cl. As shown in Fig. 3, a gas turbine GT can be positioned downstream from the scavenging fan SF to ex¬ tend the expansion stroke. The cam Cl profile then re- turns to allow the piston PI to uncover the inlet ports IP to begin cylinder scavenging between positions 'b' and *c' of Fig. 3. By the time the inlet ports IP are un¬ covered, the internal pressure of the partially evacuated cylinder has dropped to a level less than that of the inlet air which enters the cylinder through the now- uncovered inlet ports IP to sweep the remaining exhaust gas products along the cylinder CYL, through the opened ports of the divider valve DV, and then through the ex¬ haust ports EP to effectively remove the residual gas products from the cylinder and recharge the entire cyl¬ inder with fresh inlet air which also serves to cool the divider valve DV. The flow of inlet air cools CYL1, the divider valve DV, returns heat to the cylinder CYL2, and, in addition, cools the crowns of the pistons Pi and P2. Upon completion of the scavenging portion of the operating cycle, just to the right of position 'c' of Fig. 3, the profile of the cam C2 rises sufficiently to cover and close the exhaust ports EP and then enters a dwell portion while the profile of the cam Cl changes to a rising profile to advance the piston PI toward the still-open divider valve to effect compression of the fresh inlet air charge in the chamber defined between the open divider valve and the face of the piston P2. This "ram precompression" stroke is completed when the piston PI attains the crest of the rising portion, which crest - IA OMPI occurs just after the position 'd' of Fig. 4, at which point the divider valve DV is closed. The ram precom- pression stroke effectively supercharges the cylinder between the divider valve DV and the face of the Piston P2. Thereafter, and as shown in position 'e τ of Fig. 3, the profile of the cam Cl declines to retract the piston Pi to develop a partial vacuum in the chamber between the face of the piston PI and the now-closed divider valve DV. Concurrently, the profile of the cam C2 rises to cause the piston P2 to advance toward the closed divider valve and compress its fresh air charge by a selected compression ratio, such as 24:1. As shown just prior to position 'f* of Fig. 3, when the piston PI reaches its bottom dead center position, the inlet ports IP are uncovered and inlet air, either at ambient atmospheric pressure or pressurized above ambient atmospheric pressure, is aspirated into the chamber be¬ tween the face of the piston PI and the closed divider valve DV with the partial vacuum developed by the piston Pi during its retraction stroke of position 'e' of Fig. 3 assisting in the aspiration of the air charge. Fuel may be injected into the inlet air and through the uncovered inlet ports IP during aspiration, or fuel may be premixed in a carburetor with the inlet air during aspiration. After the air and fuel enter the chamber between the face of the piston PI and the divider- valve DV, the pro¬ file of the cam Cl rises to advance the piston PI toward its top dead center position to effect compression of the air/fuel mixture by a selected ratio, such as 12:1. While the piston PI is advancing toward its top dead center position, the piston P2 continues advancing toward its respective top dead center position to compress the igni¬ tion air charge and cause an adiabatic temperature in¬ crease to a temperature greater than the ignition temp- erature of the air/fuel mixture in the adjacent chamber. When both the piston Pi and the piston P2 attain their respective top dead center positions, the divider valve DV is opened to cause hot ignition-temperature air from the chamber defined between the face of the piston P2 and the divider valve DV to pass into the compressed air/fuel mixture defined between the face of the piston Pi and the divider valve. The ignition air, as described above, causes an ignition foci which develops a burning wave front that advances from the divider valve toward the the piston PI. The burning and expanding combustion gases in the aforedescribed layered or stratified combustion zone between the piston Pi and the opened divider valve than pass through the open divider valve to heat the ignition air charge and cause both pistons PI and P2 to retract during their respective expansion strokes. The operating cycle of Fig. 3 can be utilized, by varying the timing of the divider valve, to achieve vari¬ able compression braking. The compression braking cycle is initiated with the pistons PI and P2 at their respect¬ ive top dead center positions and the divider valve open¬ ed as shown in position 'a' of Fig. 3. The profiles of the cams Cl and C2 decline to cause each piston Pi and P2 to retract toward and to their respective bottom dead center positions, the piston P2 first uncovers the ex¬ haust ports with the piston Pi thereafter uncovering the inlet ports. When the inlet ports and the exhaust ports are uncovered, a fresh air charge is introduced into the cylinder. The profiles of the cams Cl and C2 then rise to cause the piston P2 to advance a short distance to close the exhaust ports. The piston Pi then advances toward its top dead center position to provide a ram assisted charging of cylinder CYL2 defined between the face of the piston P2 and the divider valve. When the piston PI attains its top dead center position, just after position 'd' in Fig. 3, the divider valve DV is closed and the profile of the cam C2 rises to cause the piston P2 to compress the ram-assisted charge while the piston PI then retracts, as shown in positions 'e* and 'f, to its bottom dead center position to uncover the inlet ports. When the piston Pi uncovers the inlet ports, the divider valve is opened to vent the air com¬ pressed by the piston P2 through the inlet ports IP. During the compression braking mode, the quantity of air admitted to the cylinder can be controlled by throttling the air flow. A more detailed description of the vari¬ able compression feature is provided below in relation to the embodiment of Fig. 4. The table of Fig. 9 presents four sets. Groups I, II, III, and IV, of physical dimensions and related cylinder pressure information for exemplary air/fuel and ignition air cylinders. The information presented in Fig. 9 relates to -the volumetric efficiency of the air/- fuel and ignition air cylinders as each compresses their respective air charges and demonstrates how shortened air/fuel strokes for a set of otherwise fixed physical dimensions can increase overall volumetric efficiency of the engine. Within each grouping, the only working vari¬ able is the stroke length of the air/fuel piston with the remaining information representing the consequential result of the stroke length change of the air/fuel piston. The information presented in each column of the table of Fig. 9 is defined as follows: CYLINDER 15 psia A/F Air-fuel cylinder charged with 15 psia at the begin- ning of the compression stroke. Air Ignition air cylinder charged with 15 psia at the beginning of the compression stroke. STROKE INCHES A/F Air-fuel cylinder piston stroke, varying in each group from 3" to 1". Air Ignition air cylinder piston stroke, remains a constant (3") in all four groups. COMPRESSION RATIO A/F Air-fuel cylinder compression ratio, a constant in each group, varies from group to group. Air Ignition air cylinder compression ratio, a constant in groups I, II and III, changes to a different value in group IV. CLEARANCE SPACE A/F Clearance or head space in air-fuel cylinder at the end of the compression stroke before the divider valve opens. Air Clearance or head space in the ignition air cylinder at the end of the compression stroke before the divider valve opens. TOTAL STROKE The combined strokes of air-fuel cylinder piston and ignition air cylinder piston; can be considered as the total expansion stroke after the divider valve opens. CLEARANCE SPACE' The combined clearance space between air-fuel cyl¬ inder piston and the ignition air cylinder piston after the divider valve opens. EXPANSION RATIO The figures in this column can be considered as a final composite compression ratio. As can be appreciated, for example, gasoline can be burned between ratios of 12:1 to 20:1. CYLINDER PRESSURE A/F Air/fuel cylinder pressure (with no heat added) before the divider valve opens. Air Ignition cylinder pressure (with no heat added) before the divider valve opens. PRESSURE FACTOR The pressure factor column indicates the quanti- tative pressure contribution of the air/fuel and ignition air cylinders to the FINAL PRESSURE column. The pressure factor is obtained for a cylinder by multiplying the cylinder pressure by a ratio of the clearance space for that cylinder divided by the combined CLEARANCE SPACE' . For example and as illustrated by the calculations below the Group IV information, the first and second rows of information in Group I indicate the air/fuel and ignition air cylinders have clearance space of 0.375" and 0.125", respectively, for a total CLEARANCE SPACE 1 of 0.500". The air/fuel cylinder contributes 75% (0.375/0.500) of the total CLEARANCE SPACE' while the ignition-air cylinder contributes 25% (0.125/0.500). Since the air/fuel cylinder provides a pressure of 120 psi prior to the divider valve opening, the air/fuel cylinder provides a contribution of 90 psi (75% x 120) and the ignition air cylinder provides a contributio - of 90 psi (25% x 360) to the final pressure of 180 psi. FINAL PRESSURE Composite pressure, the result of combining the clearance space pressures of the air-fuel and the ignition cylinders after the divider valve is opened. % GAINED ( ) The figures enclosed in parentheses indicate the percentage of pressure gain within each group (I, II, III and IV) in which the A/F cylinder stroke is the only working variable. Note that in group I a change in the A/F stroke from 3" to 1" increased the final pressure from 180 psia to 240 psia, an increase of 33.3%. [ ] The figures enclosed in brackets indicate the per¬ centage of pressure gain over an 8:1 compression ratio with a cylinder pressure of 120 psia. Note in group III in the FINAL PRESSURE column the figure 288 (psia) represents a pressure increase of 140% (all pressure figures with no heat added). It should be noted that the inventive cycle provides very large gains in final pressures and temperatures just before combustion, resulting in extremely high thermal efficiency. The addition of heat energy will provide a considerable increase in efficiency. The final pressure developed in a cylinder is a combination of the pressure contributions of the air/- fuel and ignition air cylinders and is a representation of the composite compression ratio and the composite expansion ratio. The final pressure can be varied for an application by varying those physical parameters that affect the cylinder pressure in the air/fuel and the ignition air cylinders and the pressure contribution each of these cylinders makes to the final pressure. Lengthening or shortening of the stroke of either or both of the air/fuel and/or ignition air cylinders will change the pressure developed in each of these cylinders and the pressure contribution of each to thereby change the final pressure. In the examples presented in Fig. 9, the air/fuel cylinder stroke was changed to effect a change in the final pressure. As can be appreciated by those skilled in the art, the stroke of the ignition air cylinder could have been likewise and with equal conven¬ ience changed to effect a similar change in a final pressure. In some applications, for example, it may be desirable to have a relatively long air/fuel cylinder stroke and short ignition air stroke. The various engine operating cycles described above as well as other cycles and variations thereof can be utilized in various engine configurations including the engine embodiments presented in Figs. 4 and 5. One OMPI _ embodiment of an internal combustion engine in accord¬ ance with the present invention is shown in Fig. 4 and referred to therein by the reference character 10, and one preferred structural embodiment of an engine in accordance with the present invention as shown in Fig. 5. The engine 10 (Fig. 4) is defined about a longi¬ tudinally extending axis LA and includes a plurality of cylinders CYLl...n aligned parallel to the longitudinal axis and equi-spaced about the longitudinal axis at a selected radius so that the cylinders CYLl...n are posi¬ tioned in a manner analogous to the cylinder of a revol¬ ver. Each cylinder CYLl...n includes an inlet and div¬ ider valve cartridge assembly, generally referred to by the reference character VC, that divides the cylinder into a first cylinder half CYLl and a second cylinder half CYL2. The cylinder halves CYLl and CYL2 are de¬ signed to receive a respective air/fuel charge and an ignition air charge, as described above in ' relation to Figs. 1, 2, and 3. The cylinder halves CYLl are de- signed as insertable sleeves that are carried in a left¬ side engine block section while the cylinder halves CYL2 are likewise designed as insertable sleeves and carried in a complementary right-side engine block section that is preferably bolted to the first engine block section with the valve cartridges VC retained in place in ap¬ propriately sized cartridge receiving cavities. In Fig. 4, various connecting bolts, pilot pins, and/or piloting surfaces for mating the respective engine block sections are not shown. As can be appreciated, the split-block construction, the cartridge nature of the valving, and the insertable cylinder sleeves provide an engine that can be readily assembled and disassembled. A piston PI is located on the one side of the valve cartridge VC in the cylinder half CYLl for compressing an air/fuel mixture between it and the valve cartridge VC while a piston P2 is located on the opposite side of the valve cartridge VC in the cylinder CYL2 for compres¬ sing an ignition air charge between it and the valve cartridge VC. The pistons Pi and P2 advance towards one another toward and to respective top dead center (TDC) positions and retract away from one another toward and to respective bottom dead center (BDC) positions under the control of respective axial-face cams Cl and C2 which each have rising, declining, and dwell profiles that control the engine operating cycles as described above in more detail in connection with Figs. 1, 2, and 3. The cams Cl and C2 are each supported for rotation about the longitudinal axis LA by back plates BP, only one of which is shown in Fig. 4. Each of the cams Cl and C2 carry external ring gears RGl and RG2 (as schem¬ atically shown in Fig. 4) located on the peripheral surface thereof. Pinion gears PGl -and PG2, secured to an output shaft OS, engage the ring gears RGl and RG2 " of each cam Cl and C2 and rotate the output shaft OS in response to the rotation of the cams Cl and C2. In , operation, and as explained in more detail below, the pistons PI and P2 reciprocate in their respective cyl¬ inder halves CYLl and CYL2 to cause the cams Cl and C2 to rotate, which rotation is transferred to the output shaft OS. One or more power take-off devices PTO, including engine-operated accessories, can be mounted on the en¬ gine and driven by the ring gears RGl or RC2. A com¬ pounded turbocharger assembly is located in the central - or 'core' portion of the engine and is generally desig¬ nated by the reference character CTA. The compounded turbocharger assembly CTA includes a turbine section TS which " extracts energy from the exhaust gas products and a connected compressor section CS that provides a source of pressurized inlet air for engine * combustion and com- pressed air for a pneumatically-controlled actuator system, as explained more fully below. The cams Cl and C2, as shown in Fig. 4, are of the axial face type and have a generally circular base 30, a cylindrical axially-extending portion 32 that terminates with a curvilinear trackway 34, and a radially outward- extending rim or lip 36. Both the trackway 34 and the rim 36 engage cam follower mechanisms of the various pistons PI and P2 as described more fully below. The circular base 30 of each cam Cl and C2 engages a thrust bearing TB interposed between the cam and the supporting back plates BP. The thrust bearing TB can take various forms including the tapered roller bearing arrangement shown in Fig. 4 as well as other thrust bearing config- urations including pad-type plain bearings. The cams Cl and C2 are designed to rotate about the longitudinal axis LA to control piston motion while supported for such rotation by the back plates BP. As shown by the exemplary back plate BP on the left in Fig. 4, the back plates BP are of a generally annular construction with the outside diameter surface portion including an external machined surface 58 (seals not shown) that is received or telescoped within a mating inside diameter pilot surface formed in the supporting left-side engine block section EB for relative sliding motion therebetween along the longitudinal axis LA. The central portion of the back plate BP is open to define an air inlet Al, as shown, and includes an ax¬ ially-extending back plate sleeve 50 that carries inter- nal threads at 52, such as square or Acme-type threads, for engaging complementary external threads on an inter¬ nal support cylinder 54 of the left-side engine block EB. The inward end of the back plate sleeve 50 carries a circular retainer plate RP that retains the cam Cl in place between it and the above-described thrust bearing TB. A peripheral portion of the back plate BP carries a gear tooth sector 56. A gear motor GM is mounted on the left-side block half EB and carries a pinion gear PG3 on its output shaft that engages the gear tooth sector 56 to effect variable compression. As can be appreciated from the foregoing, operation of the gear motor GM to rotate its pinion gear PG3 to,, in turn, cause limited angular rotation of the back plate BP about the longi¬ tudinal axis LA causes the back plate BP and its cam Cl to advance or retract along the longitudinal axis LA to vary the operational range of the pistons Pi to effect variable compression control. The inward axial motion of the adjustable back plate BP is limited by a shoulder formed on the support cylinder 54 while outward axial motion is limited by an annular ring located on the outward end of the support cylinder 54. The variable compression- adjustment can be performed during engine operation to optimize engine performance for different types and grades of fuels, engine speeds, loads, and environmental conditions. Each of the pistons Pi and P2 is connected to re¬ spective cams Cl or C2 by an associated cam follower mechanism, several embodiments of which are shown in Figs. 8A, 8B, 8C, 8D, and 8E, as described in more de- tail below. The illustrated cam follower mechanism of Fig. 4 for connecting the piston P2, for example, to the cam C2 includes a rolling element, such as a spheroid or ball B2, interposed between the ball receiving trackway 34 of the cam Cl and a ball-receiving thrust pad TP retained in a pad receiving cavity in the base of the piston P2. The ball B2 serves to transmit the operating forces between the cam C2 and the piston P2 while mini¬ mizing friction therebetween. The cam follower mech¬ anism described is, of course, also utilized to connect the pistons PI to their cam Cl. *§ REA OMPI _ The piston P2 includes an axially extending skirt 62 through which a threaded retainer screw 64 extends. The screw 64 passes through a guide block 66 that rides within a longitudinally extending guide slot 68 formed in the cylinder skirt (only one side of the slot being shown in Fig. 4), and is adjustably retained in place by a locknut. The inward, conically tapered end of the retainer screw 64 engages and slides relative to the under surface of the outwardly extending rim 36. Access to the retainer screws 64 and their locknuts of each cam follower to effect adjustments may be had via removable access plates AP and access openings form¬ ed in the left- and right-side block sections. As can be appreciated, the various cam followers constrain their respective pistons PI or P2 to follow the cam profiles. The valve cartridge VC assembly is located between the two cylinder halves CYLl and CYL2 and between the faces or crowns of the two pistons Pi and P2. As shown in Fig. 4 and in the enlarged detail of Fig. 6, the valve cartridge VC includes a first stationary valving disk VDl, a second stationary valving disk VD2 spaced from the first stationary valving disk VDl, and a rotatably mounted in¬ termediate disk ID positioned between the two stationary valving disks VDl and VD2. Both of the stationary valving disks VDl and VD2 can be formed, if desired, integral with the end portions of their cylinder halves CYLl and CYL2. Each of the stationary valving disks includes a plurality of equi-spaced sector-shaped through openings 82 that are in registration with one another. The rotatable inter¬ mediate disk ID includes through openings 84 of equal number and general shape as those of the stationary val¬ ving disks VDl and VD2 so that the intermediate disk ID can be moved between an open position in which the open- ings or ports 82 and 84 are in registration to permit communication across the valve cartridge VC and a closed position in which the openings 82 and 84 are out of regis¬ tration with one another. Both stationary valving disks VDl and VD2 include raised sealing rims or pads 86 that bear and seal against the rotatable intermediate disk ID. Any gas leakage across the sealing pads 86 is bypassed into bypass passageways 88 that communicate with vent opening 90 that extends through the intermediate disk ID to permit gas flow thereacross and a passageway 92 that leads to the exhaust gas chamber described more fully below. The intermediate disk ID is constrained for rotation in a combined mounting ring and inlet valve sleeve VS with mounting surfaces defined as shown. A peripheral surface portion of the rotatable intermediate disk ID carries indentations 100 that are designed to engage the teeth of a gear rack GRl which, in turn, is connected to a pneu¬ matic control actuator CAl, described more fully below, which effects opening and closing of the divider valve in timed relationship with the motion of the other engine parts. The inlet ports IP are formed through the wall por¬ tion of the cylinder half CYLl generally adjacent to the valve disk VDl. The inlet valve sleeve VS is mounted for rotation on a peripheral surface portion of the cylinder CYLl and includes inlet ports of like number, size, and shape to the inlet ports IP formed through the cylinder wall. When the inlet valve sleeve VS is rotated so that the inlet ports are in registration, the interior of the cylinder CYLl is opened to a pre-inlet mixing chamber MC of the air inlet manifold IM. The lower peripheral por¬ tion of the inlet valve sleeve VS includes indentations of like size and shape to the indentations 100. Another gear rack GR2 engages the inlet valve sleeve and is connected. to a pneumatic control actuator CA2 of the control system that effects alternate opening and closing of the inlet valve sleeve VS as described more fully below. The cylinder portion CYLl includes a circumfer¬ ential flange 110, one side of which abutts shoulder 112 and the other side of which retains the inlet valve sleeve VS. In the assembled form, as shown in Fig. 4, the cylinder CYLl is secured in place in a cylinder-receiv¬ ing bore of the left-side block, the cylinder CYL2 is adjustably mounted in place in a similar cylinder-re¬ ceiving bore in the right-side block, and the various parts of the divider valve DV are located between the two inwardly facing ends of the cylinders CYLl and CYL2. The cylinder CYL2 is adjustable to control the clearance and fit between the stationary and intermediate valving disks. Adjustment shims 12 are retained in place by threaded fasteners 11 that pass through an apertured flange of the cylinder CYL2 and engage a bulkhead on the right-side block. The compounded turbocharger assembly CTA occupies the central or core portion of the engine and includes a turbine section TS and a compressor section CS with compressor units CU1 and CU2 for providing pressurized inlet air and a compressed air supply where, as shown in the Fig. 4 embodiment, pneumatic valve controllers are used. The turbine section TS includes a scavenging fan SF and turbine units TUl and TU2 for extracting useful energy from the expanded combustion products. The compressor section CS includes a first com- pressor unit CUl secured to the rotatable shaft 19. The compressor unit CUl induces inlet air through the cen¬ tral air inlet opening Al in the back plate BP and pro¬ vides an initial pressurization thereof. A circumfer¬ ential shutter 15, controlled by a throttle actuator TA, is circumferentially mounted in the chamber defined between the compressor units CUl and CU2 and is rotat¬ able to control inlet air flow through circumferential ports 13 into internal engine air flow passageways and the inlet air manifolds (not specifically shown). A second compressor unit CU2 provides compressed air with this compressed air being provided through an outlet opening 16 to the engine pneumatic system supply, des¬ cribed more fully below. The turbine section TS includes a scavenging fan SF and first and second turbine units TUl and TU2 that are secured to the rotatably mounted shaft 19 for rotation therewith. Exhaust gases enter the exhaust gas chamber from various peripheral exhaust ports EP and pass through the scavening fan SF and turbine units TUl and TU2 where usable energy is extracted and the exhaust thereafter passing through the exhaust pipe EXH. One or more water mist injectors WMI are positioned between the turbine units TUl and TU2 and for the pur¬ pose of. injecting a water mist into the exhaust gas flow. The water mist is converted to steam and thereby (a) removes thermal energy from and cools the exhaust gas flow and (b) increases the mass flow through the turbine unit TU2 to enhance energy extraction from the exhaust gas flow. The compounded turbocharger assembly CTA is con¬ nected to the engine output shaft OS, as shown schem¬ atically in Fig. 4, by an endless belt 14, such as a chain or cogged belt, that is entrained about cooper¬ ating sprockets or pulleys including a sprocket SI se- cured to the shaft 19, an idler sprocket S2, a sprocket S3, mounted on the input shaft of an engine oil pump OP, and a sprocket S4 mounted on the engine output shaft OS. At low engine speeds, energy may be delivered from the output shaft OS to the shaft 19 to drive the inlet air compressor units CUl and CU2, while at higher engine power outputs, sufficient energy may be available from OMPI the exhaust gas flow so that sufficient power is extrac¬ ted from the exhaust gas flow to drive the compressor units CUl and CU2, with power in excess of that needed to drive the compressor section CS directed to the en- gine output shaft OS. Turbine lag or delay is elimin¬ ated because of the direct mechanical connection to the engine output shaft OS. Fuel and air are introduced into the various cyl¬ inders by an external circumferential manifold assembly MA formed as a radial extension of the left-side and right-side block sections. The manifold assembly MA in¬ cludes spaced axial openings each of which receives a fuel injector FI, such as a fuel injector of the elec¬ tromechanical plunger-type. Each fuel injector FI is connected to a source of pressurized fuel and selec¬ tively injects a quantity of fuel across a manifold chamber through a check valve CVl such as the exemplary flap-type check valve shown (the function of which is discussed below) into the preinlet port mixing chamber MC that leads into the cylinder CYLl when the sleeve inlet valve IV is open. Additionally, a liquid injector LI is mounted in a radially aligned bore to selectively inject a combustion enhancing liquid, such as water or water in solution with other substances, including meth- anol, octane enhancers where gasoline type fuels are used and cetane enhancers where diesel type fuels are used. The liquid injectors LI are each aligned so as to inject the combustion enhancing fluid into the preinlet port mixing chamber MC to intermix with the fuel charge injected by the above-described fuel injectors FI. The operation of the fuel and combustion enhancing liquid injectors, FI and LI, is typically timed. so that the desired fuel and combustion enhancing fluid will pass into the cylinder CYLl, during that time the inlet valve IV ports are open. The inlet air supply for the entire engine is main¬ tained by the above-described first compressor unit CUl with the pressurized inlet air being provided through the throttle actuator TA controlled rotary shutter valve 15. In addition to the air and fuel inlet structure described above, the external circumferential manifold assembly MA also includes an inlet air/air exhaust as¬ sembly that is utilized during compression braking as discussed more fully below. The inlet air/air exhaust assembly includes the lower portion of the mixing cham¬ ber MC and a passageway that leads from the mixing chamber MC through a check valve CV2, such as the exem¬ plary flap-type check valve shown, into an expansion chamber EC to an air muffler AM that includes a sonic wave diffusing material constrained between two aper- tured-retaining plates. Overall thermal cooling for the engine shown in Fig. 4 is achieved by flowing coolant, such as water or water with a mixture of an antifreeze liquid, through various passageways. Coolant flow, temperature, and heat removal can be controlled through conventional means (not specifically shown in Fig. 4). The operation of the various inlet air and divider valves shown in Fig. 4 is achieved through pneumatic actuation. The air supply for the pneumatic actuator system is provided from the pressurized inlet air from compressor unit CUl with this inlet air compressed by the second compressor unit CU2. This compressed air is provided through a passageway 16 to a check valve CV3 to pneumatic control system air supply line 17. A pressure regulator PR connected to air supply line 17 returns excess air to the engine inlet in the region of the first compressor unit CUl for maintaining engine air supply line pressure and returning excess pressurized OMPI air to the inlet in the interest of energy conservation. The main compressed air-supply line 17 is connected to an air receiver AR and to an engine air-starter AS through a normally closed, solinoid operated, spring return, two-way starter valve SV. The air starter AS is of conventional design and includes an air motor driving a pinion gear PG4 that engages or is engageable with the ring gear RG2 of the cam C2. A branch air line 18 is provided from the main compressed air supply line 17 to the pneumatic control actuators CAl and CA2 for each valve cartridge VC. As shown in the enlargement of Fig. 6, each pneumatic con¬ trol actuator is of the single acting type and has a piston mounted for reciprocation within a cylinder with a spring resiliently urging the piston to its retracted position. An actuator rod extends from the piston through the cylinder end plate and is secured to the above-described gear racks which, in turn, engage the oscillating inlet valve sleeve VS or the oscillating intermediate disk ID of the divider valve DV as shown in Figs. 4 and 6. The exhaust from each pneumatic actuator retains substantial energy in terms of its pressure and flow velocity. The air exhaust from each of the pneu¬ matic actuators is passed through passageways to a reg- ion adjacent to the inlet side of the compressor unit CUl where the energy is recovered by directing the air into the vicinity of the blades of the compressor unit. Additionally, the compressor unit CUl serves to assist in the removal of the exhausted air from each pneumatic actuator because of the below-atmospheric pressure at the inlet of the compressor unit CUl which provides, for faster evacuation of the air from the pneumatic actu¬ ators to permit more rapid actuator cycling. Control actuator valves CAVl and CAV2, in the form of solinoid operated, spring return, three-way pneumatic valves as described below, are utilized to control the actuators CAl and CA2. The above-described pneumatic control system is operated in response to electrical signals provided from an electronic control system which monitors engine oper¬ ating parameters and provides the appropriate timed control signals to the various fuel and combustion en¬ hancing liquid injectors, FI and LI, as described above and the pneumatic control actuators CAl and CA2 to pro¬ vide variable valve timing throughout a wide range of engine speeds, loads, and operating conditions. Various on-engine sensors are provided throughout the engine and other various off-engine sensors are provided to obtain operating parameters. The on-engine and off-engine sensors suitable for effecting engine operation are listed below in Table I. TABLE I ON-ENGINE OFF-ENGINE INPUT INPUT SENSORS SENSORS & SIGNALS Fuel Pressure Brake Transducer Fuel Temperature Acceleration Transducer A/F Ratio Throttle Position Trans¬ ducer Cam (rotary) position Inclination (downhill) sensor Detonation (knock) Atmosphere pressure Coolant Temperature Fuel selection (CYL, CYLl, CYL2) Exhaust Temperature Ignition (ON/OFF) Exhaust Pressure Compression Ratio Adjust* ments Exhaust Sensors Constituents (C0χ, N0χ, O2) Inlet Air Pressure Inlet Air Temperature Pneumatic System Air Pressure Oil Pressure - CFREAl OMPI Oil Temperature Engine RPM The output of the various on-engine and off-engine sensors are supplied through an input port to the cen- tral control system which preferably takes the form of a microcomputer, for example, a single board computer containing a processor with at least one arithmetic and logic unit (ALU) and a plurality of associated working - registers, a read-only memory (ROM) containing engine operating programs, a random access memory (RAM) for temporary data storage and manipulation, and analog/- digital converters (A/D) for converting analog output of the analog type transducers and sensors to a digital value compatible with the single board computer. The central control system processes the output of the various transducers and sensors in accordance with a previously stored engine operating program and provides output control signals, as listed below in TABLE II, to effect engine operation. TABLE II Valve Control Actuators Compression adjuster including variable timing Variable Compression Divider Valve Control Braking Actuator Air Supply Controls Fuel Delivery (injectors, etc.) Start Sequence Inlet Manifold Liquid Cooling System Control Injector Water Injectors In operation, the cams Cl and C2 rotate causing the various pistons PI and P2 to reciprocate in their re¬ spective cylinders CYLl and CYL2. Output power is dir- ected by the rotating cams Cl and C2 to the output shaft OS as described above. The compounded turbocharger assembly CTA, at low speeds, obtains the power to pres¬ surize the engine inlet air via its direct mechanical connection-to the output shaft OS. Inlet air is induc¬ ted through the engine air inlet Al with an initial. pressurization provided by the first compressor unit CUl and compressed air provided by the second compressor unit CU2 to provide control system air. The pressurized inlet air is provided through the ports 13 of the rotat¬ able throttling sleeve 15 to pass through various passages in the external manifold assembly MA to the cylinders CYLl. Exhaust gas flow is provided through the exhaust ports EP through various exhaust port passageways to the turbine section TS where the exhaust gas flow is passed through scavenging fan SF and the turbine units TUl and TU2 to extract the useable power therefrom. The water mist injectors WMI inject a controlled quantity of water into the- high temperature exhaust gas flow between the first and second turbine units TUl and TU2 to remove heat energy from the exhaust gas by flashing to steam and thereby extracting energy removed from the exhaust gas flow. The control system also provides control signals to the gear motor GM that controls the compression ratio of the engine to vary the compression ratio to optimize efficiency and match engine operation to the operating conditions and the load. This optimization and matching can be accomplished, for example, by increasing the compression ratio until such time that the detonation sensor just begins to detect the presence of pre-igni- tion during the compression stroke of the pistons PI. As can be appreciated, the use of the detonation sensor indirectly takes into account all those parameters that affect efficient engineOperation including the type and grade of fuel being consumed, the engine load, and vari- ous environmental conditions. As can be appreciated by those skilled in the art, the variable compression ratio feature permits the engine to be optimally operated under all anticipated operating conditions including, as an example, transportation applications in which a vehi- cle can experience uphill, downhill, and level travel. For an engine using a medium octane gasoline fuel in a transportation application during level travel, the control system can automatically increase the compres¬ sion ratio to a point just before detonation, for exam- pie, 12:1, to maximize engine output; upon the onset of uphill driving, the compression ratio can be lowered, for example, to 8:1, to minimize detonation; and, last¬ ly, upon the onset of downhill driving, the compression ratio can then be increased or decreased without the presence of fuel during the variable compression braking mode, described below, to vary the now-desired engine braking contribution. The engine described above utilizes a pneumatic actuator system to effect engine control. As can be appreciated, other types of actuators and actuator sys¬ tems are suitable including hydraulic, electro-mechan¬ ical, and vacuum systems. For example, hydraulic sys¬ tems are preferred in larger, slower engines and pneu¬ matic or electro-mechanical systems are preferred in smaller, higher speed engines. The engine described above includes a variable compression braking feature in which the engine can Be configured as a variable energy absorbing compressor. The variable compression braking feature is useful in transportation applications, particularly in regulating vehicle speed during downhill travel of a vehicle, but also to assist in reducing vehicle speed when desired during level or uphill travel. Generally, the variable compression braking is accomplished by charging the cylinders with various quantities of air, compressing the air charge by a selected compressios ratio and/or releasing the so compressed air before full compression in the engine. Speed can also be regulated by varying the amount of air in the cylinders during the compression braking mode by throttling the inlet air. In the compression braking mode, the divider valve DV is desirably left in an open position while the inlet valve IV for each cylinder is periodically opened when the pistons Pi and P2 are at or near their respective top dead center positions. Each of the pistons Pi and P2 then retracts towards bottom dead center under con¬ trol of their respective cams Cl and C2. During re¬ traction, pressurized inlet air passes through the low pressure check valve CVl adjacent to the fuel injector Fi until the pistons PI and P2 are at or near their bottom dead center positions thereby charging the cyl¬ inder with a fresh air charge. ' Fuel or combustion en¬ hancing fluid is not injected into the inlet air during the compression braking mode. When the pistons Pi and P2 reach their respective bottom dead center positions, the inlet valve IV is closed. On the subsequent com¬ pression stroke, the pistons PI and P2 compress the fresh air charge as they move toward and to their top dead center positions. Upon attaining the top dead center position when the air charge is fully compressed, the inlet valve IV is once again opened to allow the now-compressed air charge to vent through the passageway that starts at the lower end of the mixing chamber MC into the preinlet port mixing chamber and leads to check valve CV2. The low pressure check valve CVl adjacent to the fuel injector FI snaps closed while the second low pressure check valve CV2 then opens to allow the now- compressed air charge into the expansion chamber EC and on to the air muffler AM. In this way the cylinders are converted into a compressor in which substantial amounts of energy can be absorbed in the compression braking mode during, for example, downhill coasting of a large truck. The air compressed during the compression brak¬ ing mode can be recovered, for example, by charging an air receiver for use in the pneumatic systems of those trucks and other vehicles that utilize pneumatic brakes, starter, and accessory systems. In some applications, pilot-operated check valves may be preferred. A practical, preferred embodiment on an engine, referred to by the reference character E, in accordance with the present invention is shown in Fig. 5 and dif¬ fers from the engine 10 of Fig. 4 in that both the inlet valves and the divider valves of the various cylinders CYL are mechanically operated rather than pneumatically operated and that the engine is air cooled rather than liquid cooled. As shown in Fig. 5, the engine includes a plurality of cylinders CYLl...n (only two of which are shown in cross-section in Fig. 5) with each cylinder defined by a first cylinder portion CYLl and a second in-line cyl¬ inder portion CYL2 and with each divider valve and inlet valve located near the mid part of the combined cylinder as in the case of the Fig. 4 engine. The cylinder por- tions CYLl and CYL2 are both located in respective left and right block sections EBl and EB2 that are secured together by appropriate threaded fasteners along the medial line to constrain the divider and inlet valves between the two cylinder portions of each cylinder. The cylinders CYLl and CYL2 and the divider valve are maintained in their assembled form in a manner sim¬ ilar to that of Fig. 4. A compression spring resistancy urges the cylinder CYL2 toward the cylinder CYLl to apply a sealing force for the various mating surfaces and to compensate for wear. The back plate BP2 is secured by appropriate threaded fasteners and pilot pins to the end of the engine block while the back plate BP1 is adjustable to provide a variable compression feature. The back plate includes external threads, such as Acme or square threads, that engage internal complementary threads formed on the inside diameter surface the block EBl. A portion of the peripheral surface of the back plate BP1 includes an external gear sector which is designed to engage pinion gear PG3. A reversible gear motor drives the pinion gear PG3 to cause the back plate BP1 to rotate about the longitudinal axis and advance axially into or retract from the interior of engine block EBl to vary the compression ratio. Respective axial faced cams Cl and C2, of construction similar to those discussed above relative to the Fig. 4 embodiment, are mounted on the back plates BP1 and BP2. As shown with respect to the cam Cl, the cam includes a base that'is received within an annular thrust bearing channel BC for relative rotation, an axially-extending portion that terminates with a curvilinear trackway and a radially outward- extending retaining lip similar to that described above in relationship to Fig. 4. The cams are each retained within their annular thrust bearing channel BC by a retaining plate as shown. The external periphery of each of the cams Cl and C2 includes a radially-extending ring gear which engage the pinion gears secured to the output shaft OS. The cam follower mechanisms which secure each of the pistons Pi and P2 to their respective cams Cl and C2 are of construction similar to that disclosed above in relationship to the embodiment of Fig. 4. Each piston skirt includes a guideway that receives a guide block assembly GBA. A retainer screw passes through the guide -ξT EA OMPI block assembly and is secured in place by a locknut in a manner similar to that described above in Fig. 4. The engine E includes a compressor section in a manner similar to that of the Fig. 4 embodiment in- eluding, in the left engine block EBl, a first stage compressor for providing initial pressure to the engine air cooling system. Air is provided through the inlet air filter. This initially pressurized inlet air is provided through peripheral apertures in a rotatably mounted shutter valve. Cooling mist injectors CMI are provided on a bolt circle somewhat larger than the rotary shutter valve to selectively inject a coolant mist into the air stream of the cylinder cooling system. The mist flow rate is controlled so that the mist-laden cooling air, as it flows past the walls of the first cylinder CYLl flashes to steam to thereby absorb heat of from the cylinder- CYLl walls and thereby effectively limit the temperature in the cylinder CYLl to a desirably low limit. As the now steam-laden air passes the cylinder wall CYL2, addi¬ tional heat energy is removed; however, the quantity of additional heat energy removed from the cylinder CYL2 is less than that absorbed from the cylinder CYLl so that the cylinder CYL2 operates at a desirably higher temper- ature. The temperature of the cylinder CYLl is kept desirably low and at or below a selected temperature to limit pre-ignition. The cylinder CYL2 temperature is maintained as high as reasonably practicable since the cylinder CYL2 contains the ignition air charge. The maintenance of a high temperature in the cylinder CYL2 increases the overall thermal efficiency of engine E. The first stage compressor, as described above, assists, along with -a scavenging fan described below, in main¬ taining the desired steam flow in the cylinder jackets. Useable energy is recovered from the steam flow in the engine cooling jackets by a cooling energy recovery system CER • contained within a CER system housing located in the right-side engine block EB2, to the right of the valve control cam galley. The steam-laden air flow from the cooling jackets is directed into the CER system housing as shown in Fig. 5. The cooling energy recovery system CER includes, firstly, a scavenging fan mounted on the turbine main shaft TMS and, secondly, a steam turbine mounted adjacent to the scavenging fan. The scavenging fan serves to maintain the steam flow through the cooling jackets while the turbine extracts useable energy from the steam-laden air flow. If desired, the steam-laden air can be passed through a condenser to effect coolant recovery, for example, where a closed coolant cycle is desired. " In addition to the recovery of energy in the steam flow in the cooling jackets, useable energy is also recovered from the very high temperature and remaining pressure of the exhaust gas flow by the exhaust energy recovery system contained within an EER system housing located in the right-side engine block EB2 to the right of the cooling energy recovery system CER. The exhaust energy recovery system EER includes, firstly, a scavenging fan mounted on the turbine main shaft, sec¬ ondly, a gas turbine mounted on the turbine main shaft TMS downstream of the scavenging fan, and, thirdly, a steam turbine mounted on the turbine main shaft TMS downstream of the gas turbine. The scavenging fan assists in scavenging exhaust gas from the various cylinders during their respective scavenging phases by inducing a partial vacuum in each cylinder and the ex¬ haust gas passageways to assist the exhaust gas flow and also eliminate back pressure. The exhaust gas flow then drives the gas turbine in a conventional manner to ex¬ tract useable mechanical energy from the exhaust gas flow. A water mist injector WMI is located between the gas turbine and the steam turbine and serves to intro¬ duce a water mist spray into the high temperature ex¬ haust gases between the gas turbine and the steam tur¬ bine. The water mist removes heat energy from the ex¬ haust gas by flashing to steam thereby creating a pres- sure increase. The steam-laden exhaust gas is then passed through the steam turbine which then extracts useable energy therefrom. The exhaust gas products, downstream of the steam turbine, are vented to the atmosphere through the exhaust pipe EXH. If desired, and depending upon the particular engine application, the moisture-laden exhaust gas can be passed through a condenser to effect water recovery. Condensation will also effectively remove particulate matter entrained in the exhaust gas and remove those gaseous exhaust products that dissolve within the condensate. Valve operation is controlled by circular plate- type cams CIV and CDV located in a cam galley. The cams CIV and CDV are secured to a bearing supported sleeve BSS for rotation independently of the turbocharger main shaft TMS. The left end of the sleeve carries a gear that engages other gears (partially shown) of a gear train that is driven by a chain drive arrangement that is connected to the output shaft OS. The gear train drives the two cams CIV and CDV at a fixed rotational ratio and in timed relationship to the rotation of the cams Cl and C2. Each of the cams CIV and CDV is generally circular and include a groove or channel (partially shown) cut into one face thereof with each groove defining a closed circuit, the radius of which varies from the axis, of OMPI rotation to effect control of the inlet and divider valves. While the cam channels are not specifically shown, they include radial variations that cause opening and closing of the inlet valve and divider valve of each cylinder in timed relationship to one another and in timed relationship with the rotary position of the cams Cl and C2. An elongated cam follower is provided for each of the inlet valves and divider valves with each of the cam followers carried in a follower bushing and with one end of the cam follower carrying a pin or roller that fits within the channel of the appropriate valving cam. The opposite ends of each cam follower carry gear racks which engage, as shown in the detail of the upper valve of Figs. 7A and the lower valve in Fig. 7B, com- plementary indentations 100 in the peripheral surfaces of the rotary inlet valve sleeve and the oscillating disk portion of the divider valve in a manner similar to that of the engine 10 of Fig. 4. Accordingly, each of the valving cams CIV and CDV is connected to cam followers in equal number to the number of cylinders in the engine E. In operation, as the bearing supported sleeve BSS carrying the control cams CIV and CDV rotates, the fol¬ lowers reciprocate in their respective bushings to con¬ trol inlet valve and divider valve operation. The valve cartridge structure of the Fig. 5 engine E embodiment is similar in function to those of the Fig. 4 engine embodiment although of somewhat different mech¬ anical configuration as shown in Figs. 7A and 7B. Each divider valve includes first and second spaced sta- tionary disks VDl and VD2, each having a plurality of sector-shaped ports 82 or openings equi-spaced about the cylinder axis. An oscillatable intermediate disk ID is mounted between the two stationary disks VDl and VD2 and constrained therebetween. The oscillatable intermediate disk ID includes a number of ports 84 equal in size, number, and shape to those of the stationary valving disks VDl and VD2. The stationary valving disks VDl and VD2 each include rims or pads 86 about their through ports to bear and seal against the intermediate disk ID and prevent any leakage across the valve to prevent unintentional preignition of the fuel/air charge by the ignition temperature air in the adjacent cylinder. The inlet valve consists of a rotatable valve sleeve having a plurality of peripheral openings, intake ports IP that are designed to be in alternate registration with and out of registration with inlet ports formed through the cylinder wall of the CYLl. As shown in Figs. 7A and 7B, the inlet valve sleeve and the intermediate disk ID carry indentations that engage the gear racks of the above-described valve cam followers. The piston cam follower embodiments shown in Figs. 4 and 5 are representative of various follower arrange¬ ments. Other piston cam follower structural organi¬ zations suitable for use in the engine embodiments des- cribed include those shown in Figs. 8A-8E. The cam follower 500 shown in cross-section in Fig. 8A utilizes a tapered rolling element 502 that includes a sleeve bearing 504 journaled for relative rotation about a connecting pin 506. The connecting pin 506 is threaded at its upper end and is in threaded engagement with an internally threaded counterbore formed in the skirt portion of the piston PlO as shown. A locknut 508 maintains the adjusted position of the connecting pin 506 once the desired assembly and adjustment is com- pleted. The remote end 510 of the connecting pin 506 is tapered as shown. The cam CIO shown in Fig. 8A- is a structural variation of the cams shown in Figs. 4 and 5 and includes a trackway that is generally linear so as to engage the linear tapered surface of the rolling element 502 with the retainer lip 514 offset from the- to to to to H H tn o UI o Lπ o Ul clearance bore in a guide block through a threaded bore in the piston skirt. A locknut 548 maintains the re¬ tainer screw 546 in its adjusted position. The remote end of the retainer screw 546 is tapered as shown at " 550 and engages the underside of the retainer lip of cam C40. The reciprocating motion of the piston P40 is controlled, as in the case of the embodiments described above, by the motion of the guide block in the guide slot. The cam follower 560 of the Fig. 8E embodiment includes a spheroidal ball 562 which is located between and engages spherically-shaped mating surfaces on the " trackway of the cam C60 and a bearing pad 564 located in the base of the piston P60. The bearing pad 564 is mounted in a pad receiving cavity with clearance 574 top and bottom to allow self-alignment. The retainer as¬ sembly is defined by a spheroidal rolling element which engages a spheroidal complementary surface on the under¬ side of the cam C60 and is maintained in place by an adjustment locknut 568 and screw 566 which bears against a retaining ball pad portion of a guide block 570 which rides in the guide slot. A mounting screw 572 mounts the assembly to the skirt of the piston P60. As can be appreciated by those skilled in the art, the operating cycle(s) illustrated in Figs. 1-3 and their above-described variations are not limited to the engine configurations described in Figs. 4 and 5 but are well suited to crankshaft/connecting rod type engines. Engines with two crankshafts also can be adapted readily to the cycles of Figs. 1-3 by connecting the crankshafts with 1:1 or 1:2 drive ratios. For example, a twin crankshaft, opposed piston, two-cycle engine can imple¬ ment the operating cycle of Fig. 1 (2/2 stroke) by con¬ necting the crankshafts with a 1:1 ratio drive and also adapted to the operating cycles of the Fig. 2 (4/2 stroke) cycle and of the Fig. 3 (4V/2 stroke) cycle by connecting the crankshafts with a 1:2 ratio drive. In addition to this twin crankshaft opposed piston con¬ figuration, the H-block opposed cylinder arrangement can be operated to implement the various cycles of Figs. 1-3 with the crankshafts connected with a 1:1 or a 1:2 drive ratio and with the combustion chambers forming a "U" or similar shape. Single crankshaft engine configurations that can be adapted to the cycle of Fig. 1 include the parallel cylinder, double-row radial, opposed cylinder, V-block, and in-line configurations, among others. Modification " of the crankshaft in these configurations is required to allow piston pairs to function as the air/fuel and igni- tion air working pistons with the pistons attaining their top dead center positions substantially simul¬ taneously with the two cylinders and the combustion chambers forming a "U" or similar shape and with a divider valve.between the two cylinders to define the air/fuel chamber and the ignition air chamber. The cycles of Figs. 2 and 3 can be adapted to a split in-line crankshaft arrangement by separating the crankshaft between adjacent cylinders and reconnecting the crankshafts with a 1:2 ratio drive. in addition, the operating cycle can be adapted to a rotary (Wankel) engine configuration in which twin rotors function as the equivalent of the working pistons and the divider valve. In such an arrangement, twin rotors are mounted on a common shaft with each rotor mounted in their respective chamber with a passageway communicating between the twin chambers. The rotors are effective to compress respective air/fuel and ignition air charges to the necessary pressures with the rotors uncovering the passageway to admit the ignition temp- OMPI erature air into the air/fuel mixture to effect com¬ bustion with the subsequently expanding combustion prod¬ ucts operating against both rotors. By providing a con¬ trollable divider valve in the passageway and by varying the effective chamber sizes in a manner similar to that described above, it is possible to achieve results simi¬ lar to those of the piston/cylinder engines. After considerable research as to ' the attributes of various engine configurations, the axial-piston configu- ration was selected as the preferred embodiment for its recognized advantages. As will be apparent to those skilled in the art, various modifications and variations of the described preferred embodiment can be made without departing from the spirit and scope of the invention as presented in the appended claims and their legal equivalent. OMPI

Claims: WHAT IS CLAIMED IS: 1. An internal combustion engine, comprising: block means having at least one cylinder formed therein; first and second pistons located within said cylinder for controlled reciprocating motion to advance toward one another and retract away from one another; valve means located within said cylinder between said first and second pistons and operable between an open position permitting communication between a first chamber defined between the first piston, said cylinder, and said valve means and a second chamber defined between said second piston, said cylinder, and said valve means; control means connected to said first and second pistons for controlling the reciprocating motion thereof by causing each of said pistons to advance to a predetermined first position adjacent to said valve means and for permitting each of said pistons to retract away th'eir respective first positions to a respective second position; introducing means for introducing air into said second chamber when said second piston is in a position other than said first-mentioned predetermined position and for intro¬ ducing an air and fuel mixture of predetermined ratio into said first chamber, said control means operative to cause said first piston to compress said air and fuel in said first chamber to a predetermined pressure and to cause said second piston to compress said air in said second chamber to a predetermined pressure that is sufficient to raise the temperature of the so-compressed air to a temperature above the auto-ignition temperature of the compressed air-fuel mixture; actuator means connected to said valve means for causing said valve means, during the compression of said air in said second chamber and said air and fuel mixture in said first chamber to remain closed, and for opening said valve at a predetermined time to cause the so-compressed air from said second chamber to enter the compressed air and fuel mixture of said first chamber to cause ignition thereof, said pistons retracting in response to the formation of combustion pro¬ ducts from said so-ignited air-fuel mixture; and exhaust means for exhausting the combustion products of said so-ignited air-fuel mixture after substantial retraction of said first and second pistons. 2. An internal combustion engine, comprising: block means having at least a first and a second cylinder formed therein; a first piston located in said first cylinder for controlled reciprocating motion and a second piston located in said second cylinder for controlled reciprocating motion; * passageway means connected between said first and second cylinders to p ' ermit communication therebetween; valve means operable between an open position and a closed position between said first and second cylinders; control means connected to said first and second pis¬ tons for causing reciprocating motion thereof by causing each piston to advance in its respective cylinder to respec¬ tive advanced positions and to retract to respective re¬ tracted positions; introducing means for introducing air into said second cylinder and for introducing a predetermined ratio of air and fuel in said first cylinder, said control means causing said first piston to advance toward its advanced position to compress the air and fuel mixture therein to a predetermined ratio and causing said second piston to advance toward its advanced position to compress the air therein to a predeter¬ mined pressure sufficient to raise the temperature of the air therein to a temperature greater than the self-ignition temperature of the compressed air-fuel mixture in said first chamber; OMPI actuator means connected to said valve means for closing said valve means during the compression of the air in said second cylinder and the air and fuel in said first cylinder and for the opening thereof at a predetermined time to cause the so-compressed air from said second chamber to enter the compressed air and fuel mixture of said first chamber to cause ignition thereof; said pistons retracting in response to the formation of combustion products from said so-ignited air-fuel mixture; and exhaust means for exhausting the combustion products of the so-ignited air-fuel mixture after retraction of said pistons. 3. An internal combustion engine, comprising: block means having a plurality of cylinders formed therein and aligned parallel to and spaced from a central longitudinally extending axis; a piston pair, having first and second pistons, located in each of said cylinders for controlled reciprocating motion toward an advanced position and away therefrom to a retracted position; valve means located in each of said cylinders between said respective first and second pistons thereof and oper¬ able between an open position permitting communication between a first chamber defined between said first piston, said cylinder, and said valve means and a second chamber defined between said second piston, said cylinder, and said valve; control means connected to said first and second pistons for controlling the reciprocating motion thereof by periodically causing each of said pistons to each advance to a predetermined advanced position adjacent to said valve means and permitting each of said pistons to retract away from said first predetermined position to a respective second retracted position; ^TREAlT _OMPI " vT-Pδ " fe?N A TlO introducing means for introducing air into said second chamber when said second pistons are in a position other than said first predetermined position and for introducing an air and fuel mixture of predetermined ratio into said first chamber, said control means operative to cause said first piston to compress said fuel and air mixture in said first chamber to a first predetermined pressure and opera¬ tive to cause said second pistons to compress said air in said second cylinder to a predetermined pressure sufficient to raise the temperature thereof to a temperature above the auto-ignition temperature of the compressed air and fuel mixture; actuator means connected to said valve means for closing said valve means during the compression of the air in said second chamber and the air and fuel mixture in said first chamber and for opening said valve means at a pre¬ determined time to cause the compressed air from said second chamber to enter the compressed air-and fuel mixture of said first chamber to cause ignition thereof; said pistons re¬ tracting in response to the formation of combustion products from said so-ignited air-fuel mixture; and exhaust means for exhausting the combustion products of said ignited air and fuel mixture after retraction of said pistons. 4, The internal combustion engine claimed in claim 1, wherein: said block means includes a plurality of lineally aligned cylinders located parallel to and equispaced from a common axis. 5. The internal combustion engine claimed in claims 1 or 3, wherein: said first and second pistons reciprocate along a common axis about which said cylinder is defined. 6. The internal combustion engine claimed in claims 1, 2, or 3, wherein said valve means further comprises: at least first and second valvings disks each of which has openings therein with at least one of said disks movably mounted for rotation relative to the other to effect valve opening by placing said openings in registration with one another. 7. The internal combustion engine claimed in claims 1 or 3, wherein said control means comprises: first and second axial-faced cams mechanically coupled to one another, and first and second connecting means for connecting, respectively, said first and second pistons to said first and second axial-faced cams. 8. The internal combustion engine claimed in claim 7, wherein: . . each of said axial-faced cams comprises a circular track facing its respective connected piston and a second, oppositely facing track, and said first and second connecting means each comprises a first rolling element disposed between said first-mentioned track and a rolling element mounting means of said connected piston and a second rolling element connected between a second rolling element mounting means of said piston and said second, oppositely facing track. 9. The internal combustion engine claimed in claim 7, wherein: each of said axial-faced cams comprises a first track facing its respective connected piston and a second, oppositely facing track, and cam follower means comprising a first rolling element disposed between a rolling element mounting means of said connected piston and said first-mentioned track and a radially aligned guide pin connected to a guide pin mounting means of said connected piston engaging said second-mentioned oppositely facing track. 10. The internal combustion engine claimed in claim 7, wherein: said axial-faced cam comprises a first track facing its respective connected piston and a second, oppositely facing track, and cam following means comprising a first bearing having an outer race engaging said first-mentioned track and a second roller bearing having an outer race engaging said second-mentioned track. 11. The internal combustion engine claimed in claim 10, wherein the outer races of said first and second-mentioned bearings are coincident with a line passing through the center line of said engine. 12. The internal combustion engine claimed in claim 7, wherein said engine includes an output shaft connected between said first and said second axial-faced cams. 13. The internal combustion engine claimed in claim 12, wherein each of said first and second axial-faced cams in¬ cludes external gear teeth and said output shaft includes first and second connected pinions for engagement with said external teeth of said first and second axial-faced cams. 14. The internal combustion engine claimed in claim 1, wherein at least one of said first and second axial-faced cams is adjustable along the engine axis to vary the com¬ pression of at least one of said first and second compres¬ sion chamber. 15. The internal combustion engine claimed in claim 1, wherein said introducing means comprises: at least one intake port formed through the wall of said cylinder between said valve means and said first piston and means for opening and closing said at least one intake port. * • 16. The internal combustion engine claimed in claim 15, wherein said means for opening and closing said at least one intake port comprises an oscillatable sleeve valve. 17. The internal combustion engine claimed in claim 16, wherein said introducing means further comprises a controll¬ able fuel injector means mounted for injecting selected quantities of a hydrocarbon fuel through said intake port. 18. The internal combustion engine cl.aimed in claims 1, 2, or 3, further comprising: water injection means positioned for injecting a pre¬ determined quantity of water into the inlet air of said, engine. 19. The internal combustion engine claimed in claims 1, 2, or 3, wherein said actuator means comprises: pneumatically operated valve actuator for operating said valve means. 20. The internal combustion engine claimed in claims 1, 2, or 3, further comprising a turbine means to extract energy from the exhausted combustion products of said so-ignited air-fuel mixture, said turbine means connected to a compressor means for pressurizing inlet air introduced through said introduction means. OMPI 21. The internal combustion engine claimed in claim 20, further comprising water injection means for injecting a predetermined quantity of water into the exhausted combustion products to effect cooling thereof and expansion of the injected water and to extract additional power through said turbine means. 22. A method for extracting energy from a hydrocarbon fuel comprising the steps of: compressing a mixture of a combustion supporting gas and a hydrocarbon fuel to a predetermined pressure; compressing a quantity of the combustion supporting gas to a predetermined pressure sufficient to cause the tempera¬ ture of the so-compressed compression supporting gas to increase to a temperature greater than that of the ignition temperature of said so-compressed mixture of a combustion supporting gas and a hydrocarbon fuel; ' introducing the so-compressed combustion supporting gas into the so-compressed mixture of a compression supporting gas and hydrocarbon fuel to cause ignition thereof; and, introducing the combustion products of the so-ignited mixture of the combustion supporting gas and hydrocarbon fuel into the compressed combustion supporting gas to cause the addition of heat thereto; extracting work from the combustion products of said so-ignited mixture and the additionally heated combustion supporting gas. 23. An internal combustion engine comprising: first and second working pistons mounted for controlled reciprocation in first and second cylinder means to define a first compression chamber for compressing an air/fuel charge by a selected compression ratio and a second compression chamber for compressing an ignition air charge by a selected compression ratio sufficient to cause the temperature of the

Inventors: GERACE ANTHONY, GERACE, ANTHONY
Applicants: GERACE ANTHONY, GERACE, ANTHONY
Application year: 1984
Main classification: F02B1
IPC codes: F01B3, F02B1, F02B3, F02B25, F02B41, F02B47, F02B75, F02F7