Abstract:
An internal combustion engine of the present invention features separate compression and expansion cycles. The engine includes a separate compressor device which pressurizes air by a ratio greater than 15 to 1, at least one two stroke combustion cylinder and a compressed air conduit for transferring compressed air from the compressor to the at least one combustion cylinder. An air injection valve injects the compressed air into the combustion cylinder during the second half portion of the return stroke of the combustion cylinder. The compressed air is mixed with fuel and combusted for expansion during a power stroke. In this engine compression occurs only to a minor degree in the combustion cylinder. Accordingly, the compression ratio of the present engine may be significantly higher or lower than the volumetric expansion ratio of the combustion cylinder thus resulting in corresponding increases in either power density or thermodynamic efficiency respectively.

Description:
[0001]     This application is a continuation of application Ser. No. 10/777,796 which was filed Feb. 12, 2004.  
         [0002]     U.S. application Ser. 10/777,796 claimed the benefit of U.S. Provisional Patent Application No. 60/446,934 filed Feb. 12, 2003. 
     
    
     FIELD OF THE INVENTION  
       [0003]     The present invention relates to an internal combustion engine.  
       BACKGROUND OF THE INVENTION  
       [0004]     Many types of internal combustion engines are known in the art. It is well known that increasing the compression ratio of an internal combustion engine will result in increased thermodynamic efficiency. With many prior art engines, the compression ratio of the engine is limited by the expansion ratio of the cylinders of the engine. In other prior art engines, the compression ratio of the engine is limited to a relatively low value because auto-ignition of the fuel air mixture will occur too early in the cycle as compressed air reaches a temperature above the auto-ignition temperature of the fuel.  
       BRIEF DESCRIPTION OF THE INVENTION  
       [0005]     In an embodiment of the present invention the aforementioned problems are addressed by providing an internal combustion engine in which the compression and expansion portions of the engine&#39;s cycle and the compression and expansion ratios are independent. The present engine includes a compressor which pressurizes air by a ratio which may be substantially more than 15 to 1, a combustion cylinder including a reciprocating piston which oscillates between a top dead center position and a bottom dead center position in a power stroke and between the bottom dead center position and the top dead center position in a return stroke and a compressed air conduit for transferring compressed air from the compressor to the combustion cylinder. Pneumatic communication between the compressed air conduit and the combustion cylinder is governed by a timed valve which intermittently opens to release pressurized air into the combustion cylinder when the piston is in the second half portion of the return stroke. A fuel injector is employed to mix fuel with the pressurized air to make a fuel—air mixture which is combusted to produce hot, high pressure gaseous combustion products which expand during the power stroke. In this present engine, because the compression of air for use in the combustion portion of the cycle is conducted separately and then injected or released into the combustion cylinder when it is needed, the ratio of compression can be significantly higher or lower than the ratio of expansion. A higher expansion ratio results in a significant increase in thermodynamic efficiency while a higher compression ratio results in a significant increase in power density. Moreover, since the present engine conducts compression and expansion separately, compressed air for use in the combustion cylinder may be cooled to prevent early ignition of a fuel air mixture thus permitting a higher compression ratio.  
         [0006]     The injection of pressurized air from the compressed air conduit into the combustion cylinder preferably occurs during a relatively small portion of the combustion cylinder cycle preferably when the piston is in the second half of the return stroke. Accordingly, a timed valve such as an indexed rotary valve which presents a relatively large flow area may be used to provide timed intermittent pneumatic communication between the compressed air conduit and the combustion cylinder. Such a valve arrangement should therefore provide timed, intermittent pneumatic communication between the compressed air conduit and the combustion cylinder sufficient to allow air pressure in the compressed air conduit and the combustion chamber to substantially equalize during a relatively small portion of the combustion cylinder cycle when the piston of the combustion cylinder is in the second half of the return stroke.  
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0007]      FIG. 1  is a diagram of the internal combustion engine of the present invention.  
         [0008]      FIG. 1A  is a diagram of an embodiment of the internal combustion engine of the present invention having three combustion cylinders and two compression cylinders.  
         [0009]      FIG. 2A  shows the compression cylinder during its intake stroke.  
         [0010]      FIG. 2B  shows the compression cylinder at bottom dead center.  
         [0011]      FIG. 2C  shows the compression cylinder during its compression stroke.  
         [0012]      FIG. 3  is a timing diagram showing the timing of segments A-G of the combustion cylinder cycle shown in  FIG. 3A - FIG. 3H  and the combustion cylinder cycle shown in  FIGS. 4A-4H .  
         [0013]      FIG. 3A  shows the combustion cylinder during cycle segment A as gaseous combustion products remaining from a previous cycle are being expelled through the exhaust valve.  
         [0014]      FIG. 3B  shows the combustion cylinder during cycle segment where A and B 2  overlap when the pressurized air injection valve and the exhaust valve are both open in order to scavenge the last remaining gaseous combustion products from the previous cycle.  
         [0015]      FIG. 3C  shows the combustion cylinder during cycle segment B 2  as pressurized air is being injected into the combustion cylinder.  
         [0016]      FIG. 3D  shows the combustion cylinder during cycle segment C as fuel is being injected into the combustion cylinder.  
         [0017]      FIG. 3E  shows the combustion cylinder during cycle point D where the combustion piston is near top dead center and the fuel air mixture is being ignited by a spark plug.  
         [0018]      FIG. 3F  shows the combustion cylinder during cycle segment E where the combustion piston is at top dead center as the fuel air mixture is in the process of combustion.  
         [0019]      FIG. 3G  shows the combustion cylinder during cycle segment F during the power stroke as combustion product gases are expanding.  
         [0020]      FIG. 3H  shows the combustion cylinder during optional cycle segment G where the combustion piston is at bottom dead center as gaseous combustion products escape through the exposed exhaust port.  
         [0021]      FIG. 31  is an isometric view of the combustion cylinder with an indexed rotary valve.  
         [0022]      FIG. 3J  is an cross sectional view showing a cross section of the valve housing, the valve body and the combustion cylinder generally taken from plane A-A of  FIG. 31  except with the valve body in the position shown in  FIG. 3L   
         [0023]      FIG. 3K  is an isometric view of the indexed rotary valve with the valve housing removed for clarity as the valve body comes to rest at the end of a 90 degree rotation.  
         [0024]      FIG. 3L  is an isometric view of the indexed rotary valve with the valve housing removed for clarity as the valve body begins a next 90 rotation.  
         [0025]      FIG. 3M  is an isometric view of the indexed rotary valve with the valve housing removed for clarity as the valve body is rotating at a high speed.  
         [0026]      FIG. 3N  is an isometric view of the indexed rotary valve with the valve housing removed for clarity as the valve body comes to rest at the end of a 90 degree rotation.  
         [0027]      FIG. 3P  is an plot showing valve body rotational velocity as a function of crankshaft position for the indexed rotary valve arrangement shown in  FIGS. 3I-3N .  
         [0028]      FIG. 4A  shows the combustion cylinder including a rotary injection valve during cycle segment A as gaseous combustion products remaining from a previous cycle are being expelled through the exhaust valve.  
         [0029]      FIG. 4B  shows the combustion cylinder including a rotary injection valve during cycle portion where A and B 2  overlap when the pressurized air injection valve and the exhaust valve are both open in order to scavenge the last remaining gaseous combustion products from the previous cycle.  
         [0030]      FIG. 4C  shows the combustion cylinder including a rotary injection valve during cycle segment B 2  as pressurized air is being injected into the combustion cylinder.  
         [0031]      FIG. 4D  shows the combustion cylinder including a rotary injection valve during cycle segment C as fuel is being injected into the combustion cylinder.  
         [0032]      FIG. 4E  shows the combustion cylinder including a rotary injection valve during cycle point D where the combustion piston is near top dead center and the fuel air mixture is being ignited by a spark plug.  
         [0033]      FIG. 4F  shows the combustion cylinder including a rotary injection valve during cycle segment E where the combustion piston is at top dead center as the fuel air mixture is in the process of combustion.  
         [0034]      FIG. 4G  shows the combustion cylinder during cycle segment F during the power stroke as combustion product gases are expanding.  
         [0035]      FIG. 4H  shows the combustion cylinder during optional cycle segment G where the combustion piston is at bottom dead center as gaseous combustion products escape through the exposed exhaust port.  
         [0036]      FIG. 5  is a Pressure verses Specific Volume graph for the thermodynamic cycle of the of an embodiment of the invention internal combustion engine having an inter-cooler for cooling pressurized air.  
         [0037]      FIG. 6  is a Temperature versus Entropy graph for the thermodynamic cycle of the of an embodiment of the invention internal combustion engine having an inter-cooler for cooling pressurized air.  
         [0038]      FIG. 7  is an illustrative plot of Power versus Compression Ratio for curves of a set value for Expansion Ratio.  
         [0039]      FIG. 8  is an illustrative plot of Thermodynamic Efficiency versus Compression Ratio for curves of a set value for Expansion Ratio.  
     
    
     DETAILED DESCRIPTION  
       [0040]     Referring to the drawings,  FIG. 1  illustrates an internal combustion engine 10 in accordance with an embodiment of the invention. In  FIG. 1 , an-internal combustion engine  10  is shown including compressor  12 , compressed air conduit  50  and combustion cylinder  70 . Combustion cylinder  70  includes a cylinder  74  and a reciprocating piston  76  which is one of the mechanical arrangements for defining an internal combustion engine which features a combustion chamber that cycles between a minimum volume and a maximum volume. Combustion cylinder  70  may be one of two or more combustion cylinders coupled together on a common crankshaft  76 D. Likewise compressor  12  may include a compression cylinder  13  as shown in  FIG. 1 . Compressor  12  provides compressed air to compressed air conduit  50 . Together, compressor  12  and compressed air conduit  50  provide a source of compressed air for use by combustion cylinder  70 .  
         [0041]      FIG. 1A  schematically presents an example embodiment of the present engine  10 A having three combustion cylinders  70  associated on common crankshaft  76 D and a compressor  12  comprising two compression cylinders  13  associated on a common compressor crankshaft  18 D. In  FIG. 1A , crankshaft  76 D and compressor crankshaft  18 D are coupled by a variable ratio gear box  12 A. This variable ratio gear box may be adjusted to adjust the volume of compressed air delivered to compressed air conduit  50 . The advantage of having a the capability to control the delivery of the compressed air within conduit  50  are described in detail below but generally allow an adjustment in operating conditions between a mode having a relatively low volumetric compression ratio and a relatively high expansion ratio for maximum thermodynamic efficiency and a mode of relatively high volumetric compression ratio and a relatively low expansion ratio for maximum power density. The combustion cylinders  70  of example engine  10 A each include injection valves  72 A, exhaust valves  72 B, fuel injectors  72 C and ignition initiators  72 D.  FIG. 1A  also illustrates a timing system  300  for timing the operations of injection valves  72 A, exhaust valves  72 B, fuel injectors  72 C and ignition initiators  72 D. Such a timing system is needed for the operation of an internal combustion engine but is omitted from many of the other figures for clarity. Timing system  300 , in this example, includes a cam shaft  302 , a fuel injection timer  304  and an ignition timer  306 . Cam shaft  302  is mechanically coupled to crankshaft  76 D and carries a series of eccentric cams for governing the operations of injection valves  72 A and exhaust valves  72 B. Fuel injection timer  304  governs the operations of fuel injectors  72 C, while ignition timer  306  governs the operations of ignition initiators  72 D. Both fuel injection timer  304  and ignition timer  306  are coupled to crankshaft  76 D. Timing system  300  as presented here is only one of many possible timing systems and the selection here of particular types of components is not intended to limit the scope of the invention.  FIG. 1A  also illustrates that combustion cylinder  70  may be one of a plurality combustion cylinders coupled by a common crankshaft.  FIG. 1A  is not intended to suggest that compressor  12  must be a cylinder—piston type compressor or that compressor  12  would be limited to having two compression cylinders.  
         [0042]     Compressor  12  takes in air from the outside environment and delivers compressed air to compressed air conduit  50 . In the embodiment shown in  FIG. 1 , compressor  12  is a compression cylinder  13  which further includes a compression cylinder head  14 , a compression cylinder body  16  and a compression piston  18 . The upper surface of compression piston  18 , the inside wall of compression cylinder body  16  and compression cylinder head  14  define compression chamber  16 A which constantly changes in volume as compression piston reciprocates with compression cylinder  13 . Compression piston  18  is connected by a connecting rod  18 C to a compression crankshaft  18 D. Compression cylinder head  14  includes an intake valve  14 A and an outlet valve  14 D. Intake valve  14 A governs pneumatic communication between an intake port  14 B leading to the outside environment and compression chamber  16 A. Outlet valve  14 D governs pneumatic communication between compression chamber  16 A and an outlet port  14 E which leads to compressed air conduit  50 .  
         [0043]     Compressed air conduit  50  retains compressed air produced by compressor  12  and conveys compressed air to combustion cylinder  70 . In the embodiment shown in  FIG. 1 , compressed air conduit  50  generally includes a storage means and a cooling means so that a supply of temperature conditioned pressurized air may be available for use by combustion cylinder  70 . In the embodiment shown in  FIG. 1 , compressed air conduit  50  further includes an intake portion  52 , an insulated reservoir  54 , a heat rejecting portion 56 having heat rejecting fins  56 A, a cool compressed air valve  60 , an insulated hot air conduit  54 A, hot compressed air valve  62 , a pressure regulator  64  and an outlet portion  66 . Cool compressed air valve  60  and hot compressed air valve  62  can be adjusted in order to adjust the temperature of air in outlet portion  66  as will be described in more detail below. Pressure regulator  64  is for regulating the pressure of the pressurized air in outlet portion  66 . Preferably, reservoir  54  should encompass a volume sufficient to provide a steady supply of compressed air for use by combustion cylinder  70 .  
         [0044]     Combustion cylinder  70  receives compressed air from compressed air conduit  50  as well as fuel which is mixed with the compressed air for combustion and expansion in a power stroke. In the embodiment shown in  FIG. 1 , combustion cylinder  70  is a two stroke cylinder having a piston which oscillates in a cycle including a power stroke in which the piston moves from a top dead center position to a bottom dead center position and a return stroke in which the piston moves from the bottom dead center position to the top dead center position. Generally, the injection of compressed air from compressed air conduit  50  into combustion cylinder  70  is timed to occur during a relatively short portion of the cycle when the piston is in the second half of the return stroke. Also generally, the injection of fuel into combustion cylinder  70  is preferably timed to occur after the injection of compressed air has begun. The combustion of the fuel air mixture preferably occurs after the injection of compressed air and fuel and preferably not substantially prior to the piston reaching top dead center. In the embodiment shown in  FIG. 1 , combustion cylinder  70  further includes a combustion cylinder head  72 , a combustion cylinder body  74  and a combustion piston  76  having an upper piston surface  76 A. A connecting rod  76 C links combustion piston  76  to an associated crankshaft  76 D for the conversion of the reciprocating motion of the piston into rotational power at the crankshaft  76 D. Combustion cylinder body  74  includes a cylindrical inside wall  74 A which may be penetrated by an optional exhaust port  74 C. Exhaust port  74 C and exhaust valve  72 B are examples of typical devices or means employed for releasing exhaust from a combustion chamber. Combustion cylinder head  72  further includes an injection valve  72 A, an exhaust valve  72 B, a fuel injector  72 C and may also include an ignition initiator  72 D which in  FIG. 1  is shown as a spark plug. Combustion cylinder  70  may optionally be arranged as a Diesel cylinder which compresses a mixture of air and fuel to a sufficient pressure to cause auto ignition of the mixture. As a Diesel cylinder, combustion cylinder  70  would not need ignition initiator  72 D. Combustion cylinder head  72 , inside wall  74 A of cylinder body  74  and upper piston surface  76 A define a combustion chamber  74 B which constantly changes in volume as piston  76  moves between a bottom dead center position as shown in  FIGS. 3H  or  4 H and a top dead center which would appear to be half way between the positions shown in  FIGS. 3E and 3F  or  FIGS. 4E and 4F .  
         [0045]      FIG. 1  illustrates combustion cylinder  70  such that pressurized air valve  72 A is a conventional stem valve.  FIGS. 3A-3H  illustrate the operation of power cylinder  70  with a conventional stem valve. With a typical prior art engine, a stem valve for regulating air intake may be open during a relatively large portion of crankshaft cycle corresponding to approximately 180 degrees of crankshaft rotation. With the present engine, a pressurized air valve  72 A may be open during a relatively small portion of the crankshaft cycle corresponding to 10 to 15 degrees of the crankshaft rotation. Because of the mechanical characteristics of stem valves, the actuation of a stem valve for such a small portion of the crankshaft cycle may limit the operating RPM of power cylinder  70 . Accordingly, in order to achieve higher RPMs, it would be preferable to employ a valve arrangement capable of substantially equalizing the pressure between the pressurized portion of the system such as outlet portion  66  of compressed air conduit  50  and combustion chamber  74 B during a relatively small portion of the crankshaft cycle.  FIGS. 3J-3N  illustrate an indexed rotary valve  82  adapted for filling combustion chamber  74 B with pressurized air during a relatively small portion of the cycle. Also shown in  FIG. 31  is an example timing system  300  which includes a timing chain  300 B coupled to crankshaft  76 D for driving a cam shaft  302  for actuating exhaust valve  72 B, a timing sensor  300 A associated with drive wheel  92  of rotary valve  82  which is also driven by timing chain  300 B and a timing unit  305  which receives input from timing sensor  300 A for controlling the timing of fuel injector  72 C and ignition initiator  72 D.  
         [0046]     As can be seen with reference to  FIGS. 3I , rotary valve  82  generally includes a valve portion  84 and an indexing portion  90 . Valve portion  84 is mounted to power cylinder  70  as shown in  FIG. 31 . Valve portion  84  may be best understood by referring to FIGS.  3 J-3M. The cross section view of  FIG. 3J  is taken from plane A-A of  FIG. 31 , except that valve body  88  in  FIG. 3J  is rotated to a position corresponding to that shown in  FIG. 3M . As can be best seen in  FIG. 3J , valve portion  84 includes a valve housing  86  which rotatably carries a valve body  88 . Valve body  88  includes two intersecting passages  88 A of generally oval cross-section which are arranged at right angles with respect to each other. Valve housing  86  has a compatible longitudinal bore  86 A for carrying valve body  88  as well as bearings adapted for high speed rotation of valve body. Valve housing  86  includes a pressurized air conduit opening  86 B which opens up to a generally oval shaped inlet port  86 C. Inlet port  86 C may be generally shaped to match the shape of passages  88 A of valve body  88 . However, inlet port  86 C is preferably not sealed against valve body  88  so that passages  88 A are constantly in communication with the pressurized volume inside housing  86  and thus outlet portion  66  of pressurized conduit  50 . This constant pressurization of passages  88 A occurs regardless of their rotational position within valve housing  86 . Valve housing  86  includes an oval shaped injection port  86 D which is oval shaped to match the shape of passages  88 A. However, unlike inlet port  86 C, injection port  86 D is sealed between valve body  88  and the constantly pressurized internal volume of valve housing  86  by an injection seal  88 E. A second housing seal  88 F seals the pressurized internal volume of valve housing  86  and passages  88 A from the outside environment. The above described compatible ports and passages are preferably shaped to maximize pneumatic communication between the pressurized portion of the system and combustion chamber  74 B.  
         [0047]     The purpose of indexing portion  90  is to cause the intermittent (or “indexed”)  90  degree rotation of valve body  88  during a 90 degree portion of a complete cycle of constantly rotating crankshaft  76 D. Indexing portion  90  includes a drive wheel  92  mechanically coupled to crankshaft  76 D for constant rotation and an index wheel  94  mechanically coupled to valve body  88  for intermittent, indexed rotation. Drive wheel  92  includes a cog  92 A and a retaining disc  92 B having a scalloped portion  92 C and a non-scalloped circular retaining portion  92 D. Index wheel  94  includes slots  94 A for receiving cog  92 A and external scallops  94 B for receiving non-scalloped retaining portion  92 D of retaining disc  92 B.  FIGS. 3K-3N  illustrate the relative motions of continuously rotating drive wheel  92  and intermittently rotating index wheel  94 . Valve housing  86  has been removed in  FIGS. 3K-3N  for clarity. In  FIG. 3K , drive wheel  92  is beginning a period of rotation in which it rotates clockwise for 270 degrees while index wheel  94  remains stationary in a position that blocks communication between inlet passage  86 C and combustion cylinder  70 . In  FIG. 3L , cog  92 A of drive wheel  92  has traveled clockwise 270 degrees and begins to engage slot  94 A of index wheel  94  thus causing index wheel  94  to begin rotating in a counter clockwise direction. In  FIG. 3M , index wheel  94  is rotating at a high speed relative to crankshaft  76 D and drive wheel  92 . The relative positions of valve body  88  and valve housing  84  illustrated in  FIG. 3M  are also shown in the cross sectional view of  FIG. 3J . In  FIG. 3N , index wheel  94  has advanced 90 degrees from the position shown in  FIG. 3M  and is again stationary while continuously rotating drive wheel  92  has returned to the position shown in  FIG. 3K .  FIG. 3P  provides plot which interrelates the rotational velocity of crankshaft  76 D, which is constant, and the rotational velocity of valve body  88  which varies greatly during a 90 degree portion of the crankshaft cycle. The mechanism described here for driving the rotary valve is commonly known as a Geneva wheel mechanism and is only one of many possible ways to accomplish the above stated objective, which is, to open communication between a pressurized volume and combustion chamber  74 B in a rapid and intermittent manner during a relatively small portion of the crankshaft cycle and to open such communication sufficiently to allow the substantial equalization of air pressure between the pressurized volume of the system and the combustion chamber  
         [0048]      FIG. 1  shows compression cylinder  13  almost half way through an intake stroke and combustion cylinder  70  at the beginning of the second half of the return stroke. However, these relative positions are not intended to imply a relationship between the two cylinders. In  FIG. 1 , no direct mechanical connection is shown between compression cylinder  13  and combustion cylinder  70 . Compression cylinder  13  and combustion cylinder  70  can be coupled by a common crankshaft or could be coupled such they operate at substantially different speeds. The applicant intends however, that a portion of the power derived from the operation of combustion cylinder  70  be used to power compressor  12 .  
         [0049]      FIG. 1  illustrates compression cylinder  13  and combustion cylinder  70  as if they would be equivalent in quantity, size and shape. This would probably not be the case.  
         [0050]      FIGS. 2A -2C  illustrate the operation of compression cylinder  13 .  FIG. 2A  shows compression cylinder  13  during its intake stroke. In  FIG. 2A , intake valve  14 A is open, outlet valve  14 B is closed and compression piston  18  is descending as air is pulled into compression chamber  16 A. In  FIG. 2B , compression cylinder is at bottom dead center and intake valve  14 A and outlet valve  14 B are both closed. In  FIG. 2C , intake valve  14 A is closed and outlet valve  14 B is open as the ascending compression piston  18  is forcing compressed air into intake portion  52  of compressed air conduit  50 . This positive displacement compressor shown in  FIG. 1  and  FIGS. 2A-2C  is of a type that is well know in the art. However, it could be replaced by any suitable compressor means that is capable of delivering compressed air with a compression ratio above 15 to 1.  
         [0051]     Compressed air conduit  50  is intended to receive and store compressed air and then deliver it to combustion cylinder  70  within desired temperature and pressure ranges. Compression cylinder  13  as shown in  FIG. 2A  is intended to compress air at a ratio substantially in excess of 15 to 1. It should be noted that air at an ambient temperature and pressures (such as 20° C. and one atmosphere of pressure), when compressed at 15 to 1, will increase in temperature to a temperature that may be above the auto-ignition temperature of a desired fuel. Accordingly, compressed air conduit  50  includes a heat rejecting portion  56 A having heat rejecting fins  56 A for rejecting a portion of the heat present in the compressed air leaving compression cylinder  13 . On the other hand, insulated reservoir  54  of compressed air  50  stores compressed air with minimal heat loss. Cool compressed air valve  60  and hot compressed air valve  62  for adjusting the flow through a hot conduit  54 A can be adjusted to mix an air stream that is controlled within a pre-selected temperature range that is below the auto-ignition temperature of a desired fuel. The presence of this temperature control feature is merely a preferred feature for use with an engine that is intended for burning fuels subject to auto-ignition. In the alternative, this temperature control feature may be useful even where premature auto-ignition is not an issue.  
         [0052]      FIG. 3A- 3H  diagram the operation of combustion cylinder  70 .  FIGS. 4A-4H  diagram the operation of combustion cylinder  70  with a rotary valve  82  as shown in  FIGS. 31-3N  instead of a stem type injection valve  72 A.  FIG. 3  provides a corresponding timing diagram which shows the relative timing of the positions shown in  FIGS. 3A-3H  and  FIGS. 4A-4H . The timing diagram of  FIG. 3  can be envisioned as being divided into segments which may overlap. These segments further correspond to the various configurations shown in the other figures including  FIGS. 3A-3P  and  FIGS. 4A-4H . Segment A corresponds to  FIGS. 3A and 4A  to the extent that valve  72 B of  FIGS. 3A and 4A  are open during segment A, yet segment A also corresponds to a relatively large portion of the crankshaft cycle whereas  FIGS. 3A and 4A  only show piston  76  and connecting rod  76 C in one position rather than a range of positions. During this segment, exhaust gasses are expelled from combustion cylinder  70  as piston  76  executes a portion of its return stroke. Segment B 1  in  FIG. 3  corresponds to the intermittent rotation of valve body  88  of indexed rotary valve  82  and is only applicable to the rotary valve configuration illustrated in  FIGS. 31-3N  and  FIGS. 4A-4H . Segment B 2  is preferably centered in segment B 1 . Segment B 2  corresponds to the portion of the cycle in which one of passages  88 A of valve body  88  is in communication with injection port  86 D of valve housing  86  thus providing open communication between valve housing  86  (and thus by extension compressed air conduit  50 ) and combustion chamber  74 B. In the rotary valve case, the center of segment B 2  corresponds with the alignment of one of passages  88 A with injection port  86 D as illustrated in FIG.  3 J. Yet, for the stem valve case, segment B 2  also corresponds to the portion of the cycle when injection valve  72 A is open. Note that segment A and segment B 2  slightly overlap indicating the scavenging of exhaust gasses from combustion chamber 74B. Such scavenging is illustrated in  FIGS. 3B and 4B . If a simple stem valve is used for an injection valve, then segment B 1  is omitted and the overlapping portion of segment A and segment B 2  would correspond to  FIG. 3B . Again, if a stem type injection valve is used, then the portion of segment B 2  not overlapping with segment A would correspond to  FIG. 3C  where pressurized air is being injected into combustion cylinder  70 . Segment C corresponds to the injection of fuel shown in  FIGS. 3D and 4D . Location D corresponds to the activation of an ignition initiator or spark plug as shown in  FIGS. 3E and 4E . As has been noted above location D as well as ignition initiator  72 D are optional and may be omitted if a Diesel type engine is desired. Fuel injection of segment C of  FIG. 3  may overlap or fall completely within the air injection portion B 2  as desired by the engine designer. Those skilled in the art of engine design should appreciate that both air injection portion B 2  and fuel injection portion C should be completed prior to the action of ignition initiator  72 D or in the case of a Diesel, the air injection should be complete prior to fuel injection which will result in auto-ignition. Since the combustion piston  76  is traveling upward towards the top dead center position during these segments of the cycle, a slight recompression of the injected fuel—air mixture will occur. This recompression effect can be minimized and compensated for by proper design of the engine cycle. Segment E corresponds to the combustion phase shown in  FIGS. 3F and 4F . Segment F corresponds to the expansion portion of the cycle depicted in  FIGS. 3G and 4G . Optionally, segment G, indicates the exposure of optional exhaust port  74 C shown in  FIGS. 3A-3H  but omitted in  FIGS. 4A-4H .  
         [0053]     A timing diagram such as the diagram of  FIG. 3  is not provided here to illustrate the operation of compression cylinder  13  as shown in  FIGS. 2A-2C . This is because the timing of the intake and compression portions for compression cylinder  13  is so simple that it can even be managed with the use of spring loaded valves. However, the various process described above can be related to thermodynamic diagrams  FIG. 5  and  FIG. 6 . Although, the present engine may a compression cycle that is mechanically separated from the combustion cycle,  FIG. 5  and  FIG. 6  show how these separate mechanical cycles inter-relate in a single thermodynamic cycle.  
         [0054]      FIGS. 5 and 6  are thermodynamic plots of the type typically used by those skilled in the art to diagram thermodynamic cycles. These plots present the state of the working fluid, which in this case is air, during the course of each cycle. The paths traced between points  1 ,  2 ,  3  and  4  in  FIGS. 5 and 6  represent the standard Otto cycle of a typical prior art internal combustion engine. The paths traced between points  1 ,  2 A,  2 B, 3A and  4 A represent the thermodynamic cycle of the present internal combustion engine  10 .  
         [0055]      FIG. 5  is thermodynamic plots of pressure verses specific volume, while  FIG. 6  gives thermodynamic plots of temperature versus entropy. Specific volume is merely the inverse of density and can be expressed in cubic meters per kilogram. For many, entropy is a more difficult concept to grasp. It could be understood as the degree by which a working fluid (such as air in the present engine) deviates from the prevailing conditions of the surrounding environment. So, for example, in  FIG. 6 , state point  3 A correlates to the end of the combustion process when gasses in combustion chamber  74 B are very hot and at a very high pressure—a high entropy condition which differs greatly from ambient conditions. By contrast, state point  1  in  FIG. 6  corresponds to ambient air prior to its intake in the compression stroke of compression cylinder  13 —a condition that does not differ from the low entropy condition of the surrounding environment.  
         [0056]     As noted above, in  FIG. 5  and  FIG. 6 , the thermodynamic cycle for a typical prior art Otto cycle engine is represented by a cycle that follows a path including state points  1 ,  2 ,  3  and  4 . Compression occurs between state points 1 and  2 , combustion occurs between state points  2  and  3 , expansion of combustion gasses occurs between state points  3  and  4  and the exhaust of the gaseous combustion products occurs between state points  4  and  1 . Generally, in a typical prior art engine, thermodynamic efficiency is understood as the ratio of the useful work captured between state points  3  and  4  and the energy input needed for compression and fuel combustion occurring between state points  1  and  3 .  
         [0057]     In  FIG. 5  and  FIG. 6 , the thermodynamic cycle for the preferred embodiment of present  FIG. 1  engine is represented by the paths that travel through state points  1 ,  2 A,  2 B,  3 A and  4 A. The compression of cylinder  13  occurs between points  1  and  2 A. The optional cooling of compressed air from cylinder  12  in compressed air conduit  50  occurs between points  2 A and  2 B. Without this optional cooling, the process would proceed from point  2 A directly to point  3 A. Note that in  FIG. 6 , state point  2 B is at a temperature that is below the fuel ignition temperature. This permits spark controlled ignition as opposed to auto-ignition in an engine which uses a fuel adapted for spark ignition. Even though this cooling below the auto-ignition temperature results in a small energy loss, much of the thermodynamic benefit of the additional compression is retained. This additional compression corresponds to the paths between points  2  to  2 A in  FIGS. 5 and 6 .  
         [0058]     For example, the state points  1 ,  2 ,  3  and  4  described above for a typical Otto cycle engine could be given as follows as shown in the chart below  
                                                                                     Pressure (P)   Sp. Vol. (v)   Temperature       Point   Description   (MPa)   (m 3 /kg)   T (°K)                                1   Start of   0.100   MPa   0.829 m 3 /kg   289°   K           Compression   (14.4   psia)       (60°   F.)       2   End of Compression   1.825   MPa   0.104 m 3 /kg   663°   K       3   End of Combustion   8.739   MPa   0.104 m 3 /kg   3175°   K       4   End of Exhaust   0.475   Mpa   0.829 m 3 /kg   1382°   K               (69.0   psia)       (2028°   F.)                  
 
         [0059]     The above chart describes an example process featuring a typical 8:1 compression ratio where the heat added is 1900 KJ/Kg, heat loss is 783 KJ /Kg and the useful work is 1017 KJ/Kg. This yields a thermodynamic efficiency of 56.5%.  
         [0060]     In contrast, state points  1 ,  2 ,  2 A,  2 B,  3 A and  4 A shown in  FIGS. 5 and 6 , could, for example, be described by the second chart below:  
                                                                                     Pressure (P)   Sp. Vol. (v)   Temperature       Point   Description   (MPa)   (m 3 /kg)   T (°K)                                1    Start of   0.100   MPa   0.829 m 3 /kg   289°   K           Compression   (14.4   psia)       (60°   F.)       2A   End of Compression   6.597   MPa   0.041 m 3 /kg   957°   K       2B   Intercooler Exit   4.569   Mpa   0.041 m 3 /kg   663°   K       3A   End of Combustion   21.878   MPa   0.041 m 3 /kg   3175°   K       4A   End of Exhaust   0.329   Mpa   0.829 m 3 /kg   957°   K               (48.0   psia)       (1264°   F.)                  
 
 The above chart describes an example process which traces points  1 ,  2 A,  2 B,  3 A and  4 A shown in  FIGS. 5 and 6 . This modified process features an enhanced 20:1 compression ratio achievable with the present engine. In this high compression process, the heat added is 1800 KJ/Kg, heat loss is 690 KJ/Kg and the useful work is 1010 KJ/Kg. This yields a theoretical thermodynamic efficiency of 61.7% which is significantly greater than the theoretical 56.5% thermodynamic efficiency of the process given above having a typical 8:1 compression ratio. 
 
         [0061]     Accordingly, presented here is an engine having a means for controlling the pressure and temperature of compressed air in an Otto cycle and a means for controlling the injection of compressed air into a combustion cylinder generally during the second half of a piston return stroke so that higher thermodynamic efficiencies or power densities may be achieved.  
         [0062]     It is to be understood that while certain forms of this invention have been illustrated and described, it is not limited thereto, except in so far as such limitations are included in the following claims and allowable equivalents thereof.