Abstract:
An engine includes a crankshaft, rotating about a crankshaft axis of the engine. A power piston is slidably received within a first cylinder and is operatively connected to the crankshaft via first linkage system. The power piston reciprocates through a power stroke and an exhaust stroke of a four stroke cycle during a single rotation of the crankshaft. A compression piston is slidably received within a second cylinder and operatively connected to the crankshaft via second linkage system. The first and second linkage systems share no common mechanical link. The compression piston reciprocates through an intake stroke and a compression stroke of the same four stroke cycle during the same revolution of the crankshaft. The power piston leads the compression piston by a phase shift angle that is substantially equal to or greater than zero degrees and less than 30 degrees.

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This patent application is a continuation application of U.S. application Ser. No. 10/139,981, filed May 7, 2002 now U.S. Pat. No. 6,609,371, entitled “SPLIT FOUR STROKE ENGINE”, herein incorporated by reference in its entirety, which is a continuation application of U.S. application Ser. No. 09/909,594 (now U.S. Pat. No. 6,543,225), filed Jul. 20, 2001, entitled “SPLIT FOUR STROKE CYCLE INTERNAL COMBUSTION ENGINE”, herein incorporated by reference in its entirety. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to internal combustion engines. More specifically, the present invention relates to a four-stroke cycle internal combustion engine having a pair of offset pistons in which one piston of the pair is used for the intake and compression strokes and another piston of the pair is used for the power and exhaust strokes, with each four stroke cycle being completed in one revolution of the crankshaft. 
     BACKGROUND OF THE INVENTION 
     Internal combustion engines are any of a group of devices in which the reactants of combustion, e.g., oxidizer and fuel, and the products of combustion serve as the working fluids of the engine. The basic components of an internal combustion engine are well known in the art and include the engine block, cylinder head, cylinders, pistons, valves, crankshaft and camshaft. The cylinder heads, cylinders and tops of the pistons typically form combustion chambers into which fuel and oxidizer (e.g., air) is introduced and combustion takes place. Such an engine gains its energy from the heat released during the combustion of the non-reacted working fluids, e.g., the oxidizer-fuel mixture. This process occurs within the engine and is part of the thermodynamic cycle of the device. In all internal combustion engines, useful work is generated from the hot, gaseous products of combustion acting directly on moving surfaces of the engine, such as the top or crown of a piston. Generally, reciprocating motion of the pistons is transferred to rotary motion of a crankshaft via connecting rods. 
     Internal combustion (IC) engines can be categorized into spark ignition (SI) and compression ignition (CI) categories. SI engines, i.e. typical gasoline engines, use a spark to ignite the air-fuel mixture, while the heat of compression ignites the air fuel mixture in CI engines, i.e., typically diesel engines. 
     The most common internal-combustion engine is the four-stroke cycle engine, a conception whose basic design has not changed for more than 100 years old. This is because of its outstanding performance as a prime mover in the ground transportation industry. In a four-stroke cycle engine, power is recovered from the combustion process in four separate piston movements (strokes) of a single piston. For purposes herein, a stroke is defined as a complete movement of a piston from a top dead center position to a bottom dead center position or vice versa. Accordingly, a four-stroke cycle engine is defined herein to be an engine which requires four complete strokes of one or more pistons for every power stroke, i.e. for every stroke that delivers power to a crankshaft. 
     Referring to  FIGS. 1-4 , an exemplary embodiment of a prior art four stroke cycle internal combustion engine is shown at  10 . For purposes of comparison, the following four  FIGS. 1-4  describe what will be termed a prior art “standard engine”  10 . As will be explained in greater detail hereinafter, this standard engine  10  is an SI engine with a 4 inch diameter piston, a 4 inch stroke and an 8 to 1 compression ratio. The compression ratio is defined herein as the maximum volume of a predetermined mass of an air-fuel mixture before a compression stroke, divided by the volume of the mass of the air-fuel mixture at the point of ignition. For the standard engine, the compression ratio is substantially the ratio of the volume in cylinder  14  when piston  16  is at bottom dead center to the volume in the cylinder  14  when the piston  16  is at top dead center. 
     The engine  10  includes an engine block  12  having the cylinder  14  extending therethrough. The cylinder  14  is sized to receive the reciprocating piston  16  therein. Attached to the top of the cylinder  14  is the cylinder head  18 , which includes an inlet valve  20  and an outlet valve  22 . The cylinder head  18  cylinder  14  and top (or crown  24 ) of the piston  16  form a combustion chamber  26 . On the inlet stroke (FIG.  1 ), a fuel air mixture is introduced into the combustion chamber  26  through an intake passage  28  and the inlet valve  20 , wherein the mixture is ignited via spark plug  30 . The products of combustion are later exhausted through outlet valve  22  and outlet passage  32  on the exhaust stroke (FIG.  4 ). A connecting rod  34  is pivotally attached at its top distal end  36  to the piston  16 . A crankshaft  38  includes a mechanical offset portion called the crankshaft throw  40 , which is pivotally attached to the bottom distal end  42  of connecting rod  34 . The mechanical linkage of the connecting rod  34  to the piston  16  and crankshaft throw  40  serves to convert the reciprocating motion (as indicated by arrow  44 ) of the piston  16  to the rotary motion (as indicated by arrow  46 ) of the crankshaft  38 . The crankshaft  38  is mechanically linked (not shown) to an inlet camshaft  48  and an outlet camshaft  50 , which precisely control the opening and closing of the inlet valve  20  and outlet valve  22  respectively. 
     The cylinder  14  has a centerline (piston-cylinder axis)  52 , which is also the centerline of reciprocation of the piston  16 . The crankshaft  38  has a center of rotation (crankshaft axis)  54 . For purposes of this specification, the direction of rotation  46  of the crankshaft  38  will be in the clockwise direction as viewed by the reader into the plane of the paper. The centerline  52  of the cylinder  14  passes directly through the center of rotation  54  of the crankshaft  38 . 
     Referring to  FIG. 1 , with the inlet valve  20  open, the piston  16  first descends (as indicated by the direction of arrow  44 ) on the intake stroke. A predetermined mass of an explosive mixture of fuel (gasoline vapor) and air is drawn into the combustion chamber  26  by the partial vacuum thus created. The piston continues to descend until it reaches its bottom dead center (BDC), the point at which the piston is farthest from the cylinder head  18 . 
     Referring to  FIG. 2 , with both the inlet  20  and outlet  22  valves closed, the mixture is compressed as the piston  16  ascends (as indicated by the direction of arrow  44 ) on the compression stroke. As the end of the stroke approaches top dead center (TDC), i.e., the point at which the piston  16  is closest to the cylinder head  18 , the volume of the mixture is compressed to one eighth of its initial volume (due to an 8 to 1 compression ratio). The mixture is then ignited by an electric spark from spark plug  30 . 
     Referring to  FIG. 3 , the power stroke follows with both valves  20  and  22  still closed. The piston  16  is driven downward (as indicated by arrow  44 ) toward bottom dead center (BDC), due to the expansion of the burned gas pressing on the crown  24  of the piston  16 . Since the spark plug  30  is fired when the piston  16  is at or near TDC, i.e. at its firing position, the combustion pressure (indicated by arrow  56 ) exerted by the ignited gas on the piston  16  is at its maximum at this point. This pressure  56  is transmitted through the connecting rod  34  and results in a tangential force or torque (as indicated by arrow  58 ) on the crankshaft  38 . 
     When the piston  16  is at ifs firing position, there is a significant clearance distance  60  between the top of the cylinder  14  and the crown  24  of the piston  16 . Typically, the clearance distance is between 0.5 to 0.6 inches. For the standard engine  10  illustrated the clearance distance is substantially 0.571 inches. When the piston  16  is at its firing position conditions are optimal for ignition, i.e., optimal firing conditions. For purposes of comparison, the firing conditions of this engine  10  exemplary embodiment are: 1) a 4 inch diameter piston, 2) a clearance volume of 7.181 cubic inches, 3) a pressure before ignition of approximately 270 pounds per square inch absolute (psia), 4) a maximum combustion pressure after ignition of approximately 1200 psia and 5) operating at 1400 RPM. 
     This clearance distance  60  corresponds typically to the 8 to 1 compression ratio. Typically, SI engines operate optimally with a fixed compression ratio within a range of about 6.0 to 8.5, while the compression ratios of CI engines typically range from about 10 to 16: The piston&#39;s 16 firing position is generally at or near TDC, and represents the optimum volume and pressure for the fuel-air mixture to ignite. If the clearance distance  60  were made smaller, the pressure would increase rapidly. 
     Referring to  FIG. 4 , during the exhaust stroke, the ascending piston  16  forces the spent products of combustion through the open outlet (or exhaust) valve  22 . The cycle then repeats itself. For this prior art four stoke cycle engine  10 , four stokes of each piston  16 , i.e. inlet, compression, power and exhaust, and two revolutions of the crankshaft  38  are required to complete a cycle, i.e. to provide one power stroke. Problematically, the overall thermodynamic efficiency of the standard four stroke engine  10  is only about one third (⅓). That is ⅓ of the work is delivered to the crankshaft, ⅓ is lost in waste heat, and ⅓ is lost out of the exhaust. 
     As illustrated in  FIGS. 3 and 5 , one of the primary reasons for this low  20  efficiency is the fact that peak torque and peak combustion pressure are inherently locked out of phase.  FIG. 3  shows the position of the piston  16  at the beginning of a power stroke, when the piston  16  is in its firing position at or near TDC. When the spark plug  30  fires, the ignited fuel exerts maximum combustion pressure  56  on the piston  16 , which is transmitted through the connecting rod  34  to the crankshaft throw  40  of crankshaft  38 . However, in this position, the connecting rod  34  and the crankshaft throw  40  are both nearly aligned with the centerline  52  of the cylinder  14 . Therefore, the torque  58  is almost perpendicular to the direction of force  56 , and is at its minimum value. The crankshaft  38  must rely on momentum generated from an attached flywheel (not shown) to rotate it past this position. 
     Referring to  FIG. 5 , as the ignited gas expands in the combustion chamber  26 , the piston  16  descends and the combustion pressure  56  decreases. However, as the crankshaft throw  40  rotates past the centerline  52  and TDC, the resulting tangential force or torque  58  begins to grow. The torque  58  reaches a maximum value when the crankshaft throw  40  rotates approximately 30 degrees past the centerline  52 . Rotation beyond that point causes the pressure  56  to fall off so much that the torque  58  begins to decrease again, until both pressure  56  and torque  58  reach a minimum at BDC. Therefore, the point of maximum torque  58  and the point of maximum combustion pressure  56  are inherently locked out of phase by approximately 30 degrees. 
     Referring to  FIG. 6 , this concept can be further illustrated. Here, a graph of tangential force or torque versus degrees of rotation from TDC to BDC is shown at  62  for the standard prior art engine  10 . Additionally, a graph of combustion pressure versus degrees of rotation from TDC to BDC is shown at  64  for engine  10 . The calculations for the graphs  62  and  64  were based on the standard prior art engine  10  having a four inch stroke, a four inch diameter piston, and a maximum combustion pressure at ignition of about 1200 PSIA. As can be seen from the graphs, the point of maximum combustion pressure  66  occurs at approximately 0 degrees from TDC and the point of maximum torque  68  occurs approximately 30 degrees later when the pressure  64  has been reduced considerably. Both graphs  62  and  64  approach their minimum values at BDC, or substantially 180 degrees of rotation past TDC. 
     An alternative way of increasing the thermal dynamic efficiency of a four stoke cycle engine is to increase the compression ratio of the engine. However, automotive manufactures have found that SI engines typically operate optimally with a compression ratio within a range of about 6.0 to 8.5, while CI engines typically operate best within a compression ratio range of about 10 to 16. This is because as the compression ratios of SI or CI engines increase substantially beyond the above ranges, several other problems occur which outweigh the advantages gained. For example, the engine must be made heavier and bulkier in order to handle the greater pressures involved. Also problems of premature ignition begin to occur, especially in SI engines. 
     Many rather exotic early engine designs were patented. However, none were able to offer greater efficiencies or other significant advantages, which would replace the standard engine  10  exemplified above. Some of these early patents included: U.S. Pat. Nos. 848,029; 939,376; 1,111,841; 1,248,250; 1,301,141; 1,392,359; 1,856,048; 1,969,815; 2,091,410; 2,091,411; 2,091,412; 2,091,413; 2,269,948; 3,895,614; British Patent No. 299,602; British Patent No. 721,025 and Italian Patent No. 505,576. 
     In particular the U.S. Pat. No. 1,111,841 to Koenig disclosed a prior art split piston/cylinder design in which an intake and compression stroke was accomplished in a compression piston  12 /cylinder  11  combination, and a power and an exhaust stroke was accomplished in an engine piston  7 /cylinder  8  combination. Each piston  7  and  12  reciprocates along a piston cylinder axis which intersected the single crankshaft  5  (see  FIG. 3  therein). A thermal chamber  24  connects the heads of the compression and engine cylinders, with one end being open to the engine cylinder and the other end having a valued discharge port  19  communicating with the compressor cylinder. A water cooled heat exchanger  15  is disposed at the top of the compressor cylinder  11  to cool the air or air/fuel mixture as it is compressed. A set of spaced thermal plates  25  are disposed within the thermal chamber  24  to re-heat the previously cooled compressed gas as it passes through. 
     It was thought that the engine would gain efficiency by making it easier to compress the gas by cooling it. Thereafter, the gas was re-heated in the thermal chamber in order to increase its pressure to a point where efficient ignition could take place. Upon the exhaust stroke, hot exhaust gases were passed back through the thermal chamber and out of an exhaust port  26  in an effort to re-heat the thermal chamber. 
     Unfortunately, transfer of gas in all prior art engines of a split piston design always requires work, which reduces efficiency. Additionally, the added expansion from the thermal chamber to the engine cylinder of Koenig also reduced compression ratio. The standard engine  10  requires no such transfer process and associated additional work. Moreover, the cooling and re-heating of the gas, back and forth through the thermal chamber did not provide enough of an advantage to overcome the losses incurred during the gas transfer process. Therefore, the Koenig patent lost efficiency and compression ratio relative to the standard engine  10 . 
     For purposes herein, a crankshaft axis is defined as being offset from the piston cylinder axis when the crankshaft axis and the piston-cylinder axis do not intersect. The distance between the extended crankshaft axis and the extended piston-cylinder axis taken along a line drawn perpendicular to the piston cylinder axis is defined as the offset. Typically, offset pistons are connected to the crankshaft by well-known connecting rods and crankshaft throws. However, one skilled in the art would recognize that offset pistons may be operatively connected to a crankshaft by several other mechanical linkages. For example, a first piston may be connected to a first crankshaft and a second piston may be connected to a second crankshaft, and the two crankshafts may be operatively connected together through a system of gears. Alternatively, pivoted lever arms or other mechanical linkages may be used in conjunction with, or in lieu of, the connecting rods and crankshaft throws to operatively connect the offset pistons to the crankshaft. 
     Certain technology relating to reciprocating piston internal combustion engines in which the crankshaft axis is offset from, i.e., does not intersect with, the piston-cylinder axes is described in U.S. Pat. Nos. 810,347; 2,957,455; 2,974,541; 4,628,876; 4,945,866; and 5,146,884; in Japan patent document 60-256,642; in Soviet Union patent document 1551-880-A; and in Japanese Society of Automotive Engineers (JSAE) Convention Proceedings, date 1996, issue 966, pages 129-132. According to descriptions contained in those publications, the various engine geometries are motivated by various considerations, including power and torque improvements and friction and vibration reductions. Additionally, in-line, or straight engines in which the crankshaft axis is offset from the piston axes were used in early twentieth century racing engines. 
     However, all of the improvements gained were due to increasing the torque angles on the power stroke only. Unfortunately, as will be discussed in greater detail hereinafter, the greater the advantage an offset was to the power stroke was also accompanied by an associated increasing disadvantage to the compression stroke. Therefore, the degree of offset quickly becomes self limiting, wherein the advantages to torque, power, friction and vibration to the power stroke do not out weigh the disadvantages to the same functions on the compression stroke. Additionally, no advantages were taught or discussed regarding offsets to optimize the compression stroke. 
     By way of example, a recent prior art attempt to increase efficiency in a standard engine  10  type design through the use of an offset is disclosed in U.S. Pat. No. 6,058,901 to Lee. Lee believes that improved efficiency will result by reducing the frictional forces of the piston rings on the side walls over the full duration of two revolutions of a four stroke cycle (see Lee, column 4, lines 1016). Lee attempts to accomplish this by providing an offset cylinder, wherein the timing of combustion within each cylinder is controlled to cause maximum combustion pressure to occur when an imaginary plane that contains both a respective connection axis of a respective connecting rod to the respective piston and a respective connection axis of the connecting rod to a respective throw of the crankshaft is substantially coincident with the respective cylinder axis along which the piston reciprocates. 
     However, though the offset is an advantage during the power stroke, it becomes a disadvantage during the compression stroke. That is, when the piston travels from bottom dead center to top dead center during the compression stroke, the offset piston-cylinder axis creates an angle between the crankshaft throw and connecting rod that reduces the torque applied to the piston. Additionally, the side forces resulting from the poor torque angles on the compression stroke actually increase wear on the piston rings. Accordingly, a greater amount of power must be consumed in order to compress the gas to complete the compression stroke as the offset increases. Therefore, the amount of offset is severely limited by its own disadvantages on the compression side. Accordingly, large prior art offsets, i.e., offsets in which the crankshaft must rotate at least 20 degrees past a pistons top dead center position before the piston can reach a firing position, have not been utilized, disclosed or taught. As a result, the relatively large offsets required to substantially align peak torque to peak combustion pressure cannot be accomplished with Lee&#39;s invention. 
     Variable Compression Ratio (VCR) engines are a class of prior art CI engines designed to take advantage of varying the compression ratio on an engine to increase efficiency. One such typical example is disclosed in U.S. Pat. No. 4,955,328 to Sobotowski. Sobotowski describes an engine in which compression ratio is varied by altering the phase relation between two pistons operating in cylinders interconnected through a transfer port that lets the gas flow in both directions. 
     However, altering the phase relation to vary compression ratios impose design requirements on the engine that greatly increase its complexity and decrease its utility. For example, each piston of the pair of pistons must reciprocate through all four strokes of a complete four stroke cycle, and must be driven by a pair of crankshafts which rotate through two full revolutions per four stroke cycle. Additionally, the linkages between the pair of crankshafts become very complex and heavy. Also the engine is limited by design to CI engines due to the higher compression ratios involved. 
     Various other relatively recent specialized prior art engines have also been designed in an attempt to increase engine efficiency. One such engine is described in U.S. Pat. No. 5,546,897 to Bracken entitled “Internal Combustion Engine with Stroke Specialized Cylinders”. In Brackett, the engine is divided into a working section and a compressor section. The compressor section delivers charged air to the working section, which utilizes a scotch yoke or conjugate drive motion translator design to enhance efficiency. The specialized engine can be described as a horizontally opposed engine in which a pair of opposed pistons reciprocate in opposing directions within one cylinder block. 
     However, the compressor is designed essentially as a super charger which delivers supercharged gas to the working section. Each piston in the working section must reciprocate through all four strokes of intake, compression, power and exhaust, as each crankshaft involved must complete two full revolutions per four-stroke cycle. Additionally, the design is complex, expensive and limited to very specialized CI engines. 
     Another specialized prior art design is described in U.S. Pat. No. 5,623,894 to Clarke entitled “Dual Compression and Dual Expansion Engine”. Clarke essentially discloses a specialized two-stroke engine where opposing pistons are disposed in a single cylinder to perform a power stroke and a compression stroke. The single cylinder and the crowns of the opposing pistons define a combustion chamber, which is located in a reciprocating inner housing. Intake and exhaust of the gas into and out of the combustion chamber is performed by specialized conical pistons, and the reciprocating inner housing. 
     However, the engine is a highly specialized two-stroke system in which the opposing pistons each perform a compression stroke and a power stroke in the same cylinder. Additionally, the design is very complex requiring dual crankshafts, four pistons and a reciprocating inner housing to complete the single revolution two-stroke cycle. Also, the engine is limited to large CI engine applications. 
     Accordingly, there is a need for an improved four-stroke internal combustion engine, which can enhance efficiency by more closely aligning the torque and force curves generated during a power stroke without increasing compression ratios substantially beyond normally accepted design limits. 
     SUMMARY OF THE INVENTION 
     The present invention offers advantages and alternatives over the prior art by providing a four-stroke cycle internal combustion engine having a pair of pistons in which one piston of the pair is used for the intake and compression strokes and another piston of the pair is used for the power and exhaust strokes, with each four stroke cycle being completed in one revolution of the crankshaft. The engine enhances efficiency by more closely aligning the torque and force curves generated during a power stroke without increasing compression ratios. 
     These and other advantages are accomplished in an exemplary embodiment of the invention by providing a four stroke cycle internal combustion engine including a crankshaft rotating about a crankshaft axis of the engine. A power piston is slidably received within a first cylinder and operatively connected to the crankshaft such that the power piston reciprocates through a power stroke and an exhaust stroke of a four stroke cycle during a single rotation of the crankshaft. A compression piston is slidably received within a second cylinder and operatively connected to the crankshaft such that the compression piston reciprocates through an intake stroke and a compression stroke of the same four stroke cycle during the same rotation of the crankshaft. A power piston-cylinder axis along which the power piston reciprocates within the first cylinder has an offset from the crankshaft axis such that the power piston-cylinder axis does not intersect the crankshaft axis. A gas passage interconnects the first and second cylinders. The gas passage includes an inlet valve and an outlet valve defining a pressure chamber therebetween. The inlet valve permits substantially one way flow of compressed gas from the second cylinder to the pressure chamber and the outlet valve permits substantially one way flow of compressed gas from the pressure chamber to the first cylinder. The inlet valve and the outlet valve of the gas passage substantially maintain at least a predetermined firing condition gas pressure in the pressure chamber during the entire four stroke cycle. 
     In an alternative embodiment of the invention the engine includes a compression piston-cylinder axis along which the compression piston reciprocates within the second cylinder. The compression piston-cylinder axis has an offset from the crankshaft axis such that the compression piston-cylinder axis does not intersect the crankshaft axis. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic diagram of a representative prior art four stroke cycle engine, during the intake stoke; 
         FIG. 2  is a schematic diagram of the prior art engine of  FIG. 1  during the compression stoke; 
         FIG. 3  is a schematic diagram of the prior art engine of  FIG. 1  during the power stoke; 
         FIG. 4  is a schematic diagram of the prior art engine of  FIG. 1  during the exhaust stoke; 
         FIG. 5  is a schematic diagram of the prior art engine of  FIG. 1  when the piston is at the position of maximum torque; 
         FIG. 6 , is a graphical representation of torque and combustion pressure of the prior art engine of  FIG. 1 ; 
         FIG. 7  is a schematic diagram of an engine in accordance with the present invention during the exhaust and intake strokes; 
         FIG. 8  is a schematic diagram of the engine of  FIG. 7  when the first piston has just reached top dead center (TDC) at the beginning of a power stroke; 
         FIG. 9  is a schematic diagram of the engine of  FIG. 7  when the first piston has reached its firing position; 
         FIG. 10 , is a graphical representation of torque and combustion pressure of the engine of  FIG. 7 ; and 
         FIG. 11  is a schematic diagram of an alternative embodiment of an engine in accordance with the present invention having unequal throws and piston diameters. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Referring to  FIG. 7 , an exemplary embodiment of a four stroke internal combustion engine in accordance with the present invention is shown generally at  100 . The engine  100  includes an engine block  102  having a first cylinder  104  and a second cylinder  106  extending therethrough. A crankshaft  108  is journaled for rotation about a crankshaft axis  110  (extending perpendicular to the plane of the paper). 
     The engine block  102  is the main structural member of the engine  100  and extends upward from the crankshaft  108  to the junction with the cylinder head  112 . The engine block  102  serves as the structural framework of the engine  100  and typically carries the mounting pad by which the engine is supported in the chassis (not shown). The engine block  102  is generally a casting with appropriate machined surfaces and threaded holes for attaching the cylinder head  112  and other units of the engine  100 . 
     The cylinders  104  and  106  are openings, typically of generally circular cross section, that extend through the upper portion of the engine block  102 . Cylinders are defined herein as the chambers within which pistons of an engine reciprocate, and do not have to be generally circular in cross section, e.g., they may have a generally elliptical or half moon shape. 
     The internal walls of cylinders  104  and  106  are bored and polished to form smooth, accurate bearing surfaces sized to receive a first power piston  114 , and a second compression piston  116  respectively. The power piston  114  reciprocates along a first piston-cylinder axis  113 , and the compression piston  116  reciprocates along a second piston-cylinder axis  115 . The first and second cylinders  104  and  106  are disposed in the engine  100  such that the first and second piston-cylinder axes  113  and  115  pass on opposing sides of the crankshaft axis  110  without intersecting the crankshaft axis  110 . 
     The pistons  114  and  116  are typically cup shaped cylindrical castings of steel or aluminum alloy. The upper closed ends, i.e., tops, of the power and compression pistons  114  and  116  are the first and second crowns  118  and  120  respectively. The outer surfaces of the pistons  114 , 116  are generally machined to fit the cylinder bore closely and are typically grooved to receive piston rings (not shown) that seal the gap between the pistons and the cylinder walls. 
     First and second connecting rods  122  and  124  each include an angle bend  121  and  123  respectively. The connecting rods  122  and  124  are pivotally attached at their top distal ends  126  and  128  to the power and compression pistons  114  and  116  respectively. The crankshaft  108  includes a pair of mechanically offset portions called the first and second throws  130  and  132 , which are pivotally attached to the bottom opposing distal ends  134  and  136  of the first and second connecting rods  122  and  124  respectively. The mechanical linkages of the connecting rods  122  and  124  to the pistons  114 ,  116  and crankshaft throws  130 , 132  serve to convert the reciprocating motion of the pistons (as indicated by directional arrow  138  for the power piston  114 , and directional arrow  140  for the compression piston  116 ) to the rotary motion (as indicated by directional arrow  142 ) of the crankshaft  108 . The first piston cylinder axis  113  is offset such that it is disposed in the imaginary half plane through which the first crankshaft throw  130  rotates from its top dead center position to its bottom dead center position. The second piston cylinder axis  115  is offset in the opposing imaginary half plane. 
     Though this embodiment shows the first and second pistons  114  and  116  connected directly to crankshaft  108  through connecting rods  122  and  124  respectively, it is within the scope of this invention that other means may also be employed to operatively connect the pistons  114  and  116  to the crankshaft  108 . For example a second crankshaft may be used to mechanically link the pistons  114  and  116  to the first crankshaft  108 . 
     The cylinder head  112  includes a gas passage  144  interconnecting the first and second cylinders  104  and  106 . The gas passage includes an inlet check valve  146  disposed in a distal end of the gas passage  144  proximate the second cylinder  106 . An outlet poppet valve  150  is also disposed in an opposing distal end of the gas passage  144  proximate the top of the first cylinder  104 . The inlet check valve  146  and outlet poppet valve  150  define a pressure chamber  148  there between. The inlet valve  146  permits the one way flow of compressed gas from the second cylinder  106  to the pressure chamber  148 . The outlet valve  150  permits the one way flow of compressed gas from the pressure chamber  148  to the first cylinder  104 . Though check and poppet type valves are described as the inlet and the outlet valves  146  and  150  respectively, any valve design appropriate for the application may be used instead, e.g., the inlet valve  146  may also be of the poppet type. 
     The cylinder head  112  also includes an intake valve  152  of the poppet type disposed over the top of the second cylinder  106 , and an exhaust valve  154  of the poppet type disposed over the top to the first cylinder  104 . Poppet valves  150 , 152  and  154  typically have a metal shaft  156  with a disk  158  at one end fitted to block the valve opening. The other end of the shafts  156  of poppet valves  150 ,  152  and  154  are mechanically linked to camshafts  160 , 162  and  164  respectively. The camshafts  160 , 162  and  164  are typically a round rod with generally oval shaped lobes located inside the engine block  102  or in the cylinder head  112 . 
     The camshafts  160 , 162  and  164  are mechanically connected to the crankshaft  108 , typically through a gear wheel, belt or chain links (not shown). When the crankshaft  108  forces the camshafts  160 ,  162  and  164  to turn, the lobes on the camshafts  160 ,  162  and  164  cause the valves  150 , 152  and  154  to open and close at precise moments in the engine&#39;s cycle. 
     The crown  120  of compression piston  116 , the walls of second cylinder  106  and the cylinder head  112  form a compression chamber  166  for the second cylinder  106 . The crown  118  of power piston  114 , the walls of first cylinder  104  and the cylinder head  112  form a separate combustion chamber  168  for the first cylinder  104 . A spark plug  170  is disposed in the cylinder head  112  over the first cylinder  104  and is controlled by a control device (not shown) which precisely times the ignition of the compressed air gas mixture in the combustion chamber  168 . Though this embodiment describes a spark ignition (SI) engine, one skilled in the art would recognize that compression ignition (CI) engines are within the scope of this invention also. 
     During operation, the power piston  114  leads the compression piston  116  by a phase shift angle  172 , defined by the degrees of rotation the crankshaft  108  must rotate after the power piston  114  has reached its top dead center position in order for the compression piston  116  to reach its respective top dead center position. Preferably this phase shift is between 30 to 60 degrees. For this particular preferred embodiment, the phase shift is fixed substantially at 50 degrees. 
       FIG. 7  illustrates the power piston  114  when it has reached its bottom dead center (BDC) position and has just started ascending (as indicated by arrow  138 ) into its exhaust stroke. Compression piston  116  is lagging the power piston  114  by 50 degrees and is descending (arrow  140 ) through its intake stroke. The inlet valve  156  is open to allow an explosive mixture of fuel and air to be drawn into the compression chamber  166 . The exhaust valve  154  is also open allowing piston  114  to force spent products of combustion out of the combustion chamber  168 . 
     The check valve  146  and poppet valve  150  of the gas passage  144  are closed to prevent the transfer of ignitable fuel and spent combustion products between the two chambers  166  and  168 . Additionally during the exhaust and intake strokes, the inlet check valve  146  and outlet poppet valve  150  seal the pressure chamber  148  to substantially maintain the pressure of any gas trapped therein from the previous compression and power strokes. 
     Referring to  FIG. 8 , the power piston  114  has reached its top dead center (TDC) position and is about to descend into its power stroke (indicated by arrow  138 ), while the compression piston  116  is ascending through its compression stroke (indicated by arrow  140 ). At this point, inlet check valve  146 , outlet valve  150 , intake valve  152  and exhaust valve  154  are all closed. 
     At TDC piston  114  has a clearance distance  178  between the crown  118  of the piston  114  and the top of the cylinder  104 . This clearance distance  178  is very small by comparison to the clearance distance  60  of standard engine  10  (best seen in FIG.  3 ). This is because the power stroke in engine  100  follows a low pressure exhaust stroke, while the power stroke in standard engine  10  follows a high pressure compression stroke. Therefore, in distinct contrast to the standard engine  10 , there is little penalty to engine  100  to reduce the clearance distance  178  since there is no high pressure gas trapped between the crown  118  and the top of the cylinder  114 . Moreover, by reducing the clearance distance  178 , a more thoroughly flushing of nearly all exhaust products is accomplished. 
     In order to substantially align the point of maximum torque with maximum combustion pressure, the crankshaft  108  must be rotated approximately 40 degrees past its top dead center position when the power piston  114  is in its optimal firing position. Additionally, similar considerations hold true on the compression piston  116 , in order to reduce the amount of torque and power consumed by the crankshaft  108  during a compression stroke. Both of these considerations require that the offsets on the piston-cylinder axes be much larger than any previous prior art offsets, i.e., offsets in which the crankshaft must rotate at least 20 degrees past a pistons top dead center position before the piston can reach a firing position. These offsets are in fact so large that a straight connecting rod linking the pistons  114  and  116  would interfere with the lower distal end of the cylinders  104  and  106  during a stroke. 
     Accordingly, the bend  121  in connecting rod  122  must be disposed intermediate its distal ends and have a magnitude such that the connecting rod  122  clears the bottom distal end  174  of cylinder  104  while the power piston  114  reciprocates through an entire stroke. Additionally, the bend  123  in connecting rod  124  must be disposed intermediate its distal ends and have a magnitude such that the connecting rod  124  clears the bottom distal end  176  of cylinder  106  while the compression piston  116  reciprocates through an entire stroke. 
     Referring to  FIG. 9 , the crankshaft  108  has rotated an additional 40 degrees (as indicated by arrow  180 ) past the TDC position of power piston  114  to reach its firing position, and the compression piston  116  is Just completing its compression stroke. During this 40 degrees of rotation, the compressed gas within the second cylinder  116  reaches a threshold pressure which forces the check valve  146  to open, while cam  162  is timed to also open outlet valve  150 . Therefore, as the power piston  114  descends and the compression piston  116  ascends, a substantially equal mass of compressed gas is transferred from the compression chamber  166  of the second cylinder  106  to the combustion chamber  168  of the first cylinder  104 . When the power piston  114  reaches its firing position, check valve  146  and outlet valve  150  close to prevent any further gas transfer through pressure chamber  148 . Accordingly, the mass and pressure of the gas within the pressure chamber  148  remain relatively constant before and after the gas transfer takes place. In other words, the gas pressure within the pressure chamber  148  is maintained at least (at or above) a predetermined firing condition pressure, e.g., approximately 270 psia, for the entire four stroke cycle. 
     By the time the power piston  114  has descended to its firing position from TDC, the clearance distance  178  has grown to substantially equal that of the clearance distance  60  of standard engine  10  (best seen in FIG.  3 ), i.e., 0.571. Additionally, the firing conditions are substantially the same as the firing conditions of the standard engine  10 , which are generally: 1) a 4 inch diameter piston, 2) a clearance volume of 7.181 cubic inches, 3) a pressure before ignition of approximately 270 pounds per square inch absolute (psia), and 4) a maximum combustion pressure after ignition of approximately 1200 psia. Moreover, the angle of the first throw  130  of crankshaft  108  is in its maximum torque position, i.e., approximately 40 degrees past TDC. &#39;Therefore, spark plug  170  is timed to fire such that maximum combustion pressure occurs when the power piston  114  substantially reaches its position of maximum torque. 
     During the next 10 degrees of rotation  142  of the crankshaft  108 , the compression piston  116  will pass through to its TDC position and thereafter start another intake stroke to begin the cycle over again. The compression piston  116  also has a very small clearance distance  182  relative to the standard engine  10 . This is possible because, as the gas pressure in the compression chamber  166  of the second cylinder  106  reaches the pressure in the pressure chamber  148 , the check valve  146  is forced open to allow gas to flow through. Therefore, very little high pressure gas is trapped at the top of the power piston  116  when it reaches its TDC position. 
     The compression ratio of engine  100  can be anything within the realm of SI or CI engines, but for this exemplary embodiment it is substantially within the range of 6 to 8.5. As defined earlier, the compression ratio is the maximum volume of a predetermined mass of an air-fuel mixture before a compression stroke, divided by the volume of the mass of the air-fuel mixture at the point of ignition. For the engine  100 , the compression ratio is substantially the ratio of the displacement volume in second cylinder  106  when the compression piston  116  travels from SDC to TDC to the volume in the first cylinder  104  when the power piston  114  is at its firing position. 
     In distinct contrast to the standard engine  10  where the compression stroke and the power stroke are always performed in sequence by the same piston, the power stroke is performed by the power piston  114  only, and the compression stroke is performed by the compression piston  116  only. Therefore, the power piston  116  can be offset to align maximum combustion pressure with maximum torque applied to the crankshaft  108  without incurring penalty for being out of alignment on the compression stroke. Vice versa, the compression piston  114  can be offset to align maximum compression pressure with maximum torque applied from the crankshaft  108  without incurring penalty for being out of alignment on the power stroke. 
     Referring to  FIG. 10 , this concept can be further illustrated. Here, a graph of tangential force or torque versus degrees of rotation from TDC for power piston  114  is shown at  184  for the engine  100 . Additionally, a graph of combustion pressure versus degrees of rotation from TDC for power piston  114  is shown at  186  for engine  100 . The calculations for the graphs  184  and  186  were based on the engine  100  having firing conditions substantially equal to that of a standard engine. That is: 1) a 4 inch diameter piston, 2) a clearance volume of 7.181 cubic inches, 3) a pressure before ignition of approximately 270 pounds per square inch absolute (psia), 4) a maximum combustion pressure after ignition of approximately 1200 psia and  5 ) substantially equal revolutions per minute (RPM) of the crankshafts  108  and  38 . In distinct contrast with the graphs of  FIG. 6  for the standard prior art engine  10 , the point of maximum combustion pressure  188  is substantially aligned with the point of maximum torque  190 . This alignment of combustion pressure  186  with torque  184  results in a significant increase in efficiency. 
     Moreover, the compression piston&#39;s  116  offset can also be optimized to substantially align the maximum torque delivered to the compression piston  116  from the crankshaft  108  with the maximum compression pressure of the gas. The compression pistons  116  offset reduces the amount of power exerted in order to complete a compression stroke and further increases the overall efficiency of engine  100  relative to the standard engine  10 . With the combined power and compression piston  114 , and  116  offsets, the overall theoretical efficiency of engine  100  can be increased by approximately 20 to 40 percent relative to the standard engine. 
     Referring to  FIG. 11 , an alternative embodiment of a split four stroke engine having unequal throws and unequal piston diameters is shown generally at  200 . Because the compression and power strokes are performed by separate pistons  114 ,  116 , various enhancements can be made to optimize the efficiency of each stroke without the associated penalties incurred when the strokes are performed by a single piston. For example, the compression piston diameter  204  can be made larger than the power piston diameter  202  to further increase the efficiency of compression. Additionally, the radius  206  of the first throw  130  for the power piston  114  can be made larger than the radius  208  of the second throw  132  for the compression piston  116  to further enhance the total torque applied to the crankshaft  108 . 
     While preferred embodiments have been shown and described, various modifications and substitutions may be made thereto without departing from the spirit and scope of the invention. Accordingly, it is to be understood that the present invention has been described by way of illustration and not limitation.