Abstract:
A rotary engine having a pair of nested rotors positioned with a housing to define a central combustion chamber and four subchambers having variable volumes. Rotors are mounted about a driveshaft extending through the housing. The rotary engine utilizes tunable gas compression and expansion in order to manage emissions without needing exhaust gas recirculation or a complicated and expensive fuel injection system. The rotary engine is relatively simple and inexpensive to manufacture, has no valve train, is vibrationless, has high power density, and has a wide speed range.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims priority to U.S. Provisional Application Ser. No. 60/127,082 filed Mar. 31, 1999. 
    
    
     TECHNICAL FIELD OF THE INVENTION 
     The present invention relates generally to internal combustion engines and, more particularly, to a rotary engine. 
     BACKGROUND OF THE INVENTION 
     Central to the operation of any internal combustion engine are three components: the working volume, the power transmission, and the timing mechanisms. The working volume of a reciprocating piston engine is created by a cylindrical chamber capped at one end by the head and at the other end by a movable piston. The power of expanding gas forces the piston down and is transmitted to offset journals on the crankshaft through a connecting rod. The timing is controlled by the position of the offset journals on the crank and by a cam operated poppet valve assembly in the head. 
     History has proven that there is an intense and timeless desire to improve the internal combustion (IC) engine. Improving fuel efficiency, power density, cost, and/or durability would be very appealing to any IC engine user. Improving most of these attributes without degrading others would be a dramatic advance in engine technology. Each year hundreds of attempts and millions of dollars are poured into making small improvements to the IC engine. Designers continue, however, to spend money looking for the one elusive key that will unlock a major advance in engine technology. 
     Over the years the rotary engine has been especially attractive in this search for a dramatic technological advance because of its demonstrated ability to deliver power in a small package. One of the primary advantages of the rotary engine is the simplicity of the mechanism that defines the working volume, power transmission and timing mechanisms. The simplicity of design not only reduces the number of moving parts, therefore increasing reliability, but the more efficient packaging reduces the outside envelope of an engine of given displacement. 
     The Wankel engine designed throughout the late 1930&#39;s and early 1940&#39;s by Felix Wankel and sold commercially by a number of companies, including Mazda Motor Company, is currently the most well known rotary engine. This engine proved popular even though it had several problems, including high hydrocarbon emissions, contributed to by such factors as rotor sealing, lubrication and port configuration, only a single power pulse per revolution per stage, and poor fuel economy. But, more importantly, this engine showed that the rotary engine could successfully be used in a mass-produced automotive market if appropriate attention, time and money are spent to develop the technology. 
     While the Wankel is the most well known rotary engine, it is by far not the only rotary engine known in the prior art. Many rotary engines have been patented in the past, including U.S. Pat. No. 1,298,839 to Weed; U.S. Pat. No. 2,050,603 to Gardner; U.S. Pat. No. 2,734,489 to Tschudi; U.S. Pat. No. 3,824,963 to Eda; U.S. Pat. No. 3,854,457 to Taurozzi; U.S. Pat. No. 4,194,871 to Studenroth; U.S. Pat. No. 5,326,238 to Schukey and U.S. Pat. No. 4,604,909 to Marson. 
     The thing that all of these engines have in common is the “rotating piston” design. All of these engines use a set of rotors that move within an annular volume. The difference between these prior art engines is primarily in the design the linkage that transmits the power and timing the engine. The Tschudi Engine, for example, uses a modified Geneva mechanism to move the pistons around the volume in alternating steps. This results in an engine that takes two revolutions to complete a power cycle (i.e., transmit a power pulse to each piston), giving it a power density similar to a reciprocating piston engine. Still other prior art engines took advantage of the fact that a mechanism could be created that allowed the engine to complete one power cycle per chamber every revolution. This increased the potential power density in the engine, but the fragile mechanisms used for timing these engines led to their downfall. All of these engines were susceptible to various failure modes. 
     Therefore, the various prior art rotary engine designs indicate that the rotary engine concept has the potential to outperform reciprocating piston engines in most respects if a more practical design can be developed. The present invention is directed toward meeting this need. 
     SUMMARY OF THE INVENTION 
     The present invention relates generally to a rotary engine having a pair of nested rotors mounted about a driveshaft and having appended vanes which define four variable volume combustion chambers. The present invention utilizes tunable gas compression and expansion in order to manage emissions without needing exhaust gas recirculation (EGR) or a complicated and expensive fuel injection system. The present invention is also relatively simple and inexpensive to manufacture, has no valve train, is vibrationless, has high power density, and has a wide speed range. The design of the rotary engine of the present invention therefore offers improvements in several areas of the engine. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a perspective view of a first embodiment pair of rotors of the rotary engine of the present invention. 
     FIG. 2 is an end elevational view of a pair of rotors according to the present invention, illustrating the location of the four combustion chambers of the rotary engine. 
     FIG. 3A is a partial cross-sectional side view of the nested rotors of the present invention contained within an engine housing. 
     FIG. 3B is a partial cross-sectional side view of the arrangement of FIG.  3 A. 
     FIGS. 4A-D are schematic end elevational views of a nested rotor pair inside a housing, illustrating an Otto cycle of the rotary engine of the present invention. 
     FIG. 5 is a graph of rotor speed versus driveshaft position, illustrating a variable rotor speed aspect of the present invention. 
     FIG. 6 is a side elevational view of a first embodiment rotor of the present invention. 
     FIG. 7 is a side elevational view of a second embodiment rotor of the present invention. 
     FIG. 8 is an exploded perspective view of the second embodiment rotor of the present invention mounted upon a driveshaft with a first preferred embodiment timing mechanism of the present invention. 
     FIG. 9 is a schematic diagram of the angular relationship between rotor pins and drive flange  219  pins according to one embodiment of the present invention. 
     FIG. 10 is a plan view of a first preferred embodiment rotor position timing mechanism of the present invention, in a first position. 
     FIG. 11 is a plan view of the timing mechanism of FIG. 10 in a second position. 
     FIG. 12 is a partial cut-away view of a pair of second embodiment rotors nested on a driveshaft. 
     FIG. 13 is a cross-sectional view of the arrangement of FIG.  12 . 
     FIG. 14 is a perspective view of a preferred embodiment rotor inner seal of the present invention. 
     FIG. 15 is a partial cut-away view of rotor inner seal of FIG. 15 installed with the arrangement of FIG.  13 . 
     FIG. 16 is a cross-sectional view of a portion of rotor inner seal of FIG.  15 . 
     FIG. 17 is a schematic cross-sectional view of a prior art Wankel rotor and housing. 
     FIG. 18 is a schematic cross-sectional view of an “area contact” afforded by rotor outer seal of the present invention. 
     FIG. 19 is a schematic cross-sectional view of a preferred embodiment rotor outer seal of the present invention. 
     FIG. 20 is an exploded perspective view of a preferred embodiment rotor and rotor outer seal of the present invention. 
     FIG. 21 is a schematic cross-sectional view of a “scissors action” exhibited by the preferred embodiment rotor outer seal of the present invention. 
     FIG. 22 is a graph of rotor outer seal contact pressure versus engine speed. 
     FIG. 23 is a cross-sectional view of a preferred embodiment rotor outer seal counterbalance mechanism of the present invention. 
     FIG. 24A is a front elevational view of a preferred embodiment front drive flange of the present invention. 
     FIG. 24B is a side cross-sectional view of the front drive flange of FIG.  24 A. 
     FIG. 24C is a rear elevational view of the front drive flange of FIG.  24 A. 
     FIG. 24D is a top cross-sectional view of the front drive flange of FIG.  24 A. 
     FIG. 24E is a rear elevational view of a preferred embodiment rear drive flange of the present invention. 
     FIG. 24F is a side cross-sectional view of the rear drive flange of FIG.  24 E. 
     FIG. 24G is a front elevational view of the rear drive flange of FIG.  24 E. 
     FIG. 24H is a top cross-sectional view of the rear drive flange of FIG.  24 E. 
     FIG. 25A is a partial side sectional view of the interface of the drive flange and bearing of the present invention. 
     FIG. 25B is an exploded view of the interface of the drive flange and bearing of the present invention. 
     FIG. 26A is a side sectional view of the exhaust flange connection and housing of the present invention. 
     FIG. 26B is a perspective view of the exhaust flange connection of FIG.  26 A. 
     FIG. 27A is a top cross-sectional view of a preferred embodiment journal bearing of the present invention positioned between two housing halves. 
     FIG. 27B is a side cross-sectional view of the journal bearing of FIG. 27A positioned between a housing halve and a rotor hub. 
     FIG. 28 is a cross-sectional schematic view of a preferred embodiment central combustion site of the engine of the present invention. 
     FIG. 29A is a schematic view of a preferred embodiment cam ring and cam ring actuator of the present invention. 
     FIG. 29B is a first schematic view of the relationship between two rotatable cam rings of the present invention. 
     FIG. 29C is a second schematic view of the relationship between two rotatable cam rings of the present invention. 
     FIG. 29D is a chart plotting the cam phase angle as a function of chamber compression ratio. 
     FIG. 30 is schematic diagram showing a preferred cam profile and roller center profile for a first preferred embodiment rotor position timing mechanism of FIG.  10 . 
     FIG. 31 is a graph of rotor phase angle and rotor separation angle versus driveshaft angle for the system of FIG.  30 . 
     FIG. 32 is a graph of chamber volume versus driveshaft angle for the system of FIG.  30 . 
     FIG. 33 is a graph of chamber volume versus rotor separation angle for the system of FIG.  30 . 
     FIG. 34 is a graph of chamber volume versus driveshaft angle for a “dwell” cam profile of the present invention. 
     FIG. 35 is a graph of chamber volume versus driveshaft angle for a “spike” cam profile of the present invention. 
     FIG. 36A is a perspective view of a valve cluster of the present invention. 
     FIG. 36 b  is a partial side sectional view of a pair of valves emplaced in the housing of the present invention. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     For the purpose of promotion an understanding of the principles of the invention, reference will now be made to the embodiment illustrated in the drawings and specific language will be used to describe the same. It will nevertheless be understood that no limitation of the scope of the invention is thereby intended, such alterations and further modifications in the illustrated device, and such further applications of the principle of the invention as illustrated therein being contemplated as would normally occur to one skilled in the art to which the invention relates. 
     The present invention relates to a rotary engine, illustrated in FIG. 1, having four combustion chambers  7 A, B, C, D defined by two rotors  10 A and  10 B. The gas expansion section of rotary engine is composed of two rotors  10 A, B which are mounted upon a common drive shaft (not shown in FIG.  1 ). Rotors  10 A, B face each other and move within an enclosed cavity (not shown in FIG. 1) with oscillatory rotary motion superimposed upon uniform rotary motion. Rotors  10 A, B are interdigitated in that each of rotors  10 A, B includes two vanes  11  and  12  located 180° apart. 
     As described in greater detail hereinbelow, rotors  10 A, B are mounted to driveshaft  204  in such a way that rotors  10 A, B can be rotated approximately 90° relative to one another. As illustrated in FIG. 2, it can be observed by looking at rotors  10 A, B from one end that the spaces defined by rotor vanes  11  and  12  and the cylindrical cavity in which rotors  10 A, B rotate form four combustion chambers  7 A-D. Moving one of rotors  10 A, B with respect to the other rotor  10 B, A causes the volume of all four chambers  7 A-D to be altered. Two of the chambers  7 A, C get smaller, while two of the chambers  7 B, D get larger. By superimposing oscillatory rotary motion upon uniform rotary motion of rotors  10 A, B these four chambers  7 A-D can be associated with the four parts of the Otto cycle, namely intake, compression, expansion, and exhaust. 
     Referring to FIG. 3, the outer edges of rotor vanes  11  and  12  are preferably formed in a circular configuration so that the vanes mate with a toroidal outer casing  14 . This outer casing  14  encloses the four combustion chambers  7 A-D so that compression and expansion will occur in all four chambers  7 A-D when one rotor  10  is moved with respect to the other. As described hereinbelow in greater detail, a suitable system of seals is desirable in order to prevent gases from bleeding through the clearances between the vanes  11  and  12  and the casing  14  and into adjacent combustion chambers  7 . 
     Referring now to FIGS. 4A-D, it can be seen that if an intake port  18  and exhaust port  20  are provided in housing  14 , and if an ignition source  16 , such as a spark plug or diesel fuel injector, are also provided in housing  14 , the four portions of the Otto cycle can be accomplished in one complete rotation of rotor set if rotors  10 A, B are moved in a certain way with respect to one another. As shown in FIG. 4A, the uppermost chamber  7 A has just completed the intake cycle and has been filled with a quantity of air and fuel through the intake port  18 . During this intake portion of the Otto cycle, rotor  10 B is rotating at a rate slower than rotor  10 A, thereby causing expansion of upper chamber  7 A during the intake portion. In other words, vane  11 A is moving away from vane  11 B during this intake cycle. Referring to FIG. 4B, after intake is complete, the compression portion begins by causing rotor  10 B to rotate at a speed greater than rotor  10 A. In other words, vane  11 A and vane  11 B are now moving toward one another. This causes a reduction in the size of chamber of interest  7 A, thereby compressing the air/fuel mixture during the compression portion of the Otto cycle. Once chamber of interest  7 A has been fully compressed, it moves past ignition source  16 , which fires at this point, thereby initiating combustion. As shown in FIG. 4C, the chamber of interest  7 A undergoes expansion as the air/fuel mixture combusts, whereby much of the energy of combustion is translated into the rotational motion of driveshaft  204 . Finally, FIG. 4D illustrates chamber of interest  7 A passing exhaust port  20 , where exhaust gases from the combustion process are expelled. Vane  11 B is moving toward vane  11 A at this point, decreasing the size of chamber of interest  7 A and thereby forcing exhaust gases out of port  20 . 
     As will be apparent from a study of FIG. 4, each chamber  7 A-D may complete four parts of the Otto cycle during one revolution of rotor set  10  if rotors  10 A, B are made to accelerate and decelerate with respect to one another as they sweep through their revolution. Some motion control mechanism must therefore arrange for proper motion between rotors  10 , as is described in greater detail hereinbelow. Those having ordinary skill in the art will recognize that there is an economy of motion in the present design wherein four different chambers  7 A-D are actuated by a single movement of one of rotors  10 . This economy of motion translates into low friction, low vibration, and a high power density. 
     Driveshaft  204  to which rotors  10 A, B are coupled spins at an angular velocity of θ S . Rotor  10 A has an average velocity of θ S  also, but oscillates in a way that causes it to go faster than the driveshaft for awhile, and then slower than the driveshaft for awhile. Rotor  10 B does the inverse, as illustrated in FIG.  5 . In a preferred embodiment, the angular velocity of rotor  10 A and rotor  10 B may be made to vary sinusoidally according to the following equations: 
     
       
         {dot over (θ)} RA ={dot over (θ)} S +(1+sin 2{dot over (θ)} S )  (1) 
       
     
     
       
         {dot over (θ)} RB ={dot over (θ)} S   +( 1−sin 2{dot over (θ)} S )  (2) 
       
     
     Of course, the motion of rotors  10 A, B with respect to one another need not be limited to a sinusoidal relationship if the chosen motion control device is capable of producing other motions. 
     1. Rotor Geometry 
     As stated hereinabove, it is desired that rotor  10 A and rotor  10 B nest with one another in order to form the four combustion chambers  7 A-D. A first embodiment rotor geometry of the present invention is illustrated in FIG. 6, wherein rotor  10 B is shown in side elevation. Vanes  11 B and  12 B are coupled to a hub  102 B that has a frustoconical shape. It will be appreciated by those having ordinary skill in the art that the shape of the vanes  11 ,  12  of each rotor  10 A, B allow them to nest tightly with hub  102  of the opposite rotor  10 B, A. One consequence of the frustoconical shape of hub  102 B is that a force represented by the arrow  104  is placed upon hub  102  by the gases within each combustion chamber  7 . This force  104  has a substantial horizontal component indicated by arrow  106 , which tends to force hubs  102 A, B away from one another during operation of rotary engine. This force can assist in sealing rotor hub joint  103 . 
     An alternative design for rotors  10  is illustrated in FIG. 7, wherein the hub  102 B has a cylindrical configuration with vanes  11 B and  12 B appended therefrom. The embodiment of FIG. 7 substantially eliminates any horizontal force component resulting from pressure of the combustion chamber gases against hub  102 . Furthermore, the alternative design of FIG. 7 is less sensitive to the position of rotors  10  with respect to one another than is the frustoconical hub design of FIG.  6 . This is because any horizontal displacement of rotor  10 A away from rotor  10 B in the design of FIG. 6 will result in a gap between the vanes  11 ,  12  of one rotor  10 A, B and the hub  102  of the other rotor  10 B, A. It can be seen that the rotor design of FIG. 7 does not suffer from this problem, in that horizontal displacement of one rotor  10 A, B away from the other rotor  10 B, A does not cause vanes  11 ,  12  to pull away from the hub  102  of the other rotor  10 B, A. 
     2. Indexer Geometry 
     With reference to FIG. 8, rotors  10 A, B will each have two drive pins  202 A, B spaced 180 degrees apart and protruding from the back of rotors  10 A, B. Secured firmly to driveshaft (crankshaft)  204  is a drive flange  219  that has two similar pins  206  spaced 180 degrees apart (FIG.  9 ). Pins  206  secured to driveshaft  204  will rotate at the same angular velocity as driveshaft  204 . In this configuration, as driveshaft  204  is turned, the drive pins  206  will rotate through approximately 90 degrees before they strike rotor pins  202  and drive them at the same angular velocity as driveshaft  204 . A power transmission mechanism  207  is therefore created, but a timing mechanism is still needed. 
     As shown in FIG. 10, with drive pins  206  and rotor pins  202  spaced at 90 ° degrees, four levers that pivot on these pins are added creating a parallelogram. In this case the levers have equal lengths and the parallelogram formed therefrom is a square. This linkage forms the basis of the timing mechanism  208 . Two non-adjacent sides of the parallelogram are created by rocker levers  210  each carrying two rollers  212 . The remaining sides are created by two spacer blocks  214  having saddle bearings  216  that ride on the curved surface of the rocker lever rollers  212 . While rocker levers  210  and spacer blocks  214  are not of equal lengths, the sides of the parallelogram formed therefrom are of equal length. Timing mechanism  208  thus comprises a four bar linkage with idler rollers  212  at each comer of the parallelogram (in this case, a square). In FIG. 11, it can be seen that if two non-adjacent rollers  212  are squeezed toward the center, the remaining two rollers  212  will move apart. More importantly the angle between drive pins  206  and rotor pins  202  will change. Timing mechanism  208  allows rotors  10 A, B to slow down and speed up relative to driveshaft  204 , varying the volume in chambers  7 A-D. The final component of this mechanism is a cam ring  211  that dictates the angular shift between drive pins  206  and rotor pins  202  by providing the appropriate “squeeze” to timing mechanism  208  as it rotates. This is described hereinbelow in section  11 . The mechanism also includes an outer bearing  217 , drive flange  219  and fastener  221 . 
     3. Rotor Inner Seal 
     As shown in FIG. 12, rotary engine of the present invention will utilize two rotors nested about one another and fitted onto a driveshaft  204 . Because combustion chambers  7 A-D are defined by rotor wings  11  and  12 , rotor hubs  102 , and driveshaft  204 , it is necessary to provide a seal that will prevent the contents of one combustion chamber  7  from migrating to an adjacent, lower pressure combustion chamber  7 . 
     Because combustion chamber  7  volume is formed by two rotor hubs  102 A, B meeting together, both of which are concentric to driveshaft  204 , the seal must take into account both manufacturing tolerance variations and tolerance variations caused by wear over the life of engine  5 . Provision must therefore be made to seal combustion chambers  7  with variable clearance. 
     With reference to FIG. 13, it can be seen that there are two possible leakage paths for gases to escape from combustion chamber  7 . Path “A” is inward toward driveshaft  204 , while path “B” is circumferential toward adjacent low-pressure chambers  7 . Any viable seal design should deal with these two leakage paths and also cope with the variation in rotor-to-rotor hub clearance. 
     A preferred embodiment rotor inner seal is illustrated in FIG.  14  and indicated generally at  302 . Rotor inner seal  302  is formed from two identical halves  304  and  306 . Each seal half  304 ,  306  includes a plurality of circumferential fingers  308  which mesh with the complimentary fingers  308  of the other seal half  306 ,  304 . 
     The positioning of rotor inner seal  302  with respect to rotors  10 A, B and with respect to driveshaft  204  is illustrated in FIG.  15 . The periphery of rotor inner seal  302  is formed into a conical surface  310 , which fits into a space between rotors  10 A, B that is complimentary to the shape of rotor inner seal  302 . As shown more clearly in FIG. 16, rotor inner seal halves  304 ,  306  are pressed against the inner surfaces of rotors  10  by means of internal springs  312  which bias the halves  304 ,  306  away from an internal ring  314 . Springs  312  insure sealing at engine startup. Gas pressure loading of hubs  102  against rotor inner seal  302  augments this force during operation of engine  5 . Those having ordinary skill in the art will recognize that other biasing means known in the art may be used to apply internal pressure to rotor inner seal  302  in place of helical springs  312 . 
     4. Rotor outer Seal 
     In addition to rotor inner seal  302 , rotary engine of the present invention must also provide a seal between rotor vanes  11 ,  12  and the interior surface of housing  14 . Rotary engine of the present invention offers superior sealing at rotor  10  periphery than other prior art rotary engines. One primary advantage of the present design is that the seal at rotor  10  periphery exhibits “area” contact, unlike the prior art Wankel engine, which exhibits “line” contact. A schematic cross-sectional view of the prior art Wankel rotor  402  within a housing  404  as illustrated in FIG.  17 . Because of the peculiarities of the Wankel design, rotor  402  rocks back and forth within housing  404 , which only allows a line contact area  406  between rotor  402  and housing  404 . 
     In contrast, rotor  10  of the present invention exhibits an outer surface  408  which generally conforms to the shape of the interior of housing  14 . Furthermore, rotor  10  spins upon a (relatively) fixed axis (the centerline of driveshaft  204 ). These factors make it possible to provide a rotor outer seal  4   10  which has a relatively wide contact area  412  with the inner surface of housing  14 , which lowers unit pressure loading to reduce wear. Furthermore, this periphery seal  410  always exhibits forward motion relative to the wall of chamber  7 . Conventional crank slider engine seals experience a zero velocity condition called “ring reversal,” which causes the lubrication layer between the seal and chamber  7  surface to disappear. This creates a high wear region (i.e. top ring “turnaround” wear). The present rotary engine design will not experience this type of seal and/or chamber wall degradation, since seals  410  always have a forward velocity that promotes hydrodynamic lubrication. 
     It is desirable to minimize the “crevice volume” in a combustion chamber  7 . With reference to FIG. 19, the preferred embodiment rotor outer seal design of the present invention resolves this difficult problem by providing parallel seal mounting surfaces  414  on rotor  10 , then establishing a “shoe” seal configuration  416  that slides over these parallel surfaces  414 . 
     Rotor  10  and rotor outer seal  410  are illustrated in an exploded perspective view in FIG.  20 . In a preferred embodiment of the present invention, rotor outer seal  410  comprises two seal rails  418  connected by a web  420 . Rails  418  follow the circumferential contours of web  420 , but are formed to extend perpendicular to web  420  top and bottom surfaces, such that rails  418  extend away from web  420  in the direction of rotor  10  so as to provide a surface to mount onto parallel surfaces  414  of rotor  10 . Rails  418  further extend from web  420  in a direction away from rotor  10  in order to provide seal surfaces  418  which wipe the inner wall of housing  14 . In a preferred embodiment, the rail outer contact surfaces  418  are formed from an appropriate wear resistant coating and web  420  is formed from metal, such as steel or aluminum. Outer seal  410  can be manufactured in a variety of ways known in the art, including stamping, turning, P/M (powdered metal), casting, etc. 
     In one embodiment of the present invention, engine housing  14  is formed as a two-piece “clamshell” design. As illustrated in FIG. 21, the interface joint  422  between housing halves  423 A, B must be traversed by rotor outer seal  410 . Rotor outer seal  410  crosses interface joint  422  using a “scissors action” as indicated at  424 . This avoids damage to seal  410  and prevents the creation of audible clicking noise as seal  410  traverses the joint  422 . 
     5. Outer Seal Counterbalance 
     All rotary engines experience changes in peripheral seal contact pressure due to the centrifugal force of rotation. As illustrated in FIG. 22, the goal is to have a constant contact pressure between rotor outer seal  410  and the housing interior surface  501  independent of engine speed. However, as shown in FIG. 22, the actual outer seal contact pressure increases exponentially with increasing engine speed. 
     The rotary engine of the present invention can utilize a counterbalance mechanism to negate the engine speed effect on seal contact pressure. A preferred embodiment of the counterbalance mechanism that is illustrated in FIG. 23, and indicated at  502 . Counterbalance mechanism  502  places a shaft  504  into a space within rotor vane  11 ,  12 . One end of shaft  504  carries a counterweight  506 , while the other end  505  of shaft  504  is engaged with one end of seal  410 . The opposite end of seal  410  is coupled to rotor  10  on the opposite side of vane  11 ,  12 . Shaft  504  is allowed to pivot upon a fulcrum  508  between counterweight  506  and end  505 . 
     In operation, rotation of rotor  10  causes centrifugal force to act upon counterweight  506 , forcing counterweight  506  away from driveshaft  204 . This, in turn, causes shaft  504  to pivot upon fulcrum  508 , thereby driving end  505  of shaft  504  toward driveshaft  204 . It can thus be seen that fulcrum  508  reverses counterweight  506  through the application of centrifugal force. This allows end  505  of shaft  504  to move seal  410  toward driveshaft  204 , which is counter to the direction that centrifugal force is attempting to move seal  410 . With proper tuning of the mass of counterweight  506  and the distances between counterweight  506 , fulcrum  508  and end  505 , the counterbalance mechanism  502  will “zero out” the centrifugal forces acting upon seal  410 . Reducing or eliminating the centrifugal force helps seal  410  to experience much lower forces, and hence, reduces friction and increases the life of rotor outer seal  410 , as well as increasing the efficiency of engine  5 . Preferably, a bias spring  510  would be included to give seal  410  a constant contact pressure. Gas pressure would also act upon seal  410  in a way to increase unit loading of the seal contact. 
     6. Drive Flange Attachment 
     Because rotors  10  oscillate relative to each other as they rotate, drive flange  219  must transmit large, pulsating torques (see FIG.  8 ). Drive flange  219  must also provide for precise location, both axial and radial, relative to driveshaft  204 . Additionally, driveshaft  204  must allow pressurized lubricant to pass from the center of driveshaft  204  to drive pins  206 . 
     FIGS.  8  and  24 A- 24 H illustrate one preferred drive flange  219  geometry for accomplishing the above-mentioned goals with a minimum of complexity and expense. 
     FIGS. 24A-24D illustrate views of a preferred front drive flange  219 A, while FIGS. 24E-24H illustrate views of a preferred rear drive flange  219 B. Preferably, drive flange  219  is fabricated from powder metal precursors by powder metallurgy techniques, although drive flange  219  may be formed through other known metallurgical fabrication techniques, such as casting and machining. Drive flange  219  includes a tapered central aperture  670  adapted to receive the tapered end  672  of driveshaft  204 . The taper of central aperture  670  and driveshaft end  672  are matably matched, and are preferentially 10-30°. Central aperture  670  and driveshaft end  672  also include matable spline teeth  674 ,  676  (see FIG. 8) of a conventional geometry. Aperture spline teeth  674  and driveshaft spline teeth  676  are configured at the preferred taper angle, matching that of central aperture  670 . One spline is omitted from drive flange  219  and from driveshaft  204  to provide a definite reference angle upon assembly, defining a single orientation in which drive flange  219  and driveshaft  204  may be assembled. Spline aperture  678  as defined by omitted spline teeth  674 ,  676  also provides an inlet for lubricant. 
     7. Bearing Configuration 
     In addition to the functions discussed above, drive flange  219  is also configured to provide integral bearing surfaces  780 . As is illustrated in FIGS. 25A and B, one such bearing surface  780  serves to provide both axial and radial support. Journal and thrust bearings are known to be very cost effective and have low friction when properly lubricated. In the preferred embodiment, integral bearing surface  780  has both journal and thrust characteristics, and is preferably formed from powder metallurgy precursors. Other embodiments are contemplated wherein bearing  780  has journal only, thrust only, or other combinations of bearing characteristics. Still other embodiments are contemplated in which the bearing  780  is formed through other convenient forming techniques and/or from other convenient precursors. 
     8. Exhaust Flange Geometry 
     In contrast to the multiple exhaust ports required by a crank slider engine, rotary engine configuration of the present invention requires a single exhaust port  20 . As illustrated in FIGS. 26A and B, exhaust flange  886  connection is designed to take advantage of this requirement. Exhaust port  20  crosses the junction between upper housing  423 A and lower housing  423 B, with about half of exhaust port  20  in upper housing  423 A and about half of exhaust port  20  in lower housing  423 B. Preferably, exhaust port  20  has a very short length. Exhaust flange  886  connects using bolts oriented so as to minimize the wall area being contacted by the hot exhaust gas. A separation is arranged to further reduce the heat transfer from the hot exhaust gas back to housing  14 . Alternatively, an insulating element may be installed in the void. Exhaust flange  886  is substantially symmetrical and as such may be connected in one of two reversible orientations, allowing flexibility during engine construction and/or installation. 
     9. Journal Bearing Configuration 
     FIG. 27 illustrates a housing bearing  990  providing both radial and axial support at the junction of rotor hub  102  and housing halves  423 A, B. Preferably, housing bearing  990  is unitary and is clamped tightly between housing halves  423 A, B using a beveled seat  992 . One of the benefits of this arrangement is that rotor hub  102  does not have to traverse a step at the junction of housing halves  423 A, B, thereby avoiding premature wear. Another advantage is that a less complex bearing  990  can be provided at a lower cost. Housing bearing  990  is located positively and sealed positively by the V-profile (or a spherical profile) of housing halves  423 A, B, avoiding leakage of combustion gasses. Additionally, an oil passage  994  is thereby formed allowing lubricant to fluidly communicate with rotors  10 . 
     10. Peak Cylinder Pressure 
     As illustrated in FIG. 28, the rotary engine configuration of the present invention provides one central location  1096  for combustion to occur. Central combustion site  1096  allows an economy to be realized for the placement of combustion-related equipment, such as spark plugs and injectors. This is advantageous, since not only are fewer such devices necessary for the operation of engine  5 , but central combustion site  1096  also offers a central nexus for the placement of measurement devices  1097 , such as temperature and gas pressure sensors. A gas pressure sensor  1097  installed at central combustion site  1096  may continually feed gas pressure data to an electronic controller  1098  to provide, for example, real time adjustment of the fuel injection process. 
     Electronic controller  1098  may use a continuous gas pressure signal to continuously vary the compression ratio of engine  5  during its operation. The ability to vary the compression ratio is known to provide important benefits in areas such as “cold start” compatibility, emissions reduction, fuel economy improvement, power increase, durability increase, and weight reduction. It is therefore possible to provide measurements of a variety of engine parameters taken at central combustion location  1096  to electronic controller  1098  to adjust compression ratio, air intake temperature, spark timing or other important variables to optimize combustion. In other words, feedback to electronic controller  1098  provides for real-time adjustment of any engine process related to combustion. 
     11. Rotatable Cam Rings 
     A method of altering the compression, port timing and combustion timing ratio unique to this design is to rotate cam rings  211 , as shown in FIG.  29 A. Each cam ring  211  is operationally connected to a rotor  10  such that rotation of the cam ring  211  acts on the connected rotor  10  to change the angular relationship between the two rotors  10 . By selectively rotating one or both cam rings  211 , the mirror-image symmetrical relationship between the angular velocity of rotors  10  (as was shown in FIG. 5) may be altered. The mirror-image relationship between the angular velocity of rotors  10  holds only when cam rings  211  are positioned 90° out of phase relative to each other. When either cam ring  211  is rotated, rotor velocities are no longer completely opposed. FIGS. 29B and C show the new velocity relationship when one cam ring  211  is rotated several degrees relative to the other ring  211 . The relationship between the cam ring orientation and the compression ratio is illustrated as FIG.  29 D. It can be seen from FIGS. 29B and C that a rotation of one ring  211  several degrees relative to the other ring  211  has the effect of reducing the compression ratio, and that in fact the compression ratio is maximized when cam rings  211  are oriented 90° relative to each other. 
     By operationally arranging an actuating device  1102 , such as an electric motor equipped with a worm drive, in connection with one or both cam rings  211 , the desired cam rotation may be achieved. Referring back to FIG. 29A, one such arrangement is illustrated, wherein actuator device  1102  connected to electronic controller  1098  is operationally positioned to rotate cam ring  211 . Actuating device  1102  allows individual phasing of rotors  10  through rotation of one or both of attached cam rings  211 . Electronic controller  1098  may be provided with target compression ratio parameters that may be varied in response to operator inputs. The actual compression ratio is calculated by electronic controller  1098  using pressure input data from gas pressure sensor  1097  (see FIG.  28 ). Gas pressure sensor  1097  provides a continuous cylinder pressure signal to electronic controller  1098  for processing using logical algorithms. Electronic controller  1098  in response generates and sends a continuous or periodic signal to cam ring actuator  1102 . A closed loop feedback arrangement is thus formed to yield fast and efficient combustion control. 
     If both cam rings  211  are thusly controlled, both cam rings  211  may be rotated in the same direction by the same moment. Rotation of both cam rings  211  in the same direction by the same moment results in a change in the timing of intake port  18 , exhaust port  20  and combustion device  16 . This capability is known to have important benefits in engine performance optimization; for example, late exhaust port  20  closing provides for the retention of exhaust gasses in combustion chamber  7 . This is known as EGR (exhaustion gas retention) and reduces NOx emissions. 
     The above design yields three distinct advantages. First, a variable compression ratio may be obtained when rotors  10  are adjusted in opposing directions. Second, a variable intake charge flow (i.e. Miller cycle) may be obtained when taken toward the combustion site. Third, a variable EGR may be obtained when taken toward exhaust port  20 . 
     As discussed hereinabove in Section 2 (Indexer Geometry) and referring back to FIG. 8, a cam ring  211  may be used to dictate the angular shift between drive pins  206  and rotor pins  202  by providing the appropriate “squeeze” to timing mechanism  208  as it rotates. A typical cam profile  1150  is illustrated in FIG. 30. A cam ring  211  with an interior opening defining the cam profile  1150  will cause the roller centers of timing mechanism  208  to follow the indicated roller center profile  1152 . Rotor motion generated by cam profile  1150  follows an asymmetric wave as described by τ(θ) in FIG. 31, where T is rotor phase angle and θ is the driveshaft angle. τ(θ) for rotor A is represented by curve  1154 , while τ(θ) for rotor B is represented by curve  1156 . 
     Even though rotor phase angle functions  1154  and  1156  are asymmetric, it must be remembered that there are two rotors  10 A, B that separate the working volumes of engine  5  of the present invention. Both of these rotors  10 A, B are moving simultaneously to vary the respective volumes of chambers  7 . The profiles for both cam surfaces  1150  may or may not be the same, but one cam  211  will be positioned 90 degrees out of phase in order to produce complimentary motion of rotors  10 A, B. Therefore the chamber volume contained between rotors  10  will actually be the combination of the individual rotor motions. The angle between rotor faces, rotor separation angle, will be defined as β. An interesting outcome in the combination of these singular rotor motions is that, if a 90 degree phase exists between cams  211 , the asymmetry of both singular profiles  1154  and  1156  can cancel, resulting in β as a sinusoidal function of θ. This is illustrated as the curve  1158  in FIG.  31 . 
     Using the rotor separation angle, β, the volume for a single chamber  7  as a function of crank angle can be determined using 
     
       
           V=πβ∫w ( r ) rdr   (11.1) 
       
     
     
       
         r 2   
       
     
     
       
         r 1   
       
     
     where r is the radial distance measured from driveshaft  204  center, and w(r) is the cross-sectional width of rotor  10  as a function of r. As a simplified case, the cross-sectional variation of rotor  10  can be assumed to be constant, drastically simplifying equation 11.1 to 
     
       
         r 2   
       
     
     
       
           V=πβw|r   2 /2|  (11.2) 
       
     
     
       
         r 1   
       
     
     where r 1  and r 2  represent the radial distance to the bottom and top of rotor  10 . Using equation 11.2 and arbitrary values for the cross-sectional properties of rotor  10 , a single chamber volume may be calculated as a function of either crank angle, θ, or rotor separation angle, β. Plots of the volume as a function of both of these angles are shown in FIGS. 32 and 33, respectively. It can be seen from this cam profile example that the chamber volume changes sinusoidaly with respect to the crank angle. It can also be seen that the chamber volume can be directly proportional to the angle β. It can be assumed that a proportionality constant can be developed as a function of rotor cross-section and the distance of the cross-section from the center of the crank (driveshaft  204 ). 
     Other cam profiles may be used to achieve various benefits. There are two alternative cam profiles that are especially interesting for combustion optimization. The first is shown in FIG.  34  and is referred to as a “Dwell Profile”  1160 . The Dwell Profile  1160  is shown superimposed upon the sinusoidal profile of FIG.  32 . The Dwell Profile  1160  creates a prolonged dwell period which would be advantageous for combustion processes that are relatively slow or for engine configurations that exhibit very high operating speeds (RPM). The second profile is the “Spike Profile”  1162  shown in FIG.  35 . The Spike Profile  1162  is shown superimposed upon the sinusoidal profile of FIG.  32 . The Spike Profile  1160  avoids any (or most) dwell, creating a “spike” at top dead center (TDC). This profile would be advantageous for combustion processes that are relatively fast and that have a tendency to be explosive. Also, lower RPM engines would benefit from the Spike Profile  1162  by having reduced heat rejection. Combinations of these profiles to provide specific rotor motions or similar rotor motions are expected to offer further advantages. 
     12. Variable Intake Port Timing 
     Rotary engine configuration of the present invention provides for a single site where intake port  18  is situated, serving all combustion chambers  7 , allowing for further economy to be realized for intake process components  1232 , such as fluid control valves. FIGS. 36A and B illustrate one arrangement of fluid control valves  1232  wherein advantage of central intake port  18  is taken. A cluster  1234  of multiple outwardly opening valves is illustrated, with each valve  1232  opening outwardly in a different position in the intake region of housing  14 . By opening one or more valves  1232 , alone or in various combinations, a variety of intake port timing events and intake swirl amounts can be obtained. Thus, variable swirl ratios are obtained from the various options for which valve or valves  1232  to open. Shrouding of one or more valves  1232  is optional. 
     As is illustrated in FIG. 36B, intake valves  1232  do not need to be opened and closed for each chamber  7 , but rather may act only as flow modulation devices capable of remaining static for extended periods of time. Varying the intake port timing and/or the swirl ratio of the incoming charge gasses is known to yield important benefits for combustion and for reducing pump effort. Another important benefit of the design of cluster  1234  is that pumping losses at part load are eliminated. 
     While the invention has been illustrated and described in detail in the drawings and foregoing description, the same is to be considered as illustrative and not restrictive in character, it being understood that only the preferred embodiments have been shown and described and that all changes and modifications that come within the spirit of the invention are desired to be protected.