Patent Publication Number: US-7905221-B2

Title: Internal combustion engine

Description:
RELATED APPLICATIONS 
     This application is a Continuation in Part of U.S. Nonprovisional Ser. No. 12/108,728, filed Apr. 24, 2008, now U.S. Pat No. 7,634,988 which claimed priority of U.S. Provisional Ser. No. 60/914,273, filed Apr. 26, 2007, and Ser. No. 60/926,708, filed Apr. 27, 2007. 
    
    
     BACKGROUND OF THE INVENTION 
     The present invention relates to an apparatus and method for obtaining mechanical energy directly from the expenditure of the chemical energy of fuel burned in a combustion chamber that is an integral part of the apparatus, and more particularly to an internal-combustion engine. 
     Internal-Combustion Engine is any type of machine that obtains mechanical energy directly from the expenditure of the chemical energy of fuel burned in a combustion chamber that is an integral part of the engine. 
     In 1873 Brayton, an American, developed an engine, which had the unique features of constant-pressure combustion and complete expansion. One cylinder was used to compress air or the combustible mixture. Another cylinder was used as a working cylinder and was large enough to obtain complete expansion to atmospheric pressure. The compressor discharged the mixture into a receiver, and the mixture flowed from the receiver to the engine, being ignited and burned at constant pressure as it entered the engine. An ignition flame was supported by a mixture by-pass, and a flame-suppression grid prevented the flame from flashing back into the mixture receiver. 
     The Brayton engine could not compete with the Otto-cycle engine because of high heat and mechanical-friction losses, and it was abandoned when the Otto-engine was introduced in the United States. Although the Brayton process was abandoned for the piston engine, it is used for gas-turbine engine process. 
     Four principal types of internal-combustion engines are in general use: the Otto-cycle engine, the Diesel engine, the rotary engine, and the gas turbine. 
     The Otto-cycle engine, named after its inventor, the German technician Nikolaus August Otto, was first built in 1876 and is the familiar gasoline engine used in automobiles and airplanes. 
     The Diesel engine, (U.S. Pat. No. 542,846, granted on Jul. 16, 1895) named after the French-born German engineer Rudolf Christian Karl Diesel, operates on a different principle and usually uses oil as a fuel. It is employed in electric-generating and marine-power plants, in trucks and buses, and in some automobiles. Both Otto-cycle and Diesel engines are manufactured in two-stroke and four-stroke cycle models. 
     The essential parts of Otto-cycle and Diesel engines are the same. The combustion chamber consists of a cylinder, usually fixed, that is closed at one end and in which a close-fitting piston slides. The in-and-out motion of the piston varies the volume of the chamber between the inner face of the piston and the closed end of the cylinder. The outer face of the piston is attached to a crankshaft by a connecting rod. The crankshaft transforms the reciprocating motion of the piston into rotary motion. In multi-cylindered engines the crankshaft has one offset portion, called a crankpin, for each connecting rod, so that the power from each cylinder is applied to the crankshaft at the appropriate point in its rotation. Crankshafts have heavy flywheels and counterweights, which by their inertia minimize irregularity in the motion of the shaft. An engine may have from 1 to as many as 28 cylinders. 
     The fuel supply system of an internal-combustion engine consists of a tank, a fuel-pump, and a device for vaporizing or atomizing the liquid fuel. In Otto-cycle engines this device is either a carburetor or, more recently, a fuel-injection system. In most engines with a carburetor, vaporized fuel is conveyed to the cylinders through a branched pipe called the intake manifold and, in many engines, a similar exhaust manifold is provided to carry off the gases produced by combustion. The fuel is admitted to each cylinder and the waste gases exhausted through mechanically operated poppet valves or sleeve valves. The valves are normally held closed by the pressure of springs and are opened at the proper time during the operating cycle by cams on a rotating camshaft that is geared to the crankshaft. By the 1980s more sophisticated fuel-injection systems, also used in Diesel engines, had largely replaced this traditional method of supplying the proper mix of air and fuel. In engines with fuel injection, a mechanically or electronically controlled monitoring system injects the appropriate amount of gas directly into the cylinder or inlet valve at the appropriate time. The gas vaporizes as it enters the cylinder. This system is more fuel-efficient than the carburetor and produces less pollution. 
     In all engines some means of igniting the fuel in the cylinder must be provided. For example, the ignition system of Otto-cycle engines described below consists of a source of low-voltage, direct current electricity that is connected to the primary of a transformer called an ignition coil. The current is interrupted many times a second by an automatic switch called the timer. The pulsations of the current in the primary induce a pulsating, high-voltage current in the secondary. The high-voltage current is led to each cylinder in turn by a rotary switch called the distributor. The actual ignition device is the spark plug, an insulated conductor set in the wall or top of each cylinder. At the inner end of the spark plug is a small gap between two wires. The high-voltage current arcs across this gap yielding the spark that ignites the fuel mixture in the cylinder. 
     Because of the heat of combustion, all engines must be equipped with some type of cooling system. Some aircraft and automobile engines, small stationary engines, and outboard motors for boats are cooled by air. In this system the outside surfaces of the cylinder are shaped in a series of radiating fins with a large area of metal to radiate heat from the cylinder. Other engines are water-cooled and have their cylinders enclosed in an external water jacket. In automobiles, water is circulated through the jacket by means of a water pump and cooled by passing through the finned coils of a radiator. Some automobile engines are also air-cooled, and in marine engines seawater is used for cooling. 
     Unlike steam engines and turbines, internal-combustion engines develop no torque when starting, and therefore provision must be made for turning the crankshaft so that the cycle of operation can begin. Automobile engines are normally started by means of an electric motor or starter that is geared to the crankshaft with a clutch that automatically disengages the motor after the engine has started. Small engines are sometimes started manually by turning the crankshaft with a crank or by pulling a rope wound several times around the flywheel. Methods of starting large engines include the inertia starter, which consists of a flywheel that is rotated by hand or by means of an electric motor until its kinetic energy is sufficient to turn the crankshaft, and the explosive starter, which employs the explosion of a blank cartridge to drive a turbine wheel that is coupled to the engine. The inertia and explosive starters are chiefly used to start airplane engines. 
     Otto-Cycle Engines 
     The ordinary Otto-cycle engine is a four-stroke engine; that is, in a complete power cycle, its pistons make four strokes, two toward the head (closed head) of the cylinder and two away from the head. During the first stroke of the cycle, the piston moves away from the cylinder head while simultaneously the intake valve is opened. The motion of the piston during this stroke sucks a quantity of a fuel and air mixture into the combustion chamber. During the next stroke, the piston moves toward the cylinder head and compresses the fuel mixture in the combustion chamber. At the moment when the piston reaches the end of this stroke and the volume of the combustion chamber is at a minimum, the fuel mixture is ignited by the spark plug and burns, expanding and exerting a pressure on the piston, which is then driven away from the cylinder head in the third stroke. During the final stroke, the exhaust valve is opened and the piston moves toward the cylinder head, driving the exhaust gases out of the combustion chamber and leaving the cylinder ready to repeat the cycle. 
     The efficiency of a modern Otto-cycle engine is limited by a number of factors, including losses by cooling and by friction. In general, the efficiency of such engines is determined by the compression ratio of the engine. The compression ratio (the ratio between the maximum and minimum volumes of the combustion chamber) is usually about 8 to 1 or 10 to 1 in most modern Otto-cycle engines. Higher compression ratios, up to about 15 to 1, with a resulting increase of efficiency, are possible with the use of high-octane antiknock fuels. The efficiencies of good modern Otto-cycle engines range between 25 and 30 percent—in other words, only this percentage of the heat energy of the fuel is transformed into mechanical energy. 
     Diesel Engines 
     Theoretically, the Diesel cycle differs from the Otto cycle in that combustion takes place at constant volume rather than at constant pressure. Most Diesels are also four-stroke engines but they operate differently than the four-stroke Otto-cycle engines. The first, or suction, stroke draws air, but no fuel, into the combustion chamber through an intake valve. On the second, or compression, stroke the air is compressed to a small fraction of its former volume and is heated to approximately 440° C. (approximately 820° F.) by this compression. At the end of the compression stroke, vaporized fuel is injected into the combustion chamber and burns instantly because of the high temperature of the air in the chamber. Some Diesels have auxiliary electrical ignition systems to ignite the fuel when the engine starts and until it warms up. This combustion drives the piston back on the third, or power, stroke of the cycle. The fourth stroke, as in the Otto-cycle engine, is an exhaust stroke. 
     The efficiency of the Diesel engine, which is in general governed by the same factors that control the efficiency of Otto-cycle engines, is inherently greater than that of any Otto-cycle engine and in actual engines today is slightly more than 40 percent. Diesels are, in general, slow-speed engines with crankshaft speeds of 100 to 750 revolutions per minute (rpm) as compared to 2500 to 5000 rpm for typical Otto-cycle engines. Some types of Diesel, however, have speeds up to 2000 rpm and even higher. Because Diesels use compression ratios of 14 or more to 1, they are generally more heavily built than Otto-cycle engines, but this disadvantage is counterbalanced by their greater efficiency and the fact that they can be operated on less expensive fuel oils. 
     Two-Stroke Engines 
     By suitable design it is possible to operate an Otto-cycle or Diesel as a two-stroke or two-cycle engine with a power stroke every other stroke of the piston instead of once every four strokes. The efficiency of such engines is less than that of four-stroke engines, and therefore the power of a two-stroke engine is always less than half that of a four-stroke engine of comparable size. 
     The general principle of the two-stroke engine is to shorten the periods in which fuel is introduced to the combustion chamber and in which the spent gases are exhausted to a small fraction of the duration of a stroke instead of allowing each of these operations to occupy a full stroke. In the simplest type of two-stroke engine, sleeve valves or ports (openings in the cylinder wall that are uncovered by the piston at the end of its outward travel) replace the poppet valves. In the two-stroke cycle, the fuel mixture or air is introduced through the intake port when the piston is fully withdrawn from the cylinder. The compression stroke follows, and the charge is ignited when the piston reaches the end of this stroke. The piston then moves outward on the power stroke, uncovering the exhaust port and permitting the gases to escape from the combustion chamber. 
     Rotary Engine 
     In the 1950s the German engineer Felix Wankel developed an internal-combustion engine of a radically new design, in which a three-cornered rotor turning in a roughly oval chamber replaces the piston and cylinder. The fuel-air mixture is drawn in through an intake port and trapped between one face of the turning rotor and the wall of the oval chamber. The turning of the rotor compresses the mixture, which is ignited by a spark plug. The exhaust gases are then expelled through an exhaust port through the action of the turning rotor. The cycle takes place alternately at each face of the rotor, giving three power strokes for each turn of the rotor. Because of the Wankel engine&#39;s compact size and consequent lesser weight as compared with the piston engine, it appeared to be an important option for automobiles. In addition, its mechanical simplicity provided low manufacturing costs, its cooling requirements were low and its low center of gravity made it safer to drive. A line of Wankel-engine cars was produced in Japan in the early 1970s, and several United States automobile manufacturers researched the idea as well. However, production of the Wankel engine was discontinued as a result of its poor fuel economy and its high pollutant emissions. 
     Stratified Charge Engine 
     A modification of the conventional spark-ignition piston engine, the stratified charge engine is designed to reduce emissions without the need for an exhaust-gas re-circulation system or catalytic converter. Its key feature is a dual combustion chamber for each cylinder with a pre-chamber that receives a rich fuel-air mixture while the main chamber is charged with a very lean mixture. The spark ignites the rich mixture that in turn ignites the lean main mixture. The resulting peak temperature is low enough to inhibit the formation of nitrogen oxides, and the mean temperature is sufficiently high to limit emissions of carbon monoxide and hydrocarbon. 
     Research on modifications of conventional engines as well as alternatives to conventional engines continues. Some of these options include a modified version of the two-stroke engine, the twin engine (a combination of an internal-combustion engine and an electric engine), and the Stirling engine. 
     Stirling Engine 
     Stirling engine is a type of engine that derives mechanical power from the expansion of a confined gas at a high temperature. The engine was patented in 1816 by the Scottish clergyman Robert Stirling and was used as a small power source in many industries during the 19th and early 20th centuries. The need for automobile engines with low emission of toxic gases has revived interest in the Stirling engine, and prototypes have been built with up to 500 hp and with efficiencies of 30 to 45 percent. Common internal-combustion engines have efficiencies in the range of 20 to 30 percent. 
     The cycle that provides the work is called the Stirling cycle. It consists in its simplest form of the compression of a fixed amount of so-called working gas (hydrogen or helium) in a cool chamber. This cool compressed gas is transferred to a hot chamber, which is heated by an external burner, where the gas expands and drives a piston that delivers the work. The expanded hot gas is then cooled and returned to the cold chamber, and the cycle begins again. The engine is able to transform heat into work because the expansion of the gas at high temperature delivers more work than is required to compress the same amount of gas at low temperature. 
     An external continuous burner that can operate on gasoline, alcohol, natural gas, propane, or butane, provides the heat for the expansion chamber, and the exhaust generated has very low free carbon and toxic gas levels. The Stirling engine runs smoothly because pressure variations in the compression and expansion chambers are sinusoidal, that is, relatively gradual, rather than explosive as in internal-combustion cycles. The necessity of rapid removal of heat from the hot working gas requires a large radiator, which makes this type of engine less suited to small automobiles. 
     The Scuderi Split-Cycle Engine 
     The Scuderi Split-Cycle Engine presently under development divides (or splits) the four strokes of the Otto cycle over a paired combination of one compression cylinder and one power (or expansion) cylinder, operating in principle like the Brayton engine in 1873. These two cylinders perform their respective functions once per crankshaft revolution. 
     The concept is shown in  FIG. 1  where an intake charge is drawn into the compression cylinder  10  through an intake gas passage way and through typical poppet-style valves. Gas is compressed in the compression cylinder  10  and transferred to a compressed gas accumulator  14  and/or the power cylinder  12  through a crossover gas passage, which acts as the intake port for the power cylinder  12 . The crossover gas passage includes a set of uniquely timed valves, which maintain a pre-charged pressure in the compressed gas accumulator  14  through all four strokes of the cycle. A check valve is used to prevent reverse flow from the crossover gas passage to the compression cylinder  10 . Likewise a poppet-style valve prevents reverse flow from the power cylinder  12  to the crossover passage during the power and exhaust strokes. 
     Shortly after the piston in the power cylinder  12  reaches its top dead center position, the gas is quickly transferred to the power cylinder  12  and fired (or combusted) to produce the power stroke. The exhaust gases are pumped out of the power cylinder  12  during its return exhaust stroke through a typical poppet valve to the exhaust passage way. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic view showing the concept of the prior art Scuderi Split-Cycle Engine; 
         FIG. 2A  is a cross section view showing the first embodiment of the apparatus of the present invention; 
         FIG. 2B  is a longitudinal cross section view showing 2 adjacent cylinders of the first embodiment of the apparatus of the present invention along line  2 B- 2 B of  FIG. 2A ; 
         FIG. 3A  is a plan view showing the top of the engine block of the first embodiment of the apparatus of the present invention; 
         FIG. 3B  is a longitudinal cross section view of the engine block along line  3 B- 3 B of  FIG. 3A ; 
         FIG. 3C  is a plan view showing the bottom of the engine block of the first embodiment of the apparatus of the present invention; 
         FIG. 3D  is a cross section view of the engine block along line  3 D- 3 D of  FIG. 3A ; 
         FIG. 3E  is a cross section view along line  3 E- 3 E of  FIG. 3A  showing the nozzle for fuel injector or spark plug in the pre-combustion chamber in the engine block of the first embodiment of the apparatus of the present invention; 
         FIG. 4A  is a plan view showing the top of the engine head-block of the first embodiment of the apparatus of the present invention; 
         FIG. 4B  is a longitudinal cross section view of the engine head-block along line  4 B- 4 B of  FIG. 4A ; 
         FIG. 4C  is a plan view showing the bottom of the engine head-block of the first embodiment of the apparatus of the present invention; 
         FIG. 4D  is a cross section view of the engine head-block along line  4 D- 4 D of  FIG. 4A ; 
         FIG. 5A  is a cross section view showing the piston assembly of the first embodiment of the apparatus of the present invention; 
         FIG. 5B  is a cross section view of the piston assembly along line  5 B- 5 B of  FIG. 5A ; 
         FIG. 5C  is a top view of the piston assembly as seen from line  5 C- 5 C of  FIG. 5A ; 
         FIG. 5D  is a bottom view of the piston assembly as seen from line  5 D- 5 D of  FIG. 5A ; 
         FIG. 6  is a cross section view showing the induction cycle of the first embodiment of the apparatus of the present invention; 
         FIG. 7  is a cross section view showing the bottom center position of the piston assembly at the end of the induction cycle and at the start of the compression cycle of the first embodiment of the apparatus of the present invention; 
         FIG. 8  is a cross section view showing the compression cycle of the first embodiment of the apparatus of the present invention; 
         FIG. 9  is a cross section view showing the top center position of the piston assembly at the end of the compression cycle and at the start of the expansion cycle of the first embodiment of the apparatus of the present invention; 
         FIG. 10  is a cross section view showing the expansion cycle of the first embodiment of the apparatus of the present invention; 
         FIG. 11  is a cross section view showing the bottom center position of the piston assembly at the end of the expansion cycle and at the start of the exhaust cycle of the first embodiment of the apparatus of the present invention; 
         FIG. 12  is a cross section view showing the exhaust cycle of the first embodiment of the apparatus of the present invention; 
         FIG. 13  is a cross section view showing the scavenging of the waste gases starting with the piston assembly at about 10 degrees before the top center in the first embodiment of the apparatus of the present invention; 
         FIG. 14A  shows a typical Otto-cycle in pressure-volume plane; 
         FIG. 14B  shows a typical Diesel-cycle in pressure-volume plane; 
         FIG. 14C  shows the two-stroke super-charged Air-cycle in pressure-volume plane of the first embodiment of the apparatus of the present invention; 
         FIG. 14D  shows the four-stroke super-charged Mixed-cycle (constant-volume and constant-pressure combustion) in pressure-volume plane of the apparatus of the present invention; 
         FIG. 14E  shows the single-stroke super-charged Mixed-cycle (constant-volume and constant-pressure combustion) in pressure-volume plane of the apparatus of the present invention; 
         FIG. 15  is a cross section view showing the second embodiment of the apparatus of the present invention; 
         FIG. 16A  is a cross section view showing a flexible piston rod at the bottom center piston position with the second embodiment of the apparatus of the present invention; 
         FIG. 16B  is a cross section view showing a flexible piston rod in the mid-expansion piston position with the second embodiment of the apparatus of the present invention; 
         FIG. 17  is a cross section view showing the third embodiment of the apparatus of the present invention; 
         FIG. 18  is a cross section view showing two opposing cylinders of the third embodiment of the apparatus of the present invention; 
         FIG. 19A  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where cylinder  1  is in expansion/exhaust stroke while cylinder  2  is in intake/compression stroke; 
         FIG. 19B  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where cylinder  1  is in intake/exhaust stroke while cylinder  2  is in expansion/compression stroke; 
         FIG. 19C  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where cylinder  1  is in intake/compression stroke while cylinder  2  is in expansion/exhaust stroke; 
         FIG. 19D  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where cylinder  1  is in expansion/compression stroke while cylinder  2  is in intake/exhaust stroke; 
         FIG. 19E  shows the hatch patterns of the intake, compression, expansion and exhaust used in the figures from  19 A through  19 D; 
         FIG. 20A  shows the expansion/compression phase of a one-stroke 2-cycle cylinder during the piston up-stroke of the third embodiment of the apparatus of the present invention; 
         FIG. 20B  shows the exhaust/compression phase of a one-stroke 2-cycle cylinder during the piston up-stroke of the third embodiment of the apparatus of the present invention; 
         FIG. 20C  shows the intake-scavenging/compression phase of a one-stroke 2-cycle cylinder during the piston up-stroke of the third embodiment of the apparatus of the present invention; 
         FIG. 20D  shows the expansion/compression phase of a one-stroke 2-cycle cylinder during the piston down-stroke of the third embodiment of the apparatus of the present invention; 
         FIG. 20E  shows the exhaust/compression phase of a one-stroke 2-cycle cylinder during the piston down-stroke of the third embodiment of the apparatus of the present invention; 
         FIG. 20F  shows the intake-scavenging/compression phase of a one-stroke 2-cycle cylinder during the piston down-stroke of the third embodiment of the apparatus of the present invention; 
         FIG. 20G  shows the hatch patterns of the intake, compression, expansion and exhaust used in the figures from  20 A through  20 F; 
         FIG. 21A  is a cross section view showing two opposing cylinders of the fourth and preferred embodiment of the apparatus of the present invention, where both piston rods are connected to the same crankshaft pin; 
         FIG. 21B  is a cross section view along line B-B of  FIG. 21A  showing two opposing cylinders of the fourth and preferred embodiment of the apparatus of the present invention, where both piston rods are connected to the same crankshaft pin; 
         FIG. 21C  is a cross section view along line C-C of  FIG. 21B  showing the location of the cooling liquid flow weir pins in the cooling chambers of the fourth and preferred embodiment of the apparatus of the present invention; 
         FIG. 21D  is a cross section view along line A-A of  FIG. 21A  showing the pre-combustion chambers of the fourth and preferred embodiment of the apparatus of the present invention; 
         FIG. 22  shows the cooling liquid flow pattern over the weir pins in the cooling chambers of the fourth and preferred embodiment of the apparatus of the present invention; 
         FIG. 23A  shows the expansion/compression phase of two opposing 2-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention, where both piston rods are connected to the same crankshaft pin; 
         FIG. 23B  shows the exhaust/compression phase of two opposing 2-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention, where both piston rods are connected to the same crankshaft pin; 
         FIG. 23C  shows the intake-scavenging/compression phase of two opposing 2-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention, where both piston rods are connected to the same crankshaft pin; 
         FIG. 23D  shows the hatch patterns of the intake, compression, expansion and exhaust used in the figures from  23 A through  23 C; 
         FIG. 24A  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod to function as a compressor, where cylinder  1  is in intake/exhaust stroke while cylinder  2  is in expansion/compression stroke; 
         FIG. 24B  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod to function as a compressor, where cylinder  1  is in intake/compression stroke while cylinder  2  is in expansion/exhaust stroke; 
         FIG. 24C  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod to function as a compressor, where cylinder  1  is in expansion/compression stroke while cylinder  2  is in intake/exhaust stroke; 
         FIG. 24D  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod to function as a compressor, where cylinder  1  is in expansion/exhaust stroke while cylinder  2  is in intake/compression stroke; 
         FIG. 24E  shows the hatch patterns of the intake, compression, expansion and exhaust used in the figures from  24 A through  24 D; 
         FIG. 25A  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air pistons and ring shaped combustion pistons are formed as one unit to function as a compressor. Cylinder  1  is in intake/exhaust stroke while cylinder  2  is in expansion/compression stroke; 
         FIG. 25B  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air pistons and ring shaped combustion pistons are formed as one unit to function as a compressor. Cylinder  1  is in intake/compression stroke while cylinder  2  is in expansion/exhaust stroke; 
         FIG. 25C  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air piston heads are aligned with the ring shaped combustion pistons and formed as one unit to function as a compressor. Cylinder  1  is in expansion/compression stroke while cylinder  2  is in intake/exhaust stroke; 
         FIG. 25D  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air pistons have one common head in the middle and form one unit with the ring shaped combustion pistons to function as a compressor. Cylinder  1  is in expansion/exhaust stroke while cylinder  2  is in intake/compression stroke; and 
         FIG. 25E  shows the hatch patterns of the intake, compression, expansion and exhaust used in the figures from  25 A through  25 D. 
         FIG. 26A  shows the expansion/compression phase of a single 2-cycle cylinder. 
         FIG. 26B  shows the exhaust/compression phase of a single 2-cycle cylinder. 
         FIG. 26C  shows the intake-scavenging/compression phase of a single 2-cycle cylinder. 
         FIG. 26D  shows the hatch patterns of the intake, compression, expansion and exhaust used in the figures from  FIG. 26A  through  FIG. 26G . 
         FIG. 26E  through  FIG. 26G  show the expansion/compression phase, the exhaust/compression phase, and the intake-scavenging/compression phase of the same single 2-cycle cylinder during the piston down-stroke; 
         FIGS. 27-29  shows another embodiment where an air sump chamber is employed. 
     
    
    
     EMBODIMENTS OF THE PRESENT INVENTION 
     It is believed that a clearer understanding of the present invention will be obtained by first describing somewhat briefly the main components of the apparatus of the first embodiment of the present invention, followed by a general description of its operation. After this, there will be an introduction to basic thermodynamics followed by an air-standard analysis of the combustion-engine process comparing the performance of prior art engines with the present invention. Then there will be descriptions of further embodiments. 
     Reference is made to  FIG. 2A , which shows the cross section view, and  FIG. 2B , which shows the longitudinal cross section view of 2 adjacent cylinders. 
     As shown in  FIG. 2 , there is an engine  20  schematically shown in a cross-sectional view. In the broader scope, an engine is defined as a device to convert energy; however, in a preferred form, the engine  20  is an internal combustion engine which is shown in various embodiments further described herein. In general, the engine  20  comprises a casing  22 , a piston assembly  24 , a cooling system  26 , a fuel injection and ignition system  28  and an exhaust assembly  30 . 
     Since the engine block  32 , the head block  34  and the piston assembly primarily  24  are substantially different from the known prior art, only these three components of the first embodiment of the present invention will be described in detail. All the other components of the present invention as listed above are typically more or less of same design as in the prior art internal-combustion engines and will therefore be referred to by name and reference number only without further description. 
     Engine Block 
     Referring ahead now to  FIGS. 3A-3E  and  FIGS. 4A-4D , it can be seen that the casing  22  in general comprises an engine block  32  and a head block  34 . 
     With reference to  FIG. 3B , which shows the longitudinal cross section of two adjacent cylinders, the engine block  32  has a cylinder shape bore  36 , which forms the outside wall surface  38  of a cylindrical air chamber  40 . Radially outwards from the cylindrical air chamber  40  there is an annular shape combustion chamber  42 , which has a circular inner wall surface  44 , circular outer wall surface  46 , closed annular shaped top surface  48 , and an annular shaped open end  50  opposite of the top surface  48 . The bottom end of the circular inner wall surface  44  of the annular shape combustion chamber  42  and the bottom end of the outside wall surface  38  of the cylindrical air chamber  40  defines an annular shape surface  52 . It will be shown later in connection with the description of the piston assembly how this surface  52  fits inside the annular shape air chamber in the piston assembly. In the lower part of the circular inner wall surface  44  of the annular shape combustion chamber  42  there is a circular groove  54  for an oil scraper ring of conventional design. 
     There will now be a discussion of the cooling system  26 . In one form it forms a portion of the engine block number  32 . For the sake of clarity, the right-hand portion in  FIG. 3B  will be used to disclose the chambers which in part define the cooling system  26 . 
     Radially outward from the annular shape combustion chamber  42  there is an annular shaped cooling chamber  60 , which has a circular inner wall surface  62 , circular outer wall surface  64 , closed annular shaped top surface  66  and an annular shaped open end  68  opposite of the top surface  66 . This open end  68  is closed with an annular shaped threaded or welded cover  70 . As shown in  FIG. 3D , the cooling liquid inlet port  74  and outlet port  76  are typically threaded and connected to outside tubing for cooling liquid or gas transfer. As shown in  FIG. 3B , on the top of the engine block there is a circular groove  81  surrounding the circular shape air chamber  40  forming the bottom half of the cross section of another cooling chamber  80 . This circular groove  81 , which can be seen in  FIG. 3A , which shows the top plan view of the engine block, is connected by two additional grooves  82  and  84  opposite from each other with the outside edges  86  and  88  of the engine block  32  to form the bottom half of the cross section of the cooling liquid inlet  90  and outlet  92  channels. The inlet port  90  and outlet port  92  ends of these groves are typically threaded and connected to outside tubing for cooling liquid or gas transfer. The cooling liquid is typically water or oil, but air or other gases could as well be used for the cooling purpose. 
     With reference to  FIG. 3D , on the top of the annular shape combustion chamber  42  there are two fixed volume pre-combustion chambers  94  and  96 . The “fixed volume” description is used to differentiate these pre-combustion chambers from the main annular shape combustion chamber  42 , which volume varies with the in-and-out stroke of the piston. 
     The left side pre-combustion chamber  94  has a fuel injector nozzle  98  for fuel injection in the Diesel-cycle engine version and an additional spark plug nozzle  106  in the Otto-cycle engine version. The right side fixed volume combustion chamber  96  does not have any nozzles in it making it a supercharged combustion air supply chamber  100 . Its function will be described later in connection with the operation of the present invention. 
     Both the pre-combustion chamber  94  and the supercharged combustion air supply chamber  100  communicate with the main annular shape combustion chamber  42  through openings  102  and  104  in the bottom of the fixed volume chambers at their end just above the annular shape combustion chamber  42 . These openings  102  and  104  can be seen in the plan view in  FIG. 3A , which shows the top plan view of the engine block  32  with two adjacent cylinders. In the first embodiment of the present invention the pre-combustion chamber  94  and the supercharged combustion air supply chamber  100  are at opposite sides of the engine block. However, more than two of the fixed volume pre-combustion chambers can be used in large diameter engines of the preferred embodiments. 
     The fuel injector nozzle  98  and the spark plug nozzle  106  are shown side-by-side penetrating the engine block  32  sidewall  35  into the pre-combustion chamber  94  in  FIG. 3A . 
       FIG. 3E  shows the threaded cross section of the nozzles along line  3 E- 3 E of  FIG. 3A . 
     The bottom of the engine block is shown in plan view in  FIG. 3C . 8 holes  110  through the engine block surrounding each cylinder are shown in this view as well as in the top plan view in  FIG. 3A  for the engine assembly tie rods. 
     Head Block 
     The construction of the head block  34  is shown in  FIG. 4A  through  FIG. 4D . 
     The top of the engine head block  34  covering two adjacent cylinder regions  120  and  122  is shown in plan view in  FIG. 4A . The ambient air intake port  124  and the supercharged air discharge port  126  are in the middle of the head block just above the cylindrical air chamber  40  in the engine block  32  of  FIG. 3B . The ambient air intake port  124  is typically larger than the supercharged air discharge port  126  since the volume of the supercharged air is smaller. The supercharged air intake port  129  is shown on the topside of  FIG. 4A  just above the pre-combustion chamber  94  in the engine block (see  FIG. 3B ). The exhaust port  130  for the waste gases is shown on the bottom side of  FIG. 4A  just above the supercharged combustion air supply chamber  100  in the engine block  32 . 
       FIG. 4B  shows the longitudinal cross section view of the head block  34  along line A-A of  FIG. 4A . 
     The bottom of the engine head block in plan view is shown in  FIG. 4C . 
     In the bottom of the head block there is a circular groove  134  surrounding the top of the circular shape air chamber  40  (see  FIG. 3B ) forming the top half of the cross section of another cooling chamber  80 . This circular groove  134  is connected by two additional grooves  136  and  138  opposite from the grooves  82  and  84  (see  FIG. 3A ) and in communication with the cooling liquid inlet  90  and outlet  92  channels. The inlet port  90  and outlet port  92  ends of these groves are typically threaded and connected to outside tubing for cooling liquid or gas transfer. The cooling liquid is typically water or oil, but air or other gases could as well be used for the cooling purpose. 
     The 8 larger holes  140  around each cylinder head are for the tie rods of the engine assembly and correspond in location with holes  110  as shown in  FIG. 3C . The smaller holes  142  around the air ports are for the attachment of the respective valve blocks and overhead cam assemblies to the head block  34 . 
     The cylindrical shape recess  146  in the bottom of the head block  34  is an extension of the cylindrical air chamber  40  of the engine block  32 . 
       FIG. 4D  is a cross section view of the engine head block along line B-B of  FIG. 4A . The conical shaped recesses  148  in the head block form the seats for the valve heads. 
     Piston Assembly 
     The construction of the piston assembly is shown in  FIG. 5A  through  FIG. 5D . 
     With reference to  FIG. 5A  the piston assembly  24  comprises a cylindrical shape air piston  150  that fits loosely, typically with a 1-2 mm clearance, inside the cylindrical air chamber  40  in the engine block  32 , and an annular shape combustion piston  152  that fits tightly inside the annular shape combustion chamber  42  in the engine block  32 . The top surface  154  of the annular shape combustion piston forms the moving bottom of the annular shape combustion chamber  42  and is typically coated with very heat resistant material, which may be constructed of either metal, ceramics, or other materials. 
     The outside cylindrical wall  156  of the annular shape piston  152  has typically 3 or more circular grooves  158  for piston rings. Two or more of the upper grooves are for compression rings and one or two of the lower grooves are for oil scraper rings. The inside cylindrical wall  160  of the annular shape piston  152  has typically 3 or more circular grooves  162  for piston rings. Two or more of the upper grooves are for compression rings and one or two of the lower grooves are for oil scraping rings. The function, shape and fit of the piston rings is typically the same as in the prior art internal-combustion engines and therefore will not be described here in more detail. Oil consumption is controlled principally by the use of slotted oil rings. However, it is the combination of compression and oil scraping rings that determines the oil consumption of the engine. 
     The inside cylindrical wall  160  of the annular shape piston  152  continues downward the distance of the stroke of the piston and meets the lower part of the outer wall  164  of the cylindrical shape air piston  150  to form an annular shape air chamber  166 . The open annular shape top end  168  of this annular shape air chamber  166  is covered by the stationary annular shape face  52  that is formed between the bottom of the inner wall  44  of the combustion chamber  42  and the bottom of the inner wall  38  of the circular air chamber  40  in the engine block  32 . (See  FIG. 3B ). The in-and-out motion of the piston assembly  24  varies the volume of this annular air chamber  166  in the piston assembly. Thus, during the induction and expansion strokes ambient air is drawn in from the cylindrical air chamber  40  to be compressed to supercharged air during the compression and exhaust strokes and to exit again through the cylindrical air chamber  40  into the supercharged air accumulator  190 . 
     The small bore openings  182  below and around the periphery of the ring shape piston allow lubricating (and cooling) oil from the crankcase to enter and exit between the adjacent surfaces of the annular air chamber  166  in the piston assembly  24  and the annular shape combustion chamber  42 . 
     With reference to  FIG. 5B  the top end of the cylindrical air piston  150  has a horizontal bore  170  to receive the piston pin  176 . The piston pin bearing  174 , bearing housing  178 , and connecting rod  180  are of conventional design as used in prior art internal-combustion engines and will therefore not be described here in more detail. 
     General Description of Operation 
     1 Induction 
     Reference is made to  FIG. 6  which is a cross section view showing the induction cycle of the first embodiment. As shown in  FIG. 6 , the induction stroke, during which the piston assembly  24  is moving outwards, starts with the supercharged air intake valve  129   a  open. The supercharged air discharge valve  126   a  and the exhaust valve  130   a  are closed. Supercharged air, typically at 3 atm pressure and 105°-110° C. temperature, from the supercharged air accumulator  190  enters the pre-combustion chamber  94  and flows from there through the opening  102  at the bottom end of the pre-combustion chamber  94  into the top of the annular shape combustion chamber  42  to fill it with supercharged air. At the opposite side from the pre-combustion chamber  94  the supercharged combustion air supply chamber  100  is filled with supercharged air flowing from the top of the annular shape combustion chamber through the opening  104  at the bottom end of the supercharged combustion air supply chamber  100 . 
     The ambient air intake valve  194  at the top of the cylindrical air chamber  40  remains closed until the residual supercharged air in the air chamber has reached the ambient pressure, typically at crankshaft position about 35 degrees after top center, at which time the ambient air intake valve  194  opens. Ambient air is drawn through the ambient air intake port  124  into the cylindrical air chamber  40 . Simultaneously air is drawn into the annular air chamber  166  in the piston assembly  24  from the cylindrical air chamber  40  through the typically 1-2 mm wide annular shape clearance between the inside wall surface  38  of the cylindrical air chamber  40  and the outside wall surface  164  of the cylindrical air piston  150 . 
     The incoming air cools the inside wall  44  of the annular shape combustion chamber  42  while flowing through the typically 1-2 mm wide annular shape passage way  165  into the annular shape air chamber  166  in the piston assembly  24 . 
     2 Compression 
       FIG. 7  is a cross section view showing the bottom center position of the piston assembly at the end of the induction cycle and at the start of the compression cycle of the first embodiment of the apparatus of the present invention. 
     As shown in  FIG. 7  the induction stroke ends and the compression stroke starts typically with all valves closed and the piston assembly at the bottom center position. However, in order to attain high output at high engine speed it has been found that the intake valves should be closed appreciably after bottom dead center or after the compression stroke has started. Thus use can be made of the inertia of the flowing air to ram considerably more charge into the cylinder. 
       FIG. 7  shows the pre-combustion chamber  94 , the annular combustion chamber  42  and the supercharged combustion air supply chamber  100  filled with supercharged air, typically at 3 atm pressure and 105-110 degrees C. temperature. It also shows the cylindrical air chamber  40  in the engine block  32  and the annular air chamber  166  in the piston assembly  24  filled with ambient air. 
       FIG. 8  is a cross section view showing the compression cycle of the first embodiment of the apparatus of the present invention. 
     As shown in  FIG. 8  the in-motion of the piston assembly compresses the air in all three combustion chambers  94 ,  42 , and  100  as well as in the two air chambers, the annular air chamber  166  in the piston assembly and the cylindrical air chamber  40  in the engine block. 
     Typically at the crankshaft position about 75 degrees before top center the supercharged air pressure has reached 3 atm and the supercharged air discharge valve  196  opens to let the compressed air into the supercharged air accumulator  190  of  FIG. 2 . The supercharged air discharge valve  196  is closed again as the crankshaft reaches top dead center piston position. 
     To reach the 3 atm pressure with the supercharged air the nominal compression ratio of the air chambers is typically about 2.2:1. The nominal compression ratio (usually specified) is the ratio between the maximum and minimum volumes of the air chambers. 
     The outgoing air cools the inside wall  44  of the annular shape combustion chamber  42  while flowing through the typically 1-2 mm wide annular shape passage way  165  from the annular shape air chamber  166  in the piston assembly to the cylindrical shape air chamber  40  in the engine block  32 . 
     At the end of the compression stroke the pressure in the combustion chambers is typically 40-45 atm and the temperature 350-400 degrees C. To reach this pressure with the supercharged air at 3 atm pressure and 105-110 degrees C. temperature the nominal compression ratio of the combustion chambers is typically about 7:1. The actual compression ratio is somewhat less than the nominal value because of late intake valve closing in high-speed engines. 
     3 Expansion 
       FIG. 9  is a cross section view showing the top center position of the piston assembly at the end of the compression cycle and at the start of the expansion cycle of the first embodiment of the apparatus of the present invention. 
     As shown in  FIG. 9  the compression stroke ends and the expansion stroke starts typically with all valves closed and the piston assembly at the top center position. Typically about 2-5 degrees before top center the fuel injector  28   a  begins to introduce the fuel progressively into the supercharged air in the fixed volume pre-combustion chamber  94  prior to inflammation. At top center the spark plug  28   b  ignites the charge in Otto-cycle engine and in the Diesel-engine the temperature of the compressed air is sufficiently high to ignite the fuel. The pressure after ignition climbs typically to 90-135 atm and the temperature to 1200-1400 degrees C. 
     The flame in the combustion chambers continues to burn as long as the fuel injector feeds fuel into the pre-combustion chamber and the highly turbulent airflow from the compressed supercharged combustion air supply chamber  100  provides the oxygen for the combustion. Typically an air-fuel mixture ratio 7:1 by weight is used on the rich side and 20:1 on the lean side for gasoline engines. Rich mixtures are used to suppress combustion knock and to obtain maximum engine output, and lean mixtures are used to obtain minimum fuel consumption. 
       FIG. 10  is a cross section view showing the expansion cycle of the first embodiment of the apparatus of the present invention. 
     The apparatus of the disclosure in one form is ideally suited for using rich mixtures in the pre-combustion chamber  94  to suppress combustion knock and lean mixtures in the main annular shape combustion chamber  42  by sizing the volumes of the pre-combustion chamber  94  and the compressed supercharged combustion air supply chamber  100  relative to each other so that optimum conditions are reached between desired engine output and fuel consumption. 
     The ambient air intake valve  194  at the top of the cylindrical air chamber  40  remains closed until the residual supercharged air in the piston air chambers has reached the ambient pressure, typically at crankshaft position about 35 degrees after top center, at which time the ambient air intake valve  194  opens. Ambient air is drawn through the ambient air intake port  124  into the cylindrical air chamber  40 . Simultaneously air is drawn into the annular air chamber  166  in the piston assembly  24  from the cylindrical air chamber  40  through the typically 1-2 mm wide annular shape clearance between the inside wall surface  38  of the cylindrical air chamber  40  and the outside wall surface  164  of the cylindrical air piston  150 . 
     Again, as during the induction stroke, the incoming air cools the inside wall  44  of the annular shape combustion chamber  42  while flowing through the typically 1-2 mm wide annular shape passage way  165  into the annular shape air chamber  166  in the piston assembly  24 . 
     20-30 degrees before the bottom center the exhaust valve  130   a  opens to release the waste gases. 
     4 Exhaust 
       FIG. 11  is a cross section view showing the bottom center position of the piston assembly at the end of the expansion cycle and at the start of the exhaust cycle of the first embodiment of the apparatus of the present invention. 
       FIG. 11  shows the piston assembly  24  at the bottom center position. The ambient air intake valve  194  is closed and the exhaust stroke begins with the exhaust valve  130   a  already open. The pre-combustion chamber  94 , the combustion chamber  42  and the supercharged combustion air supply chamber  100  are full with substantially waste gases. The cylindrical air chamber  40  in the engine block  32  and the annular air chamber  166  in the piston assembly  24  are filled with ambient air. 
       FIG. 12  is a cross section view showing the exhaust cycle of the first embodiment of the apparatus of the present invention. 
     As shown in  FIG. 12  the in-motion of the piston assembly displaces the waste gases from all three combustion chambers  94 ,  42 , and  100  to flow out through the exhaust valve  130   a  into the exhaust gas accumulator  30   a.    
     Again, typically at the crankshaft position about 75 degrees before top center the supercharged air pressure in the cylindrical air chamber  40  in the engine block  32  and the annular air chamber  166  in the piston assembly  24  has reached 3 atm and the supercharged air discharge valve  196  opens to let the compressed air into the supercharged air accumulator  190  of  FIG. 2A . The supercharged air discharge valve  196  is closed again at top center piston position before the next induction stroke begins. 
     Again, as during the compression stroke, the outgoing air cools the inside wall  44  of the annular shape combustion chamber  42  while flowing through the typically 1-2 mm wide annular shape passage way  165  from the annular shape air chamber  166  in the piston assembly to the cylindrical shape air chamber  40  in the engine block  32 . 
       FIG. 13  is a cross section view showing the scavenging of the waste gases starting with the piston assembly at about 10 degrees before the top center in the first embodiment of the apparatus of the present invention. 
       FIG. 13  shows how typically at about 10 degrees before top center the supercharged air intake valve  129   a , opens to let supercharged air flow into the pre-combustion chamber  94  to scavenge the waste gases from the combustion chambers before the exhaust valve  130   a  is closed at top center and the next induction stroke begins. 
     With reference back to  FIG. 2A  there is a check valve  190   a  in the passage way  190   b  between the supercharged air accumulator  190  and the fixed volume pre-combustion chamber  94  to stop any waste gases from entering the accumulator at the start of the waste gas scavenging with the supercharged air at the end of the exhaust cycle. 
     The exhaust valve  130   a  releases the waste gases into a waste gas accumulator  30   a . The waste gas accumulator pressure is maintained substantially below the pressure in the supercharged air accumulator  190  to enable the supercharged air to scavenge the combustion chambers from waste gases at the end of the exhaust cycle. 
     The waste gases are released from the waste gas accumulator  30   a  through a gas turbine  30   b  which drives an electric generator  30   c  and/or an air compressor  30   d . The air compressor  30   d  is used to feed supercharged air into the accumulator  190  to supplement the air supply from the cylindrical air chamber of the engine or to completely eliminate the air supercharge function of the engine. 
     Before the supercharged hot air from the waste-gas-turbine-driven air compressor  30   d  enters the supercharged air accumulator  190  it is cooled with a heat exchanger  30   e  using typically either water or air for cooling to facilitate larger air charge into the engine during the induction cycle. There is a check valve  30   f  in the passage way  30   g  from the air compressor  30   d  to the supercharged air accumulator  190  to stop any back-flow from the accumulator. 
     Before the supercharged hot air from the cylindrical air chamber  40  enters the air accumulator  190  it is cooled with a heat exchanger  190   c  using typically either water or air for cooling to facilitate larger air charge into the engine during the induction cycle. There is a check valve  190   d  in the passage way  190   e  from the cylindrical air chamber  40  to the supercharged air accumulator  190  to stop any back-flow from the accumulator. 
     Basic Thermodynamics 
     It is believed that a better understanding of the air-standard analysis will be possible by referring first to the Ideal Gas Law of thermodynamics. 
     In a gas the molecules move at random, bounded only by the walls of their container. 
     Empirical laws have been developed that correlate macroscopic variables. For common gases, the macroscopic variables include pressure (P), volume (V), and temperature (T). Boyle&#39;s law states that in a gas held at a constant temperature the volume is inversely proportional to the pressure. Charles&#39;s law, or Gay-Lussac&#39;s law, states that if a gas is held at a constant pressure the volume is directly proportional to the absolute temperature. Combining these laws gives the ideal gas law: PV/T=R (per mole), also known as the equation of state of an ideal gas. The constant R on the right-hand side of the equation is a universal constant, the discovery of which is a cornerstone of modern science. 
     If V is expressed as volume per unit weight, the value of constant R will be different for different gases. If V is expressed as the volume of one molecular weight of gas, then the universal gas constant R u  is the same for all gases in any chosen system of units. Hence R=R u /M, where M is the molecular weight of the gas. 
     In general, for any amount of gas, the ideal gas equation becomes pV=NMRT, where V is now the total gas volume, N is the number of moles of gas in the volume V, M is the molecular weight, and R u =MR the universal gas constant. 
     For all ideal gases, R u =MR in lb-ft is 1,546. One pound mol of any perfect gas occupies a volume of 359 cu ft at 32 F and 1 atm. 
     The ideal gas equation of state is only approximately correct. Real gases do not behave exactly as predicted. In some cases the deviation can be extremely large. Thus, modifications of the ideal gas law, PV=RT, were proposed. Particularly useful and well known is the van der Waals equation of state: (P+a/V 2 )(V−b)=RT, where a and b are adjustable parameters determined from experimental measurements carried out on actual gases. They are material parameters rather than universal constants, in the sense that their values vary from gas to gas. 
     In thermodynamics the term “specific heat” refers to the ratio of the amount of heat transferred to raise unit mass of a material 1 deg to that required to raise unit mass of water 1 deg at some specified temperature. Gases have a different specific heat at constant pressure (c p ) from the specific heat at constant volume (c v ). 
     The ratio of these two specific heats define the constant k=c p /c v . 
     For monatomic gases, the specific heats do not vary with temperature, and k, the value of c p /c v , is 1.66. For diatomic gases (oxygen, nitrogen, etc.) the specific heats vary with temperature but for many purposes may be assumed constant over considerable ranges of temperature. For diatomic gases, k is approximately 1.40. 
     Air-Standard Analysis 
     The accurate analysis of combustion-engine processes is a complex problem. Consequently, simplifying assumptions have been introduced, resulting in the air-standard cycle analysis. This analysis implies that the medium is air and that no chemical reaction occurs during the cycle. The specific heat of the air is assumed to be constant. Also, losses by heat transfer from the apparatus to the atmosphere are assumed to be zero in this analysis. 
     The foregoing assumptions result in an analysis that is far from correct for most actual combustion-engine processes, but is of considerable value for indicating the upper limit of performance if infinitely lean air-fuel mixtures could be used. This analysis is also a simple means for indicating the relative effect of the principal variables, such as compression ratio, thermal efficiency of the cycle, and relative size of the apparatus. A measure of this is the mean effective pressure (mep), which is network per cubic inch of displacement. 
     In the air-standard analysis the medium at the end of the process is unchanged and is at the same conditions as at the beginning of the process. Thus the combustion-engine process is treated as a heat-engine cycle in this analysis. 
     In internal-combustion engines, the combustion process is assumed to occur at constant volume, at constant pressure or by a sequence of these two procedures, or in various other ways. 
     The constant-volume process is characteristic of the spark-ignition or Otto-cycle; the constant-pressure is found only in the slow-speed compression-ignition or Diesel-cycle; with both procedures, the cycle is sometimes called limited-pressure cycle and occurs in high-speed compression-ignition engines. 
     The nominal compression ratio (usually specified) is the displacement plus clearance volume divided by the clearance volume. The actual compression ratio is appreciably less than the nominal value because of late intake valve or port closing. 
     The compression pressure may be estimated from the relation p=r a   k  p m , where p m  is the intake-manifold pressure and r a  is the actual compression ratio. 
     For air the value of k is about 1.40 up to compression ratio 10:1 and about 1.39 at compression ratio 14:1. Therefore, in the analysis below, these values have been used for the mean adiabatic exponent during the compression process in the Otto- and Diesel-cycles respectively. However, since the supercharged air is at a higher temperature in the present invention the value of 1.38 has been used for the mean adiabatic exponent during the compression process. For the expansion process the mean adiabatic exponent is typically about 1.22, varying from 1.20 at the beginning to 1.25 at the end of the process for the prior art Otto- and Diesel-cycles. For the analysis of the present invention these same typical values are assumed for the mean adiabatic exponent during the expansion process. 
     Spark-ignition EnginesSpark-ignition engines have compression ratios between 4:1 and 12:1 (limited by combustion knock of the fuel-air mixture), compression pressures from below 7 atm to above 30 atm, and they operate on the Otto-cycle. The combustion pressures are usually 3.5 to 5 times the compression pressures. 
     Advantages are low first cost, low specific weight, low cranking effort required, wide variation obtainable in speed and load, high mechanical efficiency, and fairly low specific fuel consumption at high compression ratios and wide-open throttle. 
       FIG. 14A  shows the pressure-volume diagram of an ideal Otto-cycle engine with the fragmentary line. Air is admitted during the induction stroke from a to b. Air is compressed adiabatically and reversibly from b to c. Heat is added to the air during the constant-volume heating process from c to d. Adiabatic reversible expansion occurs from d to e and the exhaust valve opens at point e′. Waste gases are exhausted during the exhaust stroke from b to a. 
     The solid line shows the characteristics of the actual cycle. The hatched area inside the solid line represents the work done during the cycle. As shown in  FIG. 14A  the physical characters of this engine are: displacement 650 cu cm (40 cu in), stroke 80 mm (3.15 in), piston diameter 102 mm (4 in) and compression ratio 10:1. 
     The mean effective pressure (mep) within the hatched area is 10.5 atm (154 psi). 
     The horsepower equation of a four-stroke cycle engine is:
 
hp=mep psi  displ cu in  rpm/792,000=154×40×2,000/792,000=16 hp or 0.39 hp/cu in at 2,000 rpm (35 hp or 0.88 hp/cu in at 4500 rpm).
 
     The typical range for United States automobile engines is from 0.7 to 1.0 hp/cu in at 4,500 rpm. 
     The torque equation of a four-stroke cycle engine is:
 
Torque in lb ft=mep psi  displ cu in /(4π×12)=154×40/48π=41 lb ft.
 
     Compression-Ignition Engines 
     Compression-ignition engines have compression ratios between 11.5:1 and 22:1 and compression pressures from 27 atm to 48 atm, and they operate on the Diesel-cycle. The combustion pressure is about the same as the compression pressure for constant-pressure combustion and usually 2 times the compression pressure for mixed cycle engines (constant-volume and constant-pressure combustion). 
     Advantages are low specific fuel consumption, ability to maintain economy and thermal efficiency at part loads, low fuel cost, no pre-ignition, practically no carbon monoxide emissions except near full-load or at over-load conditions, and suitability for two-stroke operation. 
       FIG. 14B  shows the pressure-volume diagram of an ideal constant-pressure Diesel-cycle engine with the fragmented line. Air is admitted during the induction stroke from a to b. Air is compressed adiabatically and reversibly from b to c. Heat is added to the air during the constant-pressure heating process from c to d. Adiabatic reversible expansion occurs from d to e and the exhaust valve opens at point e′. Waste gases are exhausted during the exhaust stroke from b to a. 
     The solid line shows the characteristics of the actual cycle. The hatched area inside the solid line represents the work done during the cycle. As shown in  FIG. 14B  the physical characters of this engine are: displacement 650 cu cm (40 cu in), stroke 80 mm (3.15 in), piston diameter 102 mm (4 in) and compression ratio 14:1. 
     The mean effective pressure (mep) within the hatched area is 7.6 atm (112 psi).
 
hp=mep psi  displ cu in  rpm/792,000=112×40×2,000/792,000=11.3 hp or 0.28 hp/cu in at 2,000 rpm.
 
     The typical range for United States automobile Diesel engines is from 0.2 to 0.35 hp/cu in at 2,000 rpm. 
     The torque equation of a four-stroke cycle engine is:
 
Torque in lb ft=mep psi  displ cu in /(4π×12)=112×40/48π=30 lb ft.
 
     Present Invention Engines 
     Present invention engines have compression ratios typically between 6:1 and 18:1 and compression pressures typically from 40 atm to 50 atm, and they operate on the Mixed-cycle, which means constant-volume and constant-pressure combustion. The combustion pressure is usually 2 times the compression pressure. 
     Advantages are low specific weight, high mechanical efficiency, low specific fuel consumption, ability to maintain economy and thermal efficiency at part loads, wide variation obtainable in speed and load, practically no carbon monoxide, hydrocarbon or nitrogen oxide emissions, and suitability for one- and two-stroke operation. 
       FIG. 14C  shows the pressure-volume diagram of an ideal constant-volume and constant-pressure 4-cycle engine of the present invention with the fragmented line. Air is admitted during the induction stroke from a to b. Air is compressed adiabatically and reversibly from b to c. Heat is added to the air during the constant-pressure heating process from c to d. Additional heat is added to the combustion chamber during the constant-pressure heating process from d to e. Adiabatic reversible expansion occurs from e to f and the exhaust valve opens at point f′. Waste gases are exhausted during the exhaust stroke from b to a. 
     The solid line shows the characteristics of the actual cycle. The hatched area inside the solid line represents the work done during the cycle. 
     As shown in  FIG. 14C  the physical characters of this engine are: displacement 649/514 cu cm (average 35.5 cu in), stroke 80 mm (3.15 in), piston diameter 130/165 mm (5.12/6.50 in) and 138/165 mm (5.43/6.50 in), compression ratio 7:1, and supercharged air pressure 3 atm. 
     The mean effective pressure (mep) within the hatched area is 28.2 atm (415 psi).
 
hp=mep psi  displ cu in  rpm/792,000=415×35.5×2,000×2/792,000=74 hp or 2.1 hp/cu in at 2,000 rpm.
 
     The typical range for 4-cycle engines would be from 1.5 to 3.0 hp/cu in at 2,000 rpm. 
     The typical range for 4-cycle engines would be from 1.5 to 3.0 hp/cu in at 2,000 rpm. 
     The torque equation of a two-stroke cycle engine is:
 
Torque in lb ft=mep psi  displ cu in /(2π×12)=415×35.5/24π=195 lb ft.
 
       FIG. 14D  shows the pressure-volume diagram of an ideal 2-stroke air-cycle of the cylindrical air piston of the present invention used in connection with the 4-cycle engine as shown in  FIG. 14C . The physical characters of this air piston are: displacement 1230 cu cm (average 75 cu in), stroke 80 mm (3.15 in), piston diameter 140 mm (5.5 in) and compression ratio 7:1. Air is admitted during the induction stroke from a to b. Air is compressed adiabatically and reversibly from b to c and then released to the compressed air accumulator at constant pressure from c to d. Adiabatic reversible expansion occurs from d to a during the return stroke. 
     Another pressure-volume diagram is shown with letters a′, b′, c′ and d′ since the 2-cycle air piston needs only two strokes to complete a full cycle while the 4-cycle engine needs four strokes to complete a full cycle. 
     The mean pressure within the hatched area is 1.1 atm (16 psi). 
     The consumed horsepower equation of this two-stroke air cycle is:
 
hp=mep psi  displ cu in  rpm/792,000=16×75×2,000×2/792,000=6 hp at 2,000 rpm.
 
     With the above physical dimensions the air piston provides about 75% of the supercharged air volume for the four-cycle engine shown in  FIG. 14C . The balance is produced by the exhaust gas driven air compressor. 
       FIG. 14E  shows the pressure-volume diagram of an ideal constant-volume and constant-pressure one stroke 2-cycle engine of the present invention with the fragmented line. General description of the operation of such embodiment of the present invention will be given later in connection with  FIGS. 20A through 20G . 
     Air is admitted during the induction stroke from a to b. Air is compressed adiabatically and reversibly from b to c. Heat is added to the air during the constant-pressure heating process from c to d. Additional heat is added to the combustion chamber during the constant-pressure heating process from d to e. Adiabatic reversible expansion occurs from e to f and the exhaust valve opens at point f′. Waste gases are scavenged by the incoming supercharged air, and exhausted during the induction stroke from a to b. 
     The solid line shows the characteristics of the actual cycle. The hatched area inside the solid line represents the work done during the stroke. As shown in  FIG. 14E  the physical characters of this engine are: 
     displacement 352/217 cu cm (average 17.5 cu in), stroke 80 mm (3.15 in), piston diameter 130/150 mm (5.12/5.91 in) and 138/150 mm (5.43/5.91 in), compression ratio 7:1, and supercharged air pressure 3 atm. 
     The mean effective pressure (mep) within the hatched area is 27.1 atm (400 psi).
 
hp=mep psi  displ cu in  rpm/792,000=400×17.5×2,000×4/792,000=71 hp or 4.0 hp/cu in at 2,000 rpm.
 
     The typical range for 1-stroke engines would be from 3 to 6 hp/cu in at 2,000 rpm. 
     The torque equation of a one-stroke cycle engine is:
 
Torque in lb ft=mep psi  displ cu in /(π×12)=400×17.5/12π=185 lb ft.
 
     Emission Analysis 
     The key feature for the practically no carbon monoxide, hydrocarbon or nitrogen oxide emissions from the engine of present invention is the use of dual fixed volume combustion chambers for each cylinder. In the preferred embodiment of the present invention the pre-combustion chamber  294  receives a rich fuel-air mixture while the supercharged combustion air chamber  200  is charged with a very lean mixture or none at all. The rich mixture ignites the lean main mixture. The resulting peak temperature is low enough to inhibit the formation of nitrogen oxides, and the mean temperature is sufficiently high to limit emissions of carbon monoxide and hydrocarbon. The fuel-air ratio varies from rich at the pre-combustion chamber  294  to lean at the annular shape combustion chamber  242 . 
     It is the peak temperatures, which occur at the tip of the flame front, that produce most of the nitrogen oxide emissions; the lower the peak temperatures the lower the nitrogen oxide emissions. 
     When piston is racing away from the flame front it produces a cooling effect that results in lower peak temperatures and lower nitrogen oxide emissions. 
     It is a well known fact that combustion efficiencies can be improved by running lean, significantly above 14.5 to 1 air/fuel ratio. 
     The annular shape combustion chamber  242  in combination with the tangential entry of the flame front (as shown later in  FIG. 210 ) from both the pre-combustion chamber  294  and the supercharged combustion air supply chamber  200  produce a massive turbulence that results in an extremely fast burn rate (combustion duration). Burn rate is the amount of time it takes for the trapped fuel/air mixture to completely combust. 
     Burn rate is a powerful multiplier of engine efficiency. 
     Description of the Second Embodiment 
     Reference is made to  FIG. 15  which shows the cross section of the second embodiment of the apparatus of the present invention. It shows the cross section view of one cylinder comprising the following main components, which differ from the main components of the first embodiment: 
     engine block  32 , head block  34 , piston assembly  24 , supercharged air intake valve  129   a , exhaust valve  130   a , supercharged air intake valve block  129   b , exhaust valve block  130   b , fuel injector  28   a , and spark plug  28   b.    
     Since only the engine block  32 , the head block  34  and the piston assembly  24  are of unique design, only these three components of the second embodiment of the present invention will be described in detail. All the other components of the present invention as listed above are typically more or less of same design as in the prior art internal-combustion engines and will therefore be referred to by name and reference number only without further description. 
     Engine Block 
     With reference to  FIG. 15 , which shows the cross section view of one cylinder, the engine block  232  has a cylinder shape bore  236 , which forms the inner wall face  238  of the engine block  232 . This inner wall face defines a cylindrical passageway  239  through the engine block  232 . The piston connecting rod  180  travels in this passage way  239  back and forth with each stroke of the piston  224  transferring the reciprocating motion of the piston into rotary motion of the crankshaft  181 . 
     Radially outwards from the cylinder shape bore  236  there is an annular shape combustion chamber  242 , which has a circular inner wall surface  244 , circular outer wall surface  246 , closed annular shaped bottom surface  248 , and an annular shape open end  249  opposite of the bottom surface  248 . In the upper part of the circular inner wall surface  244  of the annular shape combustion chamber  242  there is a circular groove  254  for an oil scraper ring. 
     There is also an annular shape cooling chamber  260  in the engine block  232  radially outwards from the annular shape combustion chamber  242 . Each cylinder in the engine is surrounded with its own cooling chamber. Each cooling chamber has an inlet  274  and outlet  276  nozzle for forced water or oil circulation. The cooling liquid is typically water or oil, but air or other gases could as well be used for the cooling purpose. 
     With reference still to  FIG. 15 , at the bottom of the annular shape combustion chamber  242  there are two fixed volume combustion chambers  294  and  200 . The “fixed volume” description is used to differentiate these combustion chambers from the main annular shape combustion chamber  242 , which volume varies with the in-and-out stroke of the annular shape bottom end  225  of the piston assembly  224 . 
     The left side pre-combustion chamber  294  has a fuel injector nozzle  298  for fuel injection in the Diesel-cycle engine version and an additional spark plug nozzle  206  in the Otto-cycle engine version. When the right side fixed volume combustion chamber  200  does not have a nozzle in it for fuel injection it functions as a supercharged combustion air supply chamber. However, it can also be equipped with a fuel injector  228   c  in which case it will function as a lean fuel-air mixture combustion chamber  200 . The rich mixture in the pre-combustion chamber  294  ignites the lean mixture in the other fixed volume combustion chamber. 
     Both the pre-combustion chamber  294  and the supercharged combustion air supply chamber  200  communicate with the main annular shape combustion chamber  242  through openings  296  in the top of the fixed volume chambers at their end just below the annular shape combustion chamber  242 . 
     In this second embodiment of the present invention the pre-combustion chamber  294  and the supercharged combustion air supply chamber  200  are at opposite sides of the engine block. However, more than two of the fixed volume combustion chambers can be used in large diameter engines of the present invention. 
     The fuel injector nozzle  298  and the spark plug nozzle  206  are shown side-by-side penetrating the engine block  232  sidewall  226  into the pre-combustion chamber. At right a fuel injector  228   c  is shown penetrating the engine block into the lean fuel/air mixture combustion chamber  200 . 
     Head Block 
     The open end  249  of the annular shape variable combustion chamber  242  is covered with the engine head block  234  comprising a cylinder shape air chamber  240  that is closed at the top end  241  and in which a loose fitting cylinder shape piston  250  slides as described in connection with the first embodiment. 
     The in-and-out motion of the cylinder shape piston  250  varies the volume of the cylindrical air chamber  240  in the engine head block between the circular top end  251  of the cylinder-shaped piston  250  and the closed top end  241  of the cylindrical air chamber  240 . 
     The ambient air intake port  124  and the supercharged air discharge port  126  are in the middle of the head block just above the cylindrical air chamber  240  in the head block  234  as already described with the first embodiment. 
     Ambient air enters the cylindrical air chamber  240  during the piston out-motion and compressed air from the cylindrical air chamber  240  flows to a supercharged air accumulator  190  during the piston in-motion. 
     The annular shape bottom end  225  of the piston assembly  224  and the cylinder shape piston  250  are manufactured as one piece to form a single combined piston assembly  224 . The piston assembly  224  has a tube-like middle section  224   a . One end  224   b  of the middle tube-like section  224   a  is closed forming the head  250   a  of the cylinder shaped air piston  250 . At the other end  224   c  of the tube-like middle section  224   a  a flange-like section protrudes outwards from the outer surface of the tubular middle section  224   a  forming the annular shape piston head  225 . 
     There is a clearance, typically one to two millimeters, between the outside surface  264  of the moving cylindrical piston  250  and the inside surface  266  of the stationary cylindrical air chamber  240  in the engine head block  234 . This annular shape clearance space allows ambient air from the cylindrical air chamber  240  to enter the annular shape combustion chamber  242  during the air piston  250  out-motion and to let the compressed air in the annular shape combustion chamber  242  flow back to the cylindrical air chamber  240  during the air piston in-motion. This in-and-out airflow performs an efficient air-cooling function by transferring combustion heat from the walls of the annual shaped combustion chamber  242  to the air. Also it increases the thermal efficiency of the engine by transferring some of the combustion heat back to the combustion chamber with the supercharged air. 
     The inside face  251   a  of the head  250   a  of the cylinder-shaped part of the piston assembly is attached to a crankshaft  181  by a connecting rod  180  in a conventional way. The crankshaft transforms the reciprocating motion of the piston assembly into rotary motion. 
     At the closed end of the annular variable combustion chamber opposite from the engine head block end there are two or more fixed volume combustion chambers  242  and  200  in the engine block  232 . The function of these fixed volume combustion chambers was already described above in connection with the description of the first embodiment and will therefore not be repeated here. 
     The supercharged air intake valve(s)  129   a , the exhaust valve(s)  130   a , the fuel injector(s)  228   a  and  228   c , and/or spark plug(s)  206  are located at the side of the engine block  232  at the closed end of the annular combustion chamber  242  rather than on the top as was described earlier in connection with the first embodiment of the present invention. 
     The remainder of the earlier description of the first embodiment applies to this second embodiment as well. 
     Again, a four-stroke-cycle is the preferred form of this second embodiment of the present invention. The annular shape piston  225  in the combustion chamber  242  makes four strokes in a complete power cycle, two toward the head (closed head) of the combustion chamber  242  and two away from the head. However, the cylindrical piston in the cylindrical air chamber  240  in the engine head block  234  makes only two strokes in a complete supercharged air cycle sending a charge of supercharged air into the air accumulator twice during each power cycle. 
     The earlier description of the induction, compression, expansion and exhaust cycles of the first embodiment of the present invention apply also to this second embodiment of the present invention. 
     However, two major advantages are associated with this second embodiment over the first embodiment of the present invention: 
     a) The ambient air flow into the annular combustion chamber  242  from the cylindrical air chamber  240  during the compression and exhaust cycles of the engine performs an efficient air-cooling function increasing the thermal efficiency of the engine and allowing higher fuel charge per cubic inch of engine volume. 
     b) A flexible piston rod can be used to make the reciprocating masses lighter weight. Reference is made to  FIG. 16A  which is a cross section view showing a flexible piston rod  180   a  at the bottom center piston position and to  FIG. 16B  which is a cross section view showing a flexible piston rod  180   a  in the mid-expansion piston position. 
     By studying the two figures one can observe that the piston connecting rod  180   a  is always under tension during the expansion, exhaust and compression cycles allowing lighter rod construction and even the use of a flexible connecting rod. During the induction cycle the charge of supercharged air into the annular combustion chamber  240  balances some of the compression load on the connecting rod caused by the supercharge air pressure build-up in the cylindrical air chamber  240 . 
     Description of the Third Embodiment 
     Reference is made to  FIG. 17 , which is a cross section view showing the third embodiment of the apparatus of the present invention. 
     The internal combustion engine of the third embodiment of the present invention is similar to the second embodiment comprising an annular shape variable combustion chamber  242  in the engine block  232  in which a close fitting annular shape piston  225  slides. The open end  249  of the annular shape variable combustion chamber  242  is also covered with the engine head block  234  comprising a cylinder shape air chamber  240  that is closed at one end  241  and in which a loose fitting cylinder-shaped piston  250  slides. 
     However, at the open end  241   a  of the cylinder shape air chamber  240  in the engine head block  234  the inner face  266  of the air chamber fits airtight against the outer face  264  of the cylinder shaped piston  250 . This is accomplished with a set of 2 or more typical conventional piston rings  258 . 
     In the engine head block  234  facing the top  249  of the engine block  232  there is another set of two or more fixed volume combustion chambers similar to the fixed volume combustion chambers in the engine block  232  at the other closed end  248  of the annular shape variable combustion chamber  242 . The fixed volume combustion chambers in the engine head block will be later referred to as top with letter a, and the ones in the engine block as bottom fixed volume combustion chambers with letter b. Same top and bottom designation will be used in connection with respective valves, fuel injectors, spark plugs and ports. The function of these additional fixed volume combustion chambers is exactly the same as was already described above in connection with the description of the first and second embodiment and will therefore not be repeated here. 
     In this manner the fixed volume combustion chambers  294  and  200  are at both ends of the variable combustion chamber  242  thus making the close fitting annular shape piston  225  a double-acting piston. 
     At the top of the cylindrical air chamber  240  in the engine head block  234 , there are two or more valves, at least one for ambient air intake  194  and at least one for supercharged air discharge  196  to the supercharged air accumulator  190  as described earlier. 
     The function of the in-and-out motion of the cylinder-shaped piston  250  inside the cylindrical air chamber  240  has been described earlier. 
     The annular shape piston  225  and the cylinder-shaped piston  250  are manufactured as one piece to form a single combined piston assembly  224  as was described earlier in the description of the second embodiment. 
     Rest of the earlier component description of the first and second embodiment applies to this third embodiment as well. 
     Again, a four-cycle operation is the preferred form of this third embodiment of the present invention. However, while the annular shape piston  225  in the combustion chamber  242  makes four strokes to complete the four-cycle operation, it makes two expansion strokes due to the double-acting annular shape piston  225 . The third embodiment of the present invention becomes therefore a two-stroke four-cycle engine firing a power stroke once during every revolution of the crankshaft. 
     The earlier description of the induction, compression, expansion and exhaust cycles of the first and second embodiments of the present invention apply also to this third embodiment of the present invention. 
     However, three additional major advantages are associated with this third embodiment over the first and second embodiment of the present invention: 
     a) A two-stroke four-cycle engine firing a power stroke once during every revolution of the crankshaft produces twice the amount of power of a similar size four-stroke four-cycle engine. 
     b) This third embodiment of the present invention suits well for two-cycle operation making the engine a one-stroke two-cycle engine firing a power stroke twice during every revolution of the crankshaft (once every 180 degrees of the crankshaft revolution). 
     c) Operating two one-stroke two-cycle cylinders opposing each other makes the engine fire a power stroke four times during one revolution of the crankshaft: either twice every 180 degrees of the crankshaft revolution, if the two connecting rods are attached to the same crank pin or to two opposing crank pins, or once every 90 degrees of the crankshaft revolution, if the two connecting rods are attached to two crank pins that are 90 degrees apart. 
     A two cylinder engine will run as smoothly as a conventional eight cylinder engine. 
     Two-Stroke Four-Cycle Engine 
     Reference is made to  FIG. 18  which shows the cross section of two opposing cylinders of the third embodiment of the present invention. 
     Looking at  FIG. 18  in a landscape view one can observe that the left side piston assembly  224   a  and the right side piston assembly  224   b  are connected with their respective piston connecting rods  180   a  and  180   b  to crank pins  185   a  and  185   b  that are 180 degrees apart from each other in the crank shaft assembly  181 . Both pistons move at the same time either toward the crank case  183  or away from it.  FIG. 18  shows the pistons in the most inward position toward the crank case. 
     The left engine  20   a  supercharged air top intake port communicates with the supercharged air accumulator  190  through passage way  184   a , and the right engine  20   b  supercharged air top intake port communicates with the supercharged air accumulator  190  through passage way  184   b . Similarly the bottom intake ports of supercharged air are communicating with the supercharged air accumulator  190  through passage ways  184   c  and  184   d.    
     The left engine  20   a  top exhaust port communicates with the waste gas accumulator  30   a  through passage way  186   a , and the right engine  20   b  top exhaust port communicates with the waste gas accumulator  30   a  through passage way  186   b . Similarly the bottom exhaust ports are communicating with the waste gas accumulator  30   a  through passage ways  186   c  and  186   d.    
     The supercharged air discharge port  126   a  at the top of the left engine cylindrical air chamber  240   a  communicates with the compressed air accumulator  190  through a passage way  187   a  and the supercharged air discharge port  126   b  at the top of the right engine cylindrical air chamber  240   b  communicates with the compressed air accumulator  190  through a passage way  187   b . In both of these supercharged air discharge passage ways there are heat exchangers  190   c  which cool the supercharged air before it enters the supercharged air accumulator  190 . 
       FIG. 19A  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where the left engine  20   a  piston assembly  224   a  is in expansion/exhaust stroke moving away from the crank case  183  while the right engine  20   b  piston assembly  224   b  is in intake/compression stroke moving away from the crank case  183 . 
       FIG. 19B  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where the left engine  20   a  piston assembly  224   a  is in intake/exhaust stroke moving toward the crank case  183  while the right engine  20   b  piston assembly  224   b  is in expansion/compression stroke moving toward the crank case  183 . 
       FIG. 19C  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where the left engine  20   a  piston assembly  224   a  is in intake/compression stroke moving away from the crank case  183  while the right engine  20   b  piston assembly  224   b  is in expansion/exhaust stroke moving away from the crank case  183 . 
       FIG. 19D  is a cross section view showing two opposing 4-cycle cylinders of the third embodiment of the apparatus of the present invention, where the left engine  20   a  piston assembly  224   a  is in expansion/compression stroke moving toward the crank case  183  while the right engine  20   b  piston assembly  224   b  is in intake/exhaust stroke moving toward the crank case  183 . 
       FIG. 19E  shows the hatch patterns of the intake  188   a , compression  188   b , expansion  188   c  and exhaust  188   d  used in the figures from  19 A through  19 D. 
     One-Stroke Two-Cycle Engine 
       FIG. 20A  shows the expansion/compression phase of a one-stroke 2-cycle cylinder in the middle of the piston  224  up-stroke of the third embodiment of the apparatus of the present invention when top  294   a  and bottom  294   b  fixed volume pre-combustion chamber intake valves  129   a    129   b  and exhaust valves  130   a  and  130   b  in the supercharged combustion air supply chambers  200   a  and  200   b  are closed. The supercharged air discharge port  126  and the ambient air intake port  124  at the top of the cylindrical air chamber  240  are closed. 
       FIG. 20B  shows the exhaust/compression phase of a one-stroke 2-cycle cylinder of the third embodiment of the apparatus of the present invention during piston  224  up-stroke at about 80% of the expansion stroke permitting the escape of exhaust gases through the open bottom exhaust port  130   b  to the waste gas accumulator  30   a  and reducing the pressure in the cylinder. The top intake port  129   a , the top exhaust port  130   a  and the bottom intake port  129   b  are closed. The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is open to let the supercharged air flow into the supercharged air accumulator  190 . The ambient air intake port  124  at the top of the cylindrical air chamber  240  is closed. 
       FIG. 20C  shows the intake-scavenging/compression phase of a one-stroke 2-cycle cylinder of the third embodiment of the apparatus of the present invention during the piston  224  up-stroke at about 88% of the expansion stroke permitting the escape of exhaust gases through the bottom exhaust port  130   b  to the waste gas accumulator  30   a . The bottom intake port  129   b  is opened to let the incoming supercharged air from the supercharged air accumulator  190  scavenge the remaining waste gases away from the bottom pre-combustion chamber  294   b , the annular combustion chamber  242 , and the bottom supercharged air combustion chamber  200   b  into the waste gas accumulator  30   a . The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is open to let the supercharged air flow into the supercharged air accumulator  190 . The ambient air intake port  124  at the top of the cylindrical air chamber  240  is closed. 
       FIG. 20D  shows the expansion/compression phase of a one-stroke 2-cycle cylinder in the middle of the piston  224  down-stroke of the third embodiment of the apparatus of the present invention when top  294   a  and bottom  294   b  fixed volume combustion chamber intake valves  129   a  and  129   b , and exhaust valves  130   a  and  130   b  are closed. The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is closed but the ambient air intake port  124  is open. 
       FIG. 20E  shows the exhaust/compression phase of a one-stroke 2-cycle cylinder of the third embodiment of the apparatus of the present invention during piston  224  down-stroke at about 80% of the expansion stroke permitting the escape of exhaust gases through the open top exhaust port  130   a  to the waste gas accumulator  30   a  and reducing the pressure in the cylinder. The bottom intake port  129   b , the bottom exhaust port  130   b , and the top intake port  129   a  are closed. The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is closed. The ambient air intake port  124  at the top of the cylindrical air chamber  240  is open to let ambient air into the cylindrical air chamber  240 . 
       FIG. 20F  shows the intake-scavenging/compression phase of a one-stroke 2-cycle cylinder of the third embodiment of the apparatus of the present invention during the piston  224  down-stroke at about 88% of the expansion stroke permitting the escape of exhaust gases through the top exhaust port  130   a  to the waste gas accumulator  30   a . The top intake port  129   a  is opened to let the incoming supercharged air from the supercharged air accumulator  190  scavenge the remaining waste gases away from the top pre-combustion chamber  294   a , the annular combustion chamber  242 , and the top supercharged air combustion chamber  200   a  into the waste gas accumulator  30   a.    
     The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is closed. The ambient air intake port  124  at the top of the cylindrical air chamber  240  is open to let ambient air into the cylindrical air chamber. 
       FIG. 20G  shows the hatch patterns of the intake  188   a , compression  188   b , expansion  188   c  and exhaust  188   d  used in the figures from  20 A through  20 F. 
     DESCRIPTION OF THE FOURTH AND PREFERRED EMBODIMENT 
       FIG. 21A  is a cross section view showing two opposing cylinders  224   a  and  224   b  of the fourth and preferred embodiment of the apparatus of the present invention, where both piston rods  180   a  and  180   b  are connected to the same crankshaft pin  185 . 
       FIG. 21B  is a cross section view along line B-B of  FIG. 21A  showing two opposing cylinders  224   a  and  224   b  of the fourth and preferred embodiment of the apparatus of the present invention, where both piston rods  180   a  and  180   b  are connected to the same crankshaft pin  185 . 
     The internal combustion engine of the fourth and preferred embodiment of the present invention is similar to the third embodiment comprising an annular shape variable combustion chamber  242  in the engine block  232  in which a close fitting annular shape piston  225  slides. 
     Similarly, the engine head block  234  together with its separate head cover  235  comprises a cylinder shape air chamber  240  in which a cylinder-shaped piston  250  slides. 
     The function of the in-and-out motion of the cylinder-shaped piston inside the cylindrical air chamber  240  has been described earlier. 
     The annular shape piston  225  and the cylinder shape piston  250  are manufactured as one unit to form a single combined piston assembly  224  as was described earlier in the description of the second and third embodiment. 
     Similarly to the Third Embodiment of the Present Invention 
     There are fixed volume pre-combustion chambers  294   a  at the top and  294   b  at the bottom of the variable combustion chamber  242 . There are also fixed volume supercharged combustion air supply chambers  200   a  at the top and  200   b  at the bottom of the variable combustion chamber  242  thus making the close fitting annular shape piston  225  a double-acting piston. 
     However, there is an inner cooling chamber  261  similar to the outer cooling chamber  260  in the engine block  232  making the cooling chambers of this fourth embodiment different from the cooling chambers of the previous embodiments. 
     Cooling Chambers in the Engine Block 
     Reference is made to  FIG. 21A  and  FIG. 21B . There is an annular shape inner cooling chamber  261  in the engine block  232  between the annular shape combustion chamber  242  and the cylinder shape bore  236  in the middle of the engine block  232 . The open top end of this inner cooling chamber  261  is closed with an annular shape threaded or welded cover  271 . 
     There is another outer annular shape cooling chamber  260  outwards from the annular shape combustion chamber  242 . The open top end of this outer cooling chamber  260  is closed with an annular shape threaded or welded cover  270 . The outer annular cooling chamber  260  has a cooling media intake port  274  through one side of the engine block  232  and a cooling media discharge port  276  through the other side of the engine block  232 . To make the outer annular cooling chamber  260  to communicate with the inner annular cooling chamber  261   a  set of horizontal holes  263   a  and  263   b  are drilled through opposite sides of the engine block to reach the inner annular cooling chamber  261 . The outside ends of the holes  263   a  and  263   b  are capped with threaded or welded plugs  265 . 
     A set of vertical holes  267   a  and  267   b  are drilled through the bottom of the outer annular shape cooling chamber  260  to reach and communicate with the horizontal holes  263   a  and  263   b  that communicate with the inner annular cooling chamber  261 . 
     To control the flow of the cooling media reasonably evenly through both annular cooling chambers a set of vertical weir pins  269  are used as shown in  FIG. 21C  and  FIG. 22 . 
       FIG. 21C  is a cross section view along line C-C of  FIG. 21B  showing the location of the cooling liquid flow weir pins in the outer  260  and inner  261  cooling chambers of the fourth and preferred embodiment of the apparatus of the present invention. 
       FIG. 22  shows the cooling liquid flow pattern over the weir pins in the cooling chambers of the fourth and preferred embodiment of the apparatus of the present invention. 
     Copies of the cross sections of the top cylinder engine block  232  assembly from  FIG. 21A  and from  FIG. 21B  together with the copy of  FIG. 21C  are shown in  FIG. 22 . The middle of  FIG. 22  shows the outer annular cooling chamber  260  straightened out in two halves  260   a  and  260   b  as if it were a straight rather than an annular chamber, the thickness of the straightened chamber being same as the distance between the inner cylindrical face  260   c  and outer cylindrical face  260   d  of the outer annular cooling chamber  260 . 
     Between these imaginary two straightened halves  260   a  and  260   b  of the outer annular cooling chamber  260  is shown an imaginary straightened inner cooling chamber  261   ab  of the actual annular shape inner cooling camber  260 . 
     Lines  274   a  and  274   b  point to the intake port  274  where cooling media (typically water or air) enters the outer cooling chamber  260 . 
     Lines  365   a  and  366   a  point to blocking weir pin  269   a , and lines  365   b  and  366   b  point to blocking weir pin  269   b , which are separating the two halves  260   a  and  260   b  of the outer annular cooling chamber  260  from each other. 
     Line  365   c  points to over-flow weir pin  269   c  and line  365   d  points to over-flow weir pin  269   d , which divide each half of the outer cooling cambers into two quarter sections. For later description of the flow pattern of the cooling media through both annular cooling cambers the outer annular shape cooling chamber  260  quarter sections are called first quarter  361 , second quarter  362 , third quarter  363 , and fourth quarter  364 . The over-flow weir pins are shorter than the height of the outer annular cooling chamber allowing the cooling media to flow over the weir from one quarter to the other. 
     Lines  365   e  and  365   f  point to two additional over-flow weir pins  269   e  and  269   f  which divide the inner cooling camber  261  into two half sections  261   a  and  261   b . For later description of the flow pattern of the cooling media through both annular cooling cambers the inner cooling chamber half sections are called first half  261   a  and second half  261   b . These two over-flow weir pins  269   e  and  269   f  are also shorter than the height of the inner annular cooling chamber allowing the cooling media to flow over the weirs from one half to the other. 
     Lines  276   a  and  276   b  point to the discharge port  276  where cooling media (typically water or air) leaves the outer cooling chamber  260 . 
     The second quarter  362  of the first half  260   a  of the outer cooling chamber communicates with the first half  261   a  of the inner cooling chamber through the bottom passage way  366  which is a set of horizontal holes  263  as described earlier. 
     The second half  261   b  of the inner cooling chamber communicates with the third quarter  363  of the second half  260   b  of the outer cooling chamber through the passage way  367  which is a set of horizontal holes  263  as described earlier. 
     The cooling media enters the first quarter  361  of the outer annular cooling chamber  260  through intake port  274 , passes over the over-flow weir pin  269   c  into the second quarter  362  of the outer annular cooling chamber  260 . Through the bottom passage way  366  the cooling media flows from the second quarter  362  of the outer annular cooling chamber  260  to the bottom middle of the first half  261   a  of the inner annular cooling chamber  261 . The cooling media flow splits into two flows over both of the over-flow weir pins  269   e  and  269   f  in the inner cooling camber  261  and enters the top of the second half  261   b  of the inner cooling chamber. Through the bottom passage way  367  the cooling media flows from the second half  261   b  of the inner cooling chamber  261  into the third quarter  363  of the outer annular cooling chamber  260 . 
     From the third quarter  363  of the outer annular cooling chamber  260  the cooling media passes over the over-flow weir pin  269   d  into the fourth quarter  364  of the outer annular cooling chamber  360 , and finally exits from there through the discharge port  276 . In this manner the cooling media is forced to flow up and down as well as sideways through both of the annular cooling chambers ensuring efficient cooling of all surfaces to deliver the excess combustion heat away from the engine block. 
     Fixed Volume Pre-Combustion Chambers and Fixed Volume Supercharged Combustion Air Supply Chambers 
     The shape and location of the top fixed volume combustion chambers  294   a  and  200   a  in the engine head block  234  and the bottom fixed volume combustion chambers  294   b  and  200   b  in the engine block  232  at both ends of the annular combustion chamber  242  make the fourth embodiment different from the third embodiment of the present invention. 
     The shape and location of the top fixed volume pre-combustion chamber  294   a  and of the top fixed volume supercharged combustion air supply chamber  200   a  in the engine head block  234  is shown in  FIG. 21D , which is a cross section view along line A-A of  FIG. 21A . 
     The bottom fixed volume pre-combustion chamber  294   b  and the bottom fixed volume supercharged combustion air supply chamber  200   b  in the engine block  232  are of same shape as the respective top fixed volume chambers in the engine head block  234 . 
     The air or fuel mixture intake valves  229  and the waste gas exhaust valves  230 , fuel injectors  228  (and/or spark plugs) are mounted in separate housings  229   b  attached to the sides of the engine block  232 . Only the valve heads protrude into the fixed volume combustion chambers while in open position. From the valve head recesses  229   c  in each of the fixed volume combustion chambers typically two bored passage ways  295   a  and  295   b  penetrate through the engine block  232  and engine head block  234  into both top and bottom end of the annular combustion chamber  242 . The passage ways enter the annular combustion chamber preferably tangentially to create maximum flame front turbulence in the annular combustion chamber  242  during the expansion cycle. The valve head recesses together with the passage ways form the fixed volume chambers. By directing the fuel injector  228  to spray directly into passage way  295   a  as shown in  FIG. 21D  passage way  295   b  also becomes a supercharged combustion air supply chamber. 
     Rest of the earlier description of the third embodiment applies to this fourth embodiment as well. 
     Again, a four-cycle operation is the preferred form of this fourth embodiment of the present invention making the engine a two-stroke four-cycle engine firing a power stroke once during every revolution of the crankshaft with each cylinder. 
     This fourth embodiment of the present invention suits also well for two-cycle operation making the engine a one-stroke two-cycle engine firing a power stroke twice during every revolution of the crankshaft with each cylinder. 
     The following three figures will demonstrate the sequences of the strokes in a two cylinder two-cycle engine of this fourth and preferred embodiment. 
       FIG. 23A  shows the expansion/compression phase of two opposing 2-cycle cylinders. Both piston rods  180   a  and  180   b  are connected to the same crankshaft pin  185  which results in both cylinders firing a power stroke at the same time every 180 degree of crank shaft revolution. The engine power out-put is typical to a conventional 8-cylinder 4-cycle engine. 
       FIG. 23B  shows the exhaust/compression phase of two opposing 2-cycle cylinders. Both piston rods  180   a  and  180   b  are connected to the same crankshaft pin  185 . In this manner the crank case pressure will not vary since the cylindrical air pistons always move together, one moving inwards toward the crank case while the other moves outwards maintaining the same volume in the crank case. The cylindrical air pistons perform a full supercharged air compression cycle once during every revolution of the crankshaft. 
       FIG. 23C  shows the intake-scavenging/compression phase of two opposing 2-cycle cylinders. Both piston rods  180   a  and  180   b  are connected to the same crankshaft pin  185 . By connecting the piston rods to two separate crankshaft pins 90 degrees apart one 2-cylinder engine will perform one power stroke every 90 degrees of crank shaft revolution. 
       FIG. 23D  shows the hatch patterns of the intake  188   a , compression  188   b , expansion  188   c , and exhaust  188   d  used in the figures from  23 A through  23 C. 
     Supercharged Air Supply 
     The middle cylindrical supercharged air piston can be sized to produce all the necessary supercharged air volume for the engine or an exhaust-gas driven compressor can be used to reduce the size of the this cylindrical piston. This latter option gives a 5-10% better fuel economy but will add to the cost of the engine. 
     DESCRIPTION OF OTHER FEATURES OF PRESENT INVENTION 
     By aligning the centerlines of two opposing cylinders and connecting each piston pair with a straight common connecting rod, the cylindrical air chambers will work as an air compressor without having to convert the reciprocating motion of the combustion engine to a rotary motion through a crank shaft (see  FIG. 24A  through  FIG. 24E ). 
       FIG. 24A  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod  180  to function as a compressor, where piston  224   a  is in intake/exhaust stroke while piston  224   b  is in expansion/compression stroke. 
       FIG. 24B  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod  180  to function as a compressor, where piston  224   a  is in intake/compression stroke while piston  224   b  is in expansion/exhaust stroke. 
       FIG. 24C  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod  180  to function as a compressor, where piston  224   a  is in expansion/compression stroke while piston  224   b  is in intake/exhaust stroke. 
       FIG. 24D  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention connected with a common piston rod  180  to function as a compressor, where piston  224   a  is in expansion/exhaust stroke while piston  224   b  is in intake/compression stroke. 
       FIG. 24E  shows the hatch patterns of the intake  188   a , compression  188   b , expansion  188   c  and exhaust  188   d  used in the figures from  24 A through  24 D. 
     By aligning the centerlines of two opposing cylinders and using a combined cylindrical air piston assembly for both cylinders, the cylindrical air chambers will work as an air compressor without having to convert the reciprocating motion of the combustion engine to a rotary motion through a crank shaft. Three different piston assemblies are shown in the next four figures. 
       FIG. 25A  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air pistons and ring shaped combustion pistons are formed as one unit to function as a compressor. Piston  224   a  is in intake/exhaust stroke while piston  224   b  is in expansion/compression stroke. 
       FIG. 25B  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air pistons and ring shaped combustion pistons are formed as one unit to function as a compressor. Piston  224   a  is in intake/compression stroke while piston  224   b  is in expansion/exhaust stroke. 
       FIG. 25C  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air piston heads  251   a  and  251   b  are aligned with the ring shaped combustion pistons  225   a  and  225   b  and formed as one unit to function as a compressor. Cylinder  224   a  is in expansion/compression stroke while cylinder  224   b  is in intake/exhaust stroke. 
       FIG. 25D  is a cross section view showing two opposing 4-cycle cylinders of the fourth and preferred embodiment of the apparatus of the present invention in which the cylindrical shape air pistons have one common head  251  in the middle of the piston assembly  224  and form one unit with the ring shaped combustion pistons to function as a compressor. Cylinder  224   a  is in expansion/exhaust stroke while cylinder  224   b  is in intake/compression stroke. 
       FIG. 25E  shows the hatch patterns of the intake  188   a , compression  188   b , expansion  188   c , and exhaust  188   d  used in the figures from  25 A through  25 D. 
     Further, it is to be recognized that the above possible modifications are given by way of example, and yet other possible modifications could be made without departing from the basic teachings of the present invention. 
     One-Stroke Two-Cycle Engine 
     The following seven figures will demonstrate the sequences of the strokes in a single cylinder one-stroke two-cycle engine in one embodiment. 
       FIG. 26A  shows the expansion/compression phase of a single 2-cycle cylinder. A one cylinder engine will perform one power stroke every 180 degrees of crank shaft revolution and its power out-put and torque is equal or greater than a conventional 8-cylinder 4-cycle engine&#39;s with the same cylinder displacement volume per cylinder (see Air-standard Analysis and  FIG. 14  below). 
     Looking at  FIGS. 26A through 26C  and  26 E through  26 G it can be seen how the annular combustion chamber  242  comprises a void  242 A above the annular shape bottom end  225  of the piston assembly  224 . This allows for two power strokes for every rotation of the crankshaft  181 . 
       FIG. 26B  shows the exhaust/compression phase of a single 2-cycle cylinder. The cylindrical air piston performs a full supercharged air compression cycle once during every revolution of the crankshaft. 
       FIG. 26C  shows the intake-scavenging/compression phase of a single 2-cycle cylinder. By using 2 cylinders and connecting the piston rods to two separate crankshaft pins 90 degrees apart one 2-cylinder engine will perform one power stroke every 90 degrees of crank shaft revolution and its power out-put and torque is equal or greater than a conventional 16-cylinder 4-cycle engine&#39;s with the same cylinder displacement volume per cylinder (see Air-standard Analysis and  FIG. 14  below). 
       FIG. 26D  shows the hatch patterns of the intake  188   a , compression  188   b , expansion  188   c , and exhaust  188   d  used in the figures from  FIG. 26A  through  FIG. 26G . 
       FIG. 26E  through  FIG. 26G  show the expansion/compression phase, the exhaust/compression phase, and the intake-scavenging/compression phase of the same single 2-cycle cylinder during the piston down-stroke. 
     A more detailed description of this one-stroke 2-cycle cylinder engine with an annular double-acting power piston is given below. 
       FIG. 26A  shows the expansion/compression phase of a one-stroke 2-cycle cylinder in the middle of the piston  224  up-stroke when top  294   a  and bottom  294   b  fixed volume pre-combustion chamber intake valves  129   a  and  129   b  and exhaust valves  130   a  and  130   b  in the supercharged combustion air supply chambers  200   a  and  200   b  are closed. The supercharged air discharge port  126  and the ambient air intake port  124  at the top of the cylindrical air chamber  240  are closed. 
       FIG. 26B  shows the exhaust/compression phase of a one-stroke 2-cycle cylinder during piston  224  up-stroke at about 80% of the expansion stroke permitting the escape of exhaust gases through the open bottom exhaust port  130   b  to the waste gas accumulator  30   a  and reducing the pressure in the cylinder. The top intake port  129   a , the top exhaust port  130   a  and the bottom intake port  129   b  are closed. The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is open to let the supercharged airflow into the supercharged air accumulator  190 . The ambient air intake port  124  at the top of the cylindrical air chamber  240  is closed. 
       FIG. 26C  shows the intake-scavenging/compression phase of a one-stroke 2-cycle cylinder during the piston  224  up-stroke at about 88% of the expansion stroke permitting the escape of exhaust gases through the bottom exhaust port  130   b  to the waste gas accumulator  30   a . The bottom intake port  129   b  is opened to let the incoming supercharged air from the supercharged air accumulator  190  scavenge the remaining waste gases away from the bottom pre-combustion chamber  294   b , the annular combustion chamber  242 , and the bottom supercharged combustion air supply chamber  200   b  into the waste gas accumulator  30   a . The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is open to let the supercharged airflow into the supercharged air accumulator  190 . The ambient air intake port  124  at the top of the cylindrical air chamber  240  is closed. 
       FIG. 26D  shows the hatch patterns of the intake  188   a , compression  188   b , expansion  188   c  and exhaust  188   d  used in the figures from  26 A through  26 F. 
       FIG. 26E  shows the expansion/compression phase of a one-stroke 2-cycle cylinder in the middle of the piston  224  down-stroke when top  294   a  and bottom  294   b  fixed volume combustion chamber intake valves  129   a  and  129   b , and exhaust valves  130   a  and  130   b  are closed. The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is closed but the ambient air intake port  124  is open. 
       FIG. 26F  shows the exhaust/compression phase of a one-stroke 2-cycle cylinder during piston  224  down-stroke at about 80% of the expansion stroke permitting the escape of exhaust gases through the open top exhaust port  130   a  to the waste gas accumulator  30   a  and reducing the pressure in the cylinder. The bottom intake port  129   b , the bottom exhaust port  130   b , and the top intake port  129   a  are closed. The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is closed. The ambient air intake port  124  at the top of the cylindrical air chamber  240  is open to let ambient air into the cylindrical air chamber  240 . 
       FIG. 26G  shows the intake-scavenging/compression phase of a one-stroke 2-cycle cylinder during the piston  224  down-stroke at about 88% of the expansion stroke permitting the escape of exhaust gases through the top exhaust port  130   a  to the waste gas accumulator  30   a . The top intake port  129   a  is opened to let the incoming supercharged air from the supercharged air accumulator  190  scavenge the remaining waste gases away from the top pre-combustion chamber  294   a , the annular combustion chamber  242 , and the top supercharged combustion air supply chamber  200   a  into the waste gas accumulator  30   a.    
     The supercharged air discharge port  126  at the top of the cylindrical air chamber  240  is closed. The ambient air intake port  124  at the top of the cylindrical air chamber  240  is open to let ambient air into the cylindrical air chamber. 
     Now referring to  FIG. 27 , the Piston  320  oscillates back and forth within the chambers  322  and  324 . Gas compressed air rather traveling through the check valve  326  will enter within the chamber  324  as the piston  320  moves to the left in  FIG. 27 . As shown in  FIG. 28 , the air is drawn into chamber  324  through the check valve  326  from the air supply sump  330 . The air inside of the chamber  322  is now super compressed and a fuel injector (not shown) will inject fuel at this juncture and combusting an igniting device, such as a sparkplug, will ignite the mixture and the piston will travel now to the right. The position as shown in  FIG. 28 , the compressed air from the sump  330  will purge the gas in  324  and will exit out the exit port  332  where it can be seen that the rightward perimeter portion  334  of the piston  320  allows communication to the port  332  for this purging and clean air. As the pistol moves to the right this port seals off and the air begins to super compress, as shown in  FIG. 27 . 
     In left hand portion of  FIG. 27  it can be seen where the second piston  320 ′ is in an immediate position moving upwardly and the gas is within and chamber  322 ′ is combusted and expanding and applying power or a force upon the piston and the air in  324 ′ is now being super compressed to be ignited thereafter. 
     Now referring to  FIG. 29  it can be seen that another piston can be utilized for compressing the sump  330 ′. A super cooler  340  can be utilized to cool the air. It should also be noted that this mix can be retrofitted to an existing engine block. The legion in  FIG. 27  shows fresh air at  350 , compressed  352 , super compressed at  354 , combusted gas at  356  and exhausted at  358 . 
     While the present invention is illustrated by description of several embodiments and while the illustrative embodiments are described in detail, it is not the intention of the applicants to restrict or in any way limit the scope of the appended claims to such detail. Additional advantages and modifications within the scope of the appended claims will readily appear to those sufficed in the art. The invention in its broader aspects is therefore not limited to the specific details, representative apparatus and methods, and illustrative examples shown and described. Accordingly, departures may be made from such details without departing from the spirit or scope of applicants&#39; general concept.