Abstract:
An internal combustion engine having a cylindrical outer housing with inner cylindrical cylinders defining power cylinders circumferentially spaced in the engine. Each of the cylinder(s) has opposed intake and exhaust pistons. The intake and exhaust pistons use air bearings instead of conventional oil as a lubricant. The intake and exhaust pistons power a rotary cam mounted on opposed ends of the engine and a cam follower system positions the pistons. The rotary cam is operatively connected to a drive shaft or power take off shaft. A four bar linkage system is connected to the piston rod to minimize piston side loads. A jumper connects the ends of the power cylinders to pressurize air drawn into the engine and to admit heated air, indirectly heated during the combustion cycle, to charge the power cylinder for improved efficiency and to purge the pressure cylinder after combustion. The overall construction of the engine allows the engine to be made from relatively light weight materials providing a significantly improved power to weight ratio over conventional engines. The overall construction also allows for reduced fuel emissions and greater fuel efficiency.

Description:
TECHNICAL FIELD  
       [0001]     This invention relates to reciprocating internal combustion engines and particularly to an advanced version that eliminates side loadings, utilizes thermally controlled power cylinders, opposed intake and exhaust pistons, piston rings that are cooled and hydrostatically lubricated by air, and incorporates a high temperature cylinder wall which reduces engine emissions and increases engine performance.  
       BACKGROUND  
       [0002]     As is well known, diesel, gas and steam engines of the reciprocating type typically convert the linear piston motion into rotary motion by utilizing piston(s), connecting rod, and crankshaft. This conversion process obviously creates a substantial piston side load which requires oil lubrication to control friction and wear of the piston skirt and cylinder and a substantial and heavy engine case. To prevent oil breakdown and loss of lubricity the cylinder wall and piston side walls and rings generally are maintained at a temperature that is below a maximum of 350 degrees Fahrenheit. Typically, these engines must incorporate a cooling system that serves to reject at least 25 percent of the total heat energy which is dissipated into the ambient air which energy would otherwise provide shaft horsepower.  
         [0003]     As will be described in more detail hereinbelow, the engine of the present invention, unlike what is shown in the prior art, floats the piston in the cylinder with a cushion of air by absorbing the side loads that would otherwise load the pistons at locations remote from the piston. Unique to the engine of this invention is the use of air feed tubes made from a compliant material that keep the piston ring concentric to the piston and supply air to the integral piston ring depressions to hydrostatically compress the piston ring relative to the cylinder and continuously float the piston and piston ring on pockets of pressurized air. The engine of the present invention also uses a unique bearing pack connected to a four bar linkage arrangement and cam to transmit power and reduce side loads to the piston. The engine of the present invention also uses a unique jumper system for the purpose of storing base compression air which is pressurized for use in purging the combustion chamber of combusted materials and supply preheated air to the combustion chamber prior to combustion. The engine of the present invention also has a unique power cylinder that distributes heat of combustion along the length of the cylinder to allow higher operating temperatures in the power cylinder.  
         [0004]     One of the inventors of the present invention is the inventor of U.S. Pat. No. 5,551,383. This patent discloses the use of an air bearing system that relies upon pressurized air provided by a base compression cylinder that is co-annular to the power cylinder. The &#39;383 patent also employs a four bar linkage system, but the disclosed system is rather complex to replicate in a working engine and doesn&#39;t provide the benefits of the unique four bar system of the present invention. The &#39;383 patent also doesn&#39;t employ a power cylinder that manages heat to allow higher temperatures to be used in the combustion chamber and power cylinder as does the engine of the present invention. Other improvements are disclosed and claimed in the present application that are not taught or suggested in the &#39;383 patent which are patentable over the &#39;383 patent when considered either individually or in combination with other related technologies.  
         [0005]     U.S. Pat. No. 5,375,567 granted to A. Lowi, Jr. on Dec. 27, 1994, discloses a two-stroke-cycle engine that requires no cooling and utilizes twin double-harmonic cams that claim to balance reciprocating and rotary motion at all loads and speeds so as to obviate all side loads. As will be more fully detailed hereinbelow, the present invention makes no claim to the ability of operating without lubrication, Although the engine of the present invention does not require oil as a lubricant for the pistons as is the case for most piston engines it does use air as a lubircant. It also utilizes a four bar linkage system to reduce side loads. Still further, the present invention employs unique seals to seal and absorb the slight side loads that may be encountered by the pistons of the engine.  
         [0006]     Other patents that utilize opposing pistons and harmonic types of cams but do not incorporate a linkage system for minimizing or eliminating side loads are U.S. Pat. No. 2,076,334 granted to E. B. Burns on Apr. 6, 1937, and U.S. Pat. No. 1,788,140 granted to L. M. Woolson on Jan. 6, 1931.  
         [0007]     Also disclosed in the prior art are a number of patents that utilize a gas for lubrication rather than oil. For example, U.S. Pat. No. 4,455,974 granted to Shapiro et al on Jun. 26, 1984, utilizes gases generated in the engine to hydrostatically support the piston rings. Similarly, U.S. Pat. No. 4,681,326 granted to I. Kubo on Jul. 21, 1987, utilizes engine gasses to support the piston rings. U.S. Pat. No. 4,111,104 granted to Davison, Jr. on Sep. 5, 1978, utilizes engine gases to support the piston and U.S. Pat. No. 3,777,722 granted to K. W. Lenger on Dec. 11, 1973, discloses a ringless piston with air for reducing friction.  
       BRIEF DESCRIPTION OF THE INVENTION  
       [0008]     The present invention provides an improved internal combustion engine that has a low weight to power ratio, low emissions and low fuel consumption. The engine of the present invention provides numerous improvements over known internal combustion engines.  
         [0009]     The present invention includes an internal combustion engine having a housing enclosing at least one cylinder with opposed pistons mounted for reciprocation within the cylinder. Opposed power cams are mounted upon a power output shaft. Each of the power cams are operatively connected to a respective one of the opposed pistons. End plates are mounted to the housing and divide the housing into a center section and end sections. At least one cylinder with opposed pistons mounted for reciprocation within the cylinder is mounted within the center section and one each of the opposed power cams is mounted within a respective one of the end sections. The opposed pistons have connecting rods operatively interconnecting the pistons to the cams. The end plates have openings for reciprocal receipt of the connection rods. Unique seals are provided to seal the end sections from the center sections and to seal the cylinders to the end plates. The entire engine of the preferred embodiment is designed to maintain all spring rates at acceptable rates to avoid any inadvertent weakening of connections in the engine.  
         [0010]     The power cylinder of the present invention is uniquely designed to distribute heat along the length of the power cylinder to avoid tapering of the cylinder. The power cylinder has a first member defined by a first hollow tube having a predetermined length. The first hollow tube is adapted to receive at least one piston for reciprocal movement within the first hollow tube. The first hollow tube defines a combustion chamber wherein a fuel and air mixture can be introduced compressed and ignited. This first hollow tube has a high thermal expansion coefficient and low conductivity.  
         [0011]     A second member is mounted adjacent the first member. The second member has a high thermal expansion coefficient and high conductivity.  
         [0012]     A third member is positioned about the first hollow tube adjacent the combustion chamber of the first hollow tube. The third member has a low thermal expansion coefficient and low conductivity.  
         [0013]     The first, second and third members interact to reduce tapering of the first member by initially containing through the third member heat within the combustion chamber and reducing expansion of the combustion chamber and thereafter distributing heat developed in the combustion chamber along the first member by directing the heat along the second member to maintain a generally uniform temperature along the length of the first member.  
         [0014]     The engine of the present invention also uses a bearing pack that transfers the power generated by the combustion process to a pair of power cams. The bearing pack has a housing, the housing has a top surface and opposed legs extending from the top surface. The legs have facing inner surfaces and outer surfaces. A pair of axles extend out of the facing inner surfaces and a pair of pins extending out of the outer surfaces, the axles and pins are coaxial.  
         [0015]     A four bar linkage assembly is connect to the pins and guide wheels are connected to the axles. The guide wheels have a first wheel and a second wheel, the first wheel having a larger diameter than the second wheel. The opposed cams have opposed spaced apart tracks. The first wheel engages one of the tracks and the second wheel engages the other of the tracks.  
         [0016]     The pistons of the present invention are also uniquely designed to include facing combustion surfaces with outer perimeters, the outer perimeters of the combustion surfaces each having a profiled surface, the profiled surfaces mate to form a combustion chamber between the piston combustion surfaces. This allows the combustion surface to be generally closed to reduce heat loss from the combustion chamber during combustion.  
         [0017]     These and other characteristics of the present invention with its various alternatives and embodiments can be better understood with reference to the following detailed description of the invention when read in conjunction with the accompanying drawings, wherein like reference characters refer to like parts throughout the several views. 
     
    
     DESCRIPTION OF THE DRAWINGS  
       [0018]      FIG. 1  is a perspective view of the engine of the present invention.  
         [0019]      FIG. 2  is a side view of the engine of the present invention with the housing and the power cylinder shown in dotted lines.  
         [0020]      FIG. 3  is a perspective view of the power cam of the present invention.  
         [0021]      FIG. 4  is a partial perspective view of the bearing pack of the present invention.  
         [0022]      FIG. 5  is a partial cutaway view of the power cam showing the wheels of the bearing pack engaging the opposed walls of the power cam.  
         [0023]      FIG. 6  is an exploded view of the wheels of the bearing pack of the present invention.  
         [0024]      FIG. 7  is a perspective view of the base compression cylinder with the power cylinder shown in phantom.  
         [0025]      FIG. 7A  is a perspective view of the power cylinder.  
         [0026]      FIG. 7B  is a perspective view of the base compression cylinder mounted to the inlet manifold through inlet conduits.  
         [0027]      FIG. 7C  is a cutaway view illustrating the connection of a power cylinder and base compression manifold to the end plate.  
         [0028]      FIG. 7D  is perspective view of a low radio load of the circumferential spring of the present invention.  
         [0029]      FIG. 7E  is a perspective view of the ring of  FIG. 7D .  
         [0030]      FIGS. 8A-8F  are schematic views showing the operational cycle of the engine of the present invention.  
         [0031]      FIG. 9  is a perspective view of the inplate of the present invention.  
         [0032]      FIG. 10  is a cutaway view of the power cylinder showing the pistons at top dead center.  
         [0033]      FIG. 11  is an exploded view of the piston of the present invention.  
         [0034]      FIG. 12  is a partial cutaway view of the piston showing the hydrostatic bearing of the present invention.  
         [0035]      FIG. 13  is a partial perspective view of the piston ring of the present invention.  
         [0036]      FIG. 14  is a cutaway view of the piston rod seal pack.  
         [0037]      FIG. 15  is a partial perspective view of the piston rod seal pack.  
         [0038]      FIG. 16  is a cutaway view of the piston rod seal pack of the present invention.  
         [0039]      FIG. 17  is a partial cutaway view of the engine showing the R supply line connector.  
         [0040]      FIG. 18  is a partial cutaway view of the present invention showing the oil slinger of the present invention.  
         [0041]      FIG. 19  is a side view of the mounting bars and fore bar linkage of the present invention.  
         [0042]      FIG. 20  is a cutaway view of the bearing pack and cam of the present invention.  
         [0043]      FIG. 21  is a cutaway view of the engine showing two power cylinders with the rotary shaft extending between the power cylinders.  
         [0044]      FIG. 22  is a perspective view of the air bearing piston ring of the present invention.  
         [0045]      FIG. 23  is a cutaway view of the air bearing piston ring of the present invention.  
         [0046]      FIG. 24  is a sectional view through the air inlet of the air bearing piston ring.  
         [0047]      FIG. 25  is a section through the bypass slot of the air bearing piston ring of the present invention.  
         [0048]      FIG. 26  is a schematic showing of the engine of the present invention. 
     
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0049]     In its preferred embodiment, the engine of the present invention as described herein is configured with four (4) cylinders and eight (8) pistons and each paired diametrically opposed piston sets are compressing and expanding axi-symmetrically, so as to minimize or eliminate unbalance or out of plane loads at any time during the engines operating envelope for providing a relatively vibration free engine. Since each piston set “fires” twice per output shaft revolution, it produces twice the torque at half the shaft RPM. While this invention is described in the preferred embodiment to include specific parameters, it will be appreciated by one skilled in this art that other parameters including the number of pistons and attendant cylinders could be used without departing from the scope of this invention. It will be appreciated that two opposing pistons in a single cylinder will constitute the minimum number of pistons and cylinders, it will also be appreciated that more or less than four (4) cylinders and eight (8) pistons could be used.  
         [0050]      FIG. 1  is a perspective view of the engine of the present invention generally indicated by reference numeral  10  which is comprised of a modular cylindrical engine outer case assembly  12  surrounding the rotary shaft  14  for rotation about the engine&#39;s axis A. Shaft  14  extends outwardly from the fore end  16  and the aft end  18 . Surrounding the engine outer case  12  is an inlet manifold  20  and exhaust pipes  22  which are in communication with the intake pistons and exhaust pistons. Inlet conduits  24  connect the intake pistons with the manifold  20 . In the disclosed embodiment, the exhaust pipes are directly connected to ports adjacent the exhaust pistons. As will be described in greater detail herein below, the inlet port  80  disposed in the inlet manifold  20 , which may include a suitable filter, leads fresh ambient air into the piston cylinders. The exhaust ports disposed adjacent the exhaust pipes  22  discharge the spent combusted products to ambient.  
         [0051]     Fuel is admitted to the cylinders through the fuel nozzle injectors  30  which is fed fuel under pressure through a fuel line, which is not shown. Fuel from a fuel reservoir is pressurized in a well known manner from suitable injector pump(s). In the preferred embodiment the accessories would be powered by the portion of rotary shaft  14  that extends from the fore end  16  and the power for driving the load would be extracted from the shaft extending from the aft end  18 . This is, of course, optional as the power for either the accessories or load may be extracted at either end of rotary shaft  14 . It will be understood that the load that the engine drives would include without limitation, passenger cars, land vehicles, aircraft and water vehicle propellers, auxiliary power units, generators, earth moving vehicles and the like.  
         [0052]     With reference to  FIG. 2 , in the engine as illustrated, there are four equally and circumferentially spaced power cylinder assemblies  36 , these are shown by dotted lines in  FIG. 2 . The construction of the power cylinder assemblies  36  will be discussed in greater detail below. The power cylinder assembly  36  supports (8) pistons, namely four (4) intake pistons  38  opposing (4) exhaust pistons  40 .  
         [0053]     Rotary shaft  14  connects to and is rotated by the opposing power cams  46  and  48  ( FIGS. 2, 3 ). In the preferred embodiment, the cams are made of steel and aluminum, but other materials could also be used. The cams  46  and  48  are located concentrically and axially within the engine case  12 . Power cams  48  and  46  respectively are driven by the intake pistons  38  and exhaust pistons  40  respectively via the connecting rods  58  and  56 , that are operatively connected to respective bearing packs  60 . With reference to  FIG. 4 , the bearing pack  60  includes a housing  62  which is generally U-shaped having side walls  64  and  66 , legs  65  and a top  68 . In the disclosed embodiment, the top  68  and legs  65  are one piece. The connecting rods  56  and  58  are connected to a respective bearing pack  60  by a machine bolt  67  passing through the top  68  into a threaded nut  69  in rods  56  and  58 . In the preferred embodiment, the spring factor of the connection is predetermined to be greater than 3. Still further, in all connections throughout the engine  10  of the present invention, a spring rate of greater than 3 is desired. The spring rate is determined from the following equation:  
       K   =     AE   L         
        A=area of the respective parts     E=Modulas of elasticity     L=Length of the respective parts.        
 
         [0057]     By way of example, if a bolt and nut are used to connect a stack of material, the spring factor would be determined by the ratio of the spring rate of the stack of the bolt as follows:  
           K   S       K   B       &gt;     4   ⁢           ⁢       (         A   S     ⁢     E   S         L   S       )       (         A   B     ⁢     E   B         L   B       )         &gt;     4   1         
 
         [0058]     The side walls  64  and  66  each receive dual roller bearings  70  and  72  ( FIG. 5 ). Bearing  70  is a power bearing and bearing  72  is a retractor bearing. Bearings  70  and  72  are mounted coaxially on coaxial pins  73  protruding from the inside of each side wall  64  and  66 . Pins  176  protrude from the outside of each wall  64  and  66  for connection to the bars of a four bar linkage which will be discussed in greater detail below. The pins  176  are coaxial with pins  73 . In this way, a minimal bending moment is created in the operation of bearing pack  60 .  
         [0059]     Each of the bearing packs  60  are operatively mounted to the power cams  46  and  48  as shown in  FIGS. 5 and 20 . The power cams  46  and  48  have opposed facing tracks  75  and  77  upon which bearings  70  and  72  ride. In the preferred embodiment, the walls  64  or  66  are removably attached to the legs  65  and top  68  by bolts  79 . In this way, the side walls with bearings  70  and  72  attached can be assembled with the bearing  70  and  72  between and engaging tracks  75  and  77 .  
         [0060]     In the preferred embodiment, the bearings are preloaded by a retractor preload spring assembly  81  shown in  FIG. 6 . As illustrated, the power bearing  70  is mounted upon pin  73  and retractor bearing  72  is mounted upon pin  73  through preload spring  81  and retractor carrier  82 . Due to the slots  87  in spring  81 , the bearing  72  is biased with respect to pin  73  and bearing  70 . The distance between facing tracks  75  and  77  is slightly less than the combined radii of bearings  70  and  72  ( FIG. 5 ). In this way, when installed, the bearings are slightly pre-loaded. This insures contact at all times and insures tolerance take-up. Additionally, the spring  81  permits movement between bearings  70  and  72  without binding the bearings  70  and  72  with respect to tracks  75  and  77 .  
         [0061]     The roller bearings  70  are forced against and roll on tracks  77  of power cams  46  and  48  to cause them to rotate around axis A when the combustion in the power cylinder pushes the intake piston  38  and exhaust piston  40  apart ( FIGS. 1 and 2 ). The roller bearings  72  act as idlers to maintain bearing contact with tracks  75  and  77 , and the roller bearings  72  also assist in rotating the power cams around axis A when the momentum vector of each piston assembly is larger (and towards the top of the piston) than the force of acceleration being imparted to the piston assembly by the power cam. Obviously, the opposite may occur under other special circumstances. Roller bearings  70  drive the cams  46  and  48  during the power stroke of pistons  38  and  40 . The small roller bearings or retraction bearings  72  roll on the tracks  75  of the power cams  46  and  48  respectively to actuate the intake piston  38  and exhaust piston  40  to assist in pulling the intake piston  38  and exhaust piston  40  to the end of the bottom dead center of the stroke. The intake piston  38  and exhaust piston  40  are then pushed together by the large bearings or roller bearings  70  rolling on the tracks  77  (i.e. when the acceleration force vector towards the piston top is larger than the momentum vector which is opposite in direction). Under certain conditions the large roller bearings  70  may have sufficient energy to position the intake piston  38  and the exhaust piston  40  the full travel of the stroke. In other conditions the small roller bearings  72  may have to assist to position the intake and exhaust pistons to bottom dead center. The track  77  of the power cams  46  and  48  are suitably contoured to a slightly larger radius than the large bearings  70  so that the bearing outer race will hydroplane on the cam surface, minimize the contact footprint, minimize the differential velocity of individual segments within the contact footprint as the bearing follows the radially aligned tracks, provide a built in cushioning, and prevent metal to metal contact.  
         [0062]     The bearing packs are operatively supported by a four bar linkage assembly  172 ( FIG. 2 ). The assembly  172  has four bars connected between the mounting bars  174  on plate  170  and a bearing  178  ( FIG. 2 ), mounted to pin  176  ( FIG. 4 ). As should be appreciated, the bearing  178  is coaxial with pin  73  ( FIG. 4 ) and bearings  70  and  72  ( FIG. 4 ). In this way, there is no bending moment in the pistons  38  and  40 . The four bar linkage  172  allows the linear movement of pistons  38  and  40  to be translated into rotational movement of the cams  46  and  48  with none of the reactive forces caused by the pistons driving the cams being translated through the pistons  38  and  40  to the cylinder wall  98 .  
         [0063]     The four bar linkage assembly  172  is further illustrated in  FIGS. 19 and 20 . The four bar linkage assembly  172  of the present invention guides the pistons  38  and  40  and reacts the side load of the cams  46  and  48  to the link mount points I at  180  and II at  182 . Since the center coupler link or bearing  178  has a revolute center, none of the cams  46  and  48  side load is transmitted to the piston  38  and  40  and piston ring  130 . There is a precise relationship between the geometry of the links  184  to give the desired straight line motion of the piston ring  130  and pistons  38  and  40 . The relationship is the combination of the lengths of the four bar linkage assembly  172  components and the location of the mounting points  180  and  182 .  
         [0064]     By way of example, in one engine design, the pistons  38  and  40  have a stroke of 2.0 inches. To allow for tolerances and possible travel outside of the design range, an additional 0.1 inches was added to each end of the stroke. The lengths of the components were selected for a stroke of 2.2 inches. The linkage for the engine would have the following dimensions:  
                                                       Coupler link 178 Stroke:   2.20 inches           Mount Point 180:   X = 1.7726, Y = 0.880           Mount Point 182:   X = −1.7726, Y = −0.880           Link Length 184:   1.990 inches           Coupler Link 178 Length   1.770 inches                      
 
         [0065]     The ratio of these components must be maintained as the engine stroke is scaled up or down. Each component length is scaled linearly with the change in stroke. If the stroke is doubled, all the values must be doubled. To reduce the stresses on the components, all links are in double shear.  
         [0066]     The engine&#39;s operating cycle is best illustrated by the schematic drawings of  FIGS. 8A-8F  where  FIG. 8A  illustrates top dead center.  FIG. 8B  illustrates the power stroke cycle,  FIG. 8C  is exhausting the power cylinder,  8 D- 8 E is charging the base compression and purging the power cylinder and  8 F illustrates the compression stroke cycle.  
         [0067]     As shown in  FIG. 8A  the intake  38  and exhaust  40  pistons are located at the top dead center of their strokes and intake piston  38  and exhaust piston  40  are at the end of the compression stroke and in the power stroke and positioned as close to each other for correct compression ratio. As is apparent from the foregoing, the air in the working portion or compression chamber  29  of the power cylinder (the volume between intake and exhaust pistons) is fully compressed and fuel is timely introduced through the fuel nozzle injector  30  to cause an explosion forcing the pistons to separate. At this point of the cycle the inlet check valve  85  is open since the air on the upstream and downstream sides of the check valve  85  is basically at the same pressure or the pressure upstream is still slightly greater than the intermediate inlets  86  and  88  and the transfer tube intake or jumper  84 . To maintain the sealability of the chamber represented by the contiguous volume under the pistons  38  and  40  (including inlets  88 ,  86 , and intake jumpers  84 ), an o-ring seal  83  is used. This seal also accommodates the thermal differentials between the cylinder  36  and the plate  170 . Also the pressures on the back sides of intake piston  38  and exhaust piston  40  are similar since they are in fluid communication with inlet port  80  via the intake jumpers  84  and the inlet passages  86  and  88 . The exhaust port  90  is closed off by exhaust piston  40 .  
         [0068]     With reference to  FIG. 21  in the preferred embodiment, the power cylinder  36  is supported radially by a low spring rate o-ring seal  83  and is constrained in the engine by features on the endplate. These features maintain concentricity with the connecting rods  56  and  58  and rod seal hole  190  in the endplate  170 . The axial location of the cylinder can be adjusted by a shim  251  to account for the tolerance of the machined components. The intake and exhaust endplates  170  are identical so the shim  251  is also used to fill the area on the intake side in the location where the spring  250  resides on the exhaust endplate  170 .  
         [0069]     The power cylinder  36  is loaded against the intake endplate  170  by the spring  250  that reacts on the other end to the exhaust endplate  170 . The spring  250  (currently a wave spring) has sufficient spring force to overcome an inadvertant rub of the piston ring  130  or piston skirt on the power cylinder inner wall  98 .  
         [0070]     Referring next to  FIG. 8B , as both pistons are translating back toward the dead end of the stroke, i.e. bottom dead center, the check valve  85  is closed so that gas cannot flow towards inlet port  80  and the pressures behind intake piston  38  and exhaust piston  40  increase. The pressure of the combusted products between pistons  38  and  40  in compression chamber  29  of the power cylinder assembly  36  decreases. The exhaust port  90  remains closed at this point of the cycle.  
         [0071]     Referring next to  FIG. 8C , the pistons are still moving apart and travelling toward bottom dead center and the exhaust ports  90  are opening and inlet passages  86  are blocked from communicating with the compression chamber  29  and are closed off by the check valve  85  so that air cannot flow towards the inlet  80 . Consequently, pressure of the intake air continues to build under the piston. The exhaust gases are leaving the compression chamber. At this point of the cycle the pressure of fluid in the assembly compression chamber  29  of the power cylinder assembly  36  is quickly reducing towards its lowest value.  
         [0072]     At the bottom dead center of the stroke as seen in  FIG. 8D , the exhaust ports  90  are fully opened and the inlet passages  86  and  88  are in full communication with the compression chamber. This scavenges or purges the compression chamber  29  of power cylinder assembly  36  by allowing air trapped in the intake jumper  84  under a pressure higher than the compression chamber  29  to fill the power cylinder. It will be appreciated that prior to charging the power cylinder assembly  36 , the air captured in jumper  84  is preheated by being in indirect heat exchange with the combustion products during the combustion process with a consequential increase in engine efficiency, because this allows the piston top to run at higher temperatures which allows high power density levels, and the thermal tapering of the cylinder is decreased. However, this percentage must be managed, too much is a detriment to overall engine efficiency.  
         [0073]     It will be noted that in  FIGS. 8B and 8C  the air trapped behind the intake piston  38  and exhaust piston  40  is blocked from the compression chamber  29  by the intake piston  38  and that the intake jumpers  84  are only in communication with the volumes behind the intake and exhaust pistons  38  and  40 . This air is completely trapped while the pistons are still in their power stroke. Hence, the power stoke further compresses this air. Since the pistons are close to the end of their stroke during the remaining portion of the power stroke as is viewed in the schematics depicted in  FIGS. 8B and 8C  and  8 D the movement of the intake and exhaust pistons creates a very high pressure of remaining base compression air.  
         [0074]      FIGS. 8E and 8F  depict the compression cycle where the pistons are actuated by the power cams  46  and  48  toward top dead center which is the transition point of the power stroke ( FIG. 8A ). As the intake piston  38  and exhaust piston  40  move toward each other and pass over the inlets  86  and  88  and exhaust ports  90 , the air trapped in the compression chamber  29  of the power cylinder compresses which causes the pressure to increase until it reaches the maximum value at the end of the stroke (top dead center)  FIG. 8A . Once the pistons cross over the inlets  86  &amp;  88 , the back ends of the intake pistons  38  and exhausts piston  40  remain open to the inlet pressure and since the back pressure of the check valves  85  equals the ambient pressure and is greater than the pressure below the pistons  38  and  40 , these check valves remain open and the back ends of the intake piston  38  and exhaust piston  40  draw in ambient air.  
         [0075]     It should be appreciated by those of ordinary skill in the art that by changing the shape of the power cams  46  and  48 , the engine&#39;s characteristics can be changed. For example, by adjusting the length of the flats in the power cams  46  and  48 , the acceleration, velocity, emissions, power, etc., can be altered.  
         [0076]     It will be appreciated from the foregoing that engine  10 , does not require valving, such as the poppet type valves used for opening and closing the intake and exhaust ports inasmuch as these ports in this engine are opened and closed by virtue of the intake and exhaust pistons.  
         [0077]     With reference to  FIG. 7 , the base compression cylinder assembly is generally shown at  95  as partially surrounding cylinder  36  and includes a base compression cylinder manifold  92  which is mounted through tie bolts  93  and nuts  91  to an air intake manifold  89 . The tie bolts  93  extend through the intake jumpers  84 . The power cylinder assembly  36  is mounted within the manifold openings  94 . The power cylinder assembly  36  is shown in  FIG. 7   a  and will be discussed in greater detail below. In the disclosed embodiment, the intake check valves  85  ( FIGS. 8A, 8B ,  8 C,  8 D;  8 E) are reed valves which mount in inlets  80  of the air intake manifold  89 . The tie bolts  93  and nuts  91  connect the power cylinder assembly  36  and the base compression assembly  95  together into a single package. Each of these packages are then mounted between the plates  170  ( FIG. 9 ). The plates  170  are then bolted to the outer case  12 .  
         [0078]     With reference to  FIG. 7   c , the preferred method for mounting the power cylinder assembly  36  within the manifolds  89  and  92  will be discussed. The openings  94  of the manifolds  89  and  92  have slots  106  and  108  for receipt of O-rings  110 . Positioned between the manifolds  89  and  92  and the power cylinder assembly  36  are low radial load circumferential springs  112  and  114 . As illustrated, spring  112  is preferably a straight spring and spring  114  is preferably a V-shaped spring. The V-shaped spring  114  is shown in more detail in  FIG. 7   d . Spring  114  is preferably made from an annular inner sleeve  115  brazed to an annular outer sleeve  117  at one end  119  thereof. Both springs  112  and  114  are preferably made of low thermal conductivity material. In  FIG. 7C , the springs  112  and  114  are pinned by pins  116  to the cylinder assembly  36  and more particularly to second tube  102 . The pins  116  are preferably press fit into openings in the power cylinder assembly  36  and more particularly second tube  102 . With respect to spring  112  an opening is also provided in the manifold for receipt of pin  116 . The use of springs  112  and  114  in conjunction with O-rings  110  provides a proper seal, but also allows for balancing of the spring rates between the coupled elements. A further O-ring  110  is mounted between plate  170  and power cylinder assembly  36  to complete the seal. A shim is shown in this figure.  
         [0079]     With reference to  FIG. 7   b , the interconnection of the inlet manifold  20  or air collector through the inlet conduits  24  to the air intake manifold  89  is illustrated.  
         [0080]     With reference to  FIGS. 7A and 10 , power cylinder assembly  36  will be discussed. The power cylinder assembly  36  is constructed of three different materials to both block and facilitate heat transfer and in particular to maintain the cylinder at a fairly consistent and average temperature along its entire length. It is critical that the power cylinder temperature be maintained as close to uniform as possible to prevent distortion of the power cylinder assembly  36 . If, as is typical of standard internal combustion engines, the cylinder temperatures are not maintained, the diameter of the cylinder  36  could vary, creating difficulties with wear, piston gas bypass, loss of efficiency; and in the present engine, difficulties with the air bearings which will be described in greater detail below. Typical internal combustion engine cylinders expand in hotter regions creating a coning or tapering effect. By maintaining the cylinder temperature within a narrower range along its length, there is no appreciable coning or tapering effect.  
         [0081]     As will be appreciated, the temperature of the cylinder  36  is going to be higher than that found in typical internal combustion engines. This does not create a problem for this engine for several reasons. The cylinder  36  doesn&#39;t contain oil as a lubricant, air is used, the thermal mass and thermal conductivity are lower, and the frictional heating created by piston side loads and oil shearing is not an issue with this engine. Therefore, higher temperatures are acceptable. The key is to apportion the heat generated by combustion along the length of the cylinder  36  in a generally uniform manner.  
         [0082]     Heat apportionment is achieved by using three separate materials in cylinder  36 . The first material is steel, preferably A286. In the disclosed embodiment a steel tube  100  is used. As illustrated in  FIG. 7A , inlets  86  are formed in tube  100 . This material has a high alpha (thermal expansion coefficient) and low conductivity. Mounted about tube  100  is a second tube  102  of material made from copper. It should be appreciated the tube  102  does not have to be a tube but could be strips of material etc. Tube  102  is in two parts and is mounted adjacent the combustion chamber forming a gap between the two parts at the combustion chamber. Copper has a high alpha and high conductivity. The third material is a stainless steel sleeve  104  mounted about the combustion chamber and partially covering tube  102 . This sleeve  104  has a low alpha and very low conductivity.  
         [0083]     With this construction, heat generated within the combustion chamber is predominantly blocked by the sleeve  104 . A majority of any heat absorbed into sleeve  104  is directed down the copper tube  102 . The copper tube  102  being highly conductive, quickly normalizes the temperatures between any heat transferred from the low conductivity sleeve  104  and the cold end of the cylinder  100  which is farthest away from the combustion zone. Thus, this construction minimizes tapering caused by thermal stresses. The stainless steel sleeve  104  reduces if not eliminates tapering at the combustion chamber because of its very high alpha and resists heat because of its low conductivity. In this way, tapering is reduced or eliminated and balanced between the different thermal regions along the cylinder  36  and at the combustion chamber heat is maintained within the combustion chamber during the initial stages of the power stroke where it can provide useful work. It should also be noted that the comparative alphas are critical so that the sleeve to sleeve mechanical contact is maintained at various temperatures and thermal gradients. In this preferred embodiment, the mechanical contact is primarily the result of thermal assembly with a minor role played by brazing. Obvious, to one skilled in the art this same result could be obtained using other mechanically constrained interfaces or braze/solders.  
         [0084]     With reference to  FIGS. 11 and 13 , the piston assemblies  120  will be described. In the preferred embodiment, the piston assemblies  120  include a piston top  122 , a ring pack top  124 , ring pack bottom  126 , a piston base  128 , and a piston ring  130 . The various components which make up the piston assemblies  120  are bolted together by bolts (not shown) that are inserted into protrusions  132  extending from piston top  122  through protrusions  134  extending from piston base  128 . The protrusions  132  extend through openings  136  on ring pack top  124  and bottom  126 . The protrusions  132  and  134  and the bolts once again allow the spring rates to be managed.  
         [0085]     The piston tops  122  are specially constructed to reduce the radiant heat effect against the cylinder walls  98 . As illustrated in  FIG. 12 , the face  42  of the exhaust piston  40  and the face  44  of the intake piston  36  have overlapping surfaces  142  and  144  about the outer perimeter of the faces.  
         [0086]     These overlapping surfaces  142  and  144  shield the cylinder wall from radiant heat generated during combustion and traps the radiant heat in the combustion chamber. It should be appreciated that the overlapping surfaces could have a different shape. For example, the protruding portion  142  could be on intake piston  38  and portion  144  could be on piston  40 .  
         [0087]     To further manage combustion in engine  10 , a fuel injection cavity  164  is provided in the piston top. The cavity  164  is longer than it is wide to direct fuel along the path for complete combustion and to insure that no liquid fuel contacts a solid surface prior to combustion.  
         [0088]     Continuing with reference to  FIGS. 11 and 12 , the piston ring  130  of the present invention is preferably a split piston ring with hydrostatic lift pockets  244 . Pressurized air flows to the circumferentially spaced pockets  244  formed in each of the split rings via small diameter flex tubes  148 . The flex tubes  148  are generally U-shaped to allow for a defined spring rate curve. Flex tubes  148  are plugged into a central air manifold  150  which receives pressurized air which is piped from the piston base  128 . The pressurized air flows from the connecting rods  56  and  58  and into the openings in manifold  150 . See  FIG. 13 . A retaining wall  154  receives the free ends of flex tubes  148  which are in turn fed through the seal  158 , and then plugged into ring  130 . See  FIGS. 11, 12  and  13 . The retaining wall  154  has a plurality of slots  156  for receipt of tubes  148  and seals  158 . The slots  156  allow for lateral movement of tubes  148 . The seals  158  have a face plate  160  and relatively long body  162 . The body  162  receives the tube  148  and fits in slot  156  and face plate  160  engages wall  154 . The long body  162  acts as an effective seal within slot  156  and against the tube  148 . The face plate seals the opening in slot  156 . This combination of sealing methods seals combustion gasses out of the cavity in pistons  38  and  40 .  
         [0089]     The air supplied by the tubes  148  flows into the lift pockets  244  in the ring  130  to lift the pistons  38  and  40  off the cylinder wall. The pockets  244  are equally spaced or arranged for optimum positioning around the circumference of the piston rings so that the air admitted compresses the piston ring relative to the cylinder and additionally locates the piston. Each of the tubes  148  are bent in a generally U-shaped configuration and since one end is affixed to the piston and the other end is affixed to the piston ring, the pressure in the tubes and the stress in the tube walls will create a force that together with the hydrostatic lifting forces will space and float the piston and piston rings relative to the walls of the power cylinders. Tubes  148  are made from a suitable flexible and resilient material (either metal or a composite material) that exhibit good compliant characteristics so as to have a sufficient spring rate to properly load the piston rings as was described immediately above.  
         [0090]     As is apparent from the foregoing the air for the hydrostatic bearings lubricates and cools the piston rings and provides additional combustion air albeit a small amount. In addition the hydrostatic bearings float the piston and piston rings which serve to minimize the side loadings and friction. The side loadings are further eliminated by use of the four bar linkage system. The centering action of the hydrostatic bearings also serves to minimize blowby between the ring and the cylinder.  
         [0091]     The air bearing piston ring  130  is a low mass airflow device but, more importantly, only requires 100 to 200 psig to operate. These lower actuation pressures result in several benefits. One is a very low parasitic power loss (e.g. the power required to supply this air would run from 0.6 to 3.2 hp, respectively for a 50 hp cylinder). Secondly, pressurizing ambient air to 100 or 200 psig results in less heat added to the air. Therefore its temperature is not increased substantially. Consequently, the value of the pressurized air as a coolant for the piston ring  130  is enhanced. This, in turn, allows higher power levels to be run before the material limits of the ring  130  are reached.  
         [0092]     Lower actuation pressure is achieved by depressed trenches  242  that surround the support pockets  244 . These trenches  242  communicate with the low pressure side of the pistons  38  and  40  (e.g. opposite the combustion side). Trenches  242  consist of circumferential grooves  246  and by-pass slots  248 . Slots  248  are typically at a lower pressure, therefore the air bearing effect works. To maximize the amount of lift for the provided actuation pressure, the flowable area of the pockets  244  is maximized for the circumference.  
         [0093]     During the dynamic combustion process where very high pressures are incurred above the air bearing piston ring  130 , a very complex set of events are evolving that are more effectively managed by this newly configured &amp; optimized ABPR. For example, this ABPR operating in a 50 hp cylinder could provide a ring to cylinder clearance of 0.00090″ with 0 psig above the ring  130  and a low 0.00010″ or even 0.00000″ at combustion at top dead center with supply pressures of 10 psig to 200 psig, if the designer should desire. However, at 0.00000″ of clearance, special care must be taken to plan for the minimal wear that would eventually occur in this zone very near top dead center. This happens because at combustion, the hydrostatic life force must be significantly large to lift the ring off the cylinder.  
         [0094]     This levitation force during combustion must be large because in opposition are the loads of the air supply tubes or lines  204  and the net combustion pressure on the backside of the ring  130 . The combustion pressure on the frontside of the ring  130  is not as critical since its effective area of its impact is smaller than the combustion pressure area on the backside. This frontside area is represented as surface A in  FIG. 22 . This net force pushes the ABPR radially outward and reduces the ring-to-cylinder clearance. Also, any high pressure combustion leakage (blowby) is quickly short circuited by the trenches  242  and prevents backfeeding into the inclusions. This short circuiting maintains the cooling effect of the supply air at the ABPR, prevents combustion product contamination of the pads/inclusions, and insures high frequency response of the ABPR levitation system to changes in its force equilibrium.  
         [0095]     This wear would be minimal for several reasons: no piston side loads, seat-out (0″ clearance) occurs at very low piston velocities and wear is a function of frictional force and velocity differentials.  
         [0096]     The air bearing piston ring  130  is lubricated with air rather oil. The purpose of the piston ring  130  is to minimize the leakage of the combustion pressure and this requires a very small radial clearance between the ring  130  and cylinder wall  98 . The clearance between the piston ring  130  and the wall  98  is small and is just large enough to levitate the load, i.e., the piston  38  and  40 . The unique feature of the piston ring  130  is that is self equalizing. For example, if too much air is leaking out around the edges of the lift pocket or support pockets  244 , the clearance decreases to reduce the airflow rate to equal the air supply rate. Conversely, if the air leakage out of the edges of the lift pockets  244  is too small, the air pressure increases and lifts the load to attain the correct airflow. It has been found that the functional clearance with airbearing piston rings  130  is small and in the correct range to provide acceptable combustion pressure ceiling. This clearance is similar to a typical radial clearance when oil is separating a piston ring from a cylinder wall and, more importantly, the clearances are centered around the piston. In a normal oiled piston, the piston side loads create an eccentricity that increases blowby. Thus the small, centered air filled clearance is functionally equivalent the eccentrically distributed, oil filled clearance of an oil lubed piston and ring assembly.  
         [0097]     An optimum combination of functional variables exist to provide the best sealing and minimum rubbing contact for virtually no wear. The prime variables that control the function of the piston ring  130  are: 
        1. Pressure differential between secondary air supply pressure versus piston ring pocket sink pressure=ΔP S       2. Air flow rate #/second=Q S       3. Air pocket area and quantity of pockets A N       4. Perimeter length around air pockets (P L )     5. Spring rate of piston ring (k R )     6. Spring rate and preload of air supply tubes set by offset bend dimension (k t )     7. Coefficient of friction between piston groove (f G ) land and piston ring seal surfaces.     8. Air supply tube inside air flow area (A t )     9. Low pressure sink groove size and depth around the piston ring pocket (A S )     10. Piston ring to cylinder radial clearance ring (R C )     11. Combustion Pressure maximum value and range  
         P   C     ⁢         Q   S     =     Δ   ⁢           ⁢     Ph   2     ⁢   hb         12   ⁢   μ   ⁢           ⁢   l           
    12. Q S =pending air flow Clearance=f(ΔP S ,Q S ,A N ,k R ,k t ,f g ,A T ,A S ,P C ,μ,A CI ,A CD )     13. Area of ring inside diameter exposed to combustion pressure A CI       14. Area of ring outside diameter exposed to combustion pressure A CO.          
 
         [0112]     All of the above listed variables are adjustable to meet differing requirements. For each requirement there is an optimum combination to minimize either wear or leakage. For example, if abrasive fuel is a requirement, these variables can be adjusted to float the ring on an air film through the entire stroke, with a small increase in blow by. For maximum efficiency and minimum blow by, the ring seals land would contact the cylinder when exposed to high pressures from initial combustion at top dead center. To minimize the wear at the region of contact or the cylinder can have a suitable coating over a very short travel at top dead center and then float the ring the rest of the stroke, with a small wear penalty. Both the intensity of the ring to cylinder contact pressure and, or length of stroke with contact are adjustable. For example, increasing the second air pressure ΔP S  alone will reduce the stroke length of high piston ring contact pressure and slightly reduce the contact pressure.  
         [0113]     With reference to  FIGS. 14, 15  and  16 , the piston rod seal pack  188  will be described. The piston rod seal pack  188  seals the piston rods or connecting rods  56  and  58  with respect to the openings  190  in plate  170 . The connecting rods  56  and  58  reciprocate within seal packs  188  mounted in openings  190 .  
         [0114]     The seal packs  188  are constructed to seal the rods  56  and  58  with respect to openings  190  and to allow for slight lateral movement of the rods  56  and  58  and the seal pack  188  relative to the plate  170 . The seal packs  188  have an outer casing  192  and two inner casings  194  and  196 . The casings are held together with a snap ring  198 . It should be appreciated that a single housing could be used, but for ease of manufacture, three separate casings were used. To use the high grade seals in a one piece installation would have contorted the seals beyond their elastic limits and decreased their sealing effectiveness and durability. (It should also be noted that it may be possible for a production environment and for some applications to use engineering materials that could be overmolded around the seals.) The casings contain seals  199  which seal against the connecting rod. An O-ring  202  is mounted between casings  194  and  196  and shares a groove in casings  192  and  196 . A second O-ring  202  is mounted between plate  170  and casing  192 . The second O-ring  200  shares a groove formed partly in each of these parts. Having shared grooves for supporting the O-rings allows for better sealing and for some movement with respect to the sealed components, and provides for high strength retention to oppose the oscillating movements of the rods  56  and  58  since the O-ring is practically in pure shear.  
         [0115]     With reference to  FIG. 17 , the air supply line  204  for feeding pressurized air to the ring  130  will be described. The air supply line  204  connects to the bearing pack  60 . The pack  60  has internal channels which direct air to the connecting rods  56  and  58  which have a channel  206  for routing air to manifold  150 . The line  204  has a connector  208  for connecting the bearing pack line  210  to the outer casing line  212 . Outer casing line  212  is connected to casing  12  and to a supply of pressurized air. A standard connector  211  is used to connect line  212  to casing  12 . The connector  208  has a long body with an O-ring cavity  216 . The long body facilitates sealing because a long tube in a long hole minimizes leakage. The use of an O-ring further reduces leakage. As a result, the connector  208  seals against leakage while simultaneously allowing tube  210  to move or reciprocate with respect to tube  212  as a result of movement in the bearing pack  60 .  
         [0116]     With reference to  FIG. 18 , the connection between the cams  46  and  48  and the output or rotary shaft  14  is illustrated at  220 . In the disclosed embodiment, the connection at  220  is through gear teeth  222  on both the cam and shaft. Additionally, the oil slinger  230  is illustrated in  FIG. 18 . The oil slinger  230  is preferably part of the cams  46  and  48 . Slinger  230  extends along the rotary shaft  14 . In the most preferred design, slinger  230  is a hollow tube with a plurality of openings which allow it to be slung out to lubricate the bearing packs  60 . The holes face the bearing packs  60 . Oil is fed to the slingers  230  through feed line  232 . Tube  232  is connected to an oil reservoir and pump to direct oil to slinger  230 .  
         [0117]     With reference to  FIG. 21 , the shaft seal tube  234  of the present invention will be described. Shaft seal tube  234  is concentric with and receives the rotary shaft  14 . The shaft seal tube  234  is connected between the end plates  170 . The seal tube  234  is a cylinder with a flange  236  on both ends. The tube  234  is sealed with an O-ring  236  to each of the end plates  170 . Tube  234  eliminates the need for a seal from the stationary endplate to the rotating rotary shaft  14 . The seal tube  234  also provides support to the endplate  170  and adds stiffness. The end plates  170  are supported on the outside diameter by the center case  238  and the bearing case  240 . The center case  238  and the bearing case  240  define the outer case assembly  12 . The endplate  170  is supported on the inside diameter by the shaft seal tube  234 . The loads from the pistons  38  and  40  are transmitted to the end plates  170 . With the end plates  170  supported at the inside and outside diameters, the link mount points  180  and  182  deflections are minimized. The minimized deflection of the link mount points  180  and  182  is important to maintain the straight line motion of the bearing packs  60  attached to the pistons  38  and  40 .  
         [0118]     Although this invention has been shown and described with respect to detailed embodiments thereof, it will be appreciated and understood by those skilled in the art that various changes in form and detail thereof may be made without departing from the spirit and scope of the claimed invention.