CA1209925A - Internal combustion engine and operating cycle - Google Patents

Abstract

"INTERNAL COMBUSTION ENGINE AND OPERATING CYCLE"
ABSTRACT OF THE DISCLOSURE

A cylindrical rotor is mounted for rotation in a stator and the rotor includes a plurality of cylindrical bores disposed annularly around the axis thereof which are generally parallel to the stator bore. Pistons reciprocate within these bores and connecting rods or links of the pistons engage within the profile of an annular cam secured to the stator so that as the rotor rotates, the pistons reciprocate due to the cam profile. The cycle can be of a diesel type with a fuel injection or of the gasoline/air mixture type with conventional spark plugs. Of importance, is the ease with which the expansion ratio can be made greater than the compression ratio thus util-izing more of the energy normally expelled and wasted in the exhaust gases with the conventional cycles. The thermodynamic cycle in conjunction with the novel engine provides a dwell at the top or combustion stroke to provide a combustion temper-ature of not greater than 4000°C for improved engine efficiency and pollution control. Engines operating in accordance with the present thermodynamic cycle utilize the formula:

where:
k is the ratio of specific heats;
C is the compression ratio;
E is the expansion ratio;
Tb is the temperature at the end of compression;
and Tc is the temperature at the end of combustion and is from about 3000°C to 4000°R.

Description

~Z099~S

BACKGROUND OF THE INVENTION
Conventional engines whether they are of the reciprocating pistons type or the rotary type, utilize the Otto cycle or the diesel cycle or the dual combustion cycle.
These conventional engines suffer from one principal dis-advantage, namely that the expansion stroke is the same length as the compression stroke so that a considerable amount of the energy is wasted and expelled as hot exhaust gases under considerable pressure. Furthermore conventional engines require a separate combustion chamber for each piston.
Engines constructed in accordance with one embodi-ment of the invention utilize a novel thermodynamic cycle that provides a constant volume combustion by employing a dwell at the end of the compression stroke. The dwell of the engine at the top of the compression stroke achieves a combustion temperature of less than 4000~R and is prefer-ably maintained in a range of about 3300 and 3500R that not only results in improved engine efficiency but aLso provides pollution control advantages.
The present invention overcomes the disadvantages and limitations of prior art thermodynamic cycles by utiliz-ing an improved thermodynamic cycle in which the expansion ratio or stroke is longer than the compression stroke thereby , .

~Z~9~ZS

converting some of the energy normally expelled and wasted in the exhaust gases, to useful work or horsepower.
The greater expansion ratio compared to that of the compression ratio is achieved within one cylinder and piston in contrast to the "Brayton" cycle which expands the gases to atmospheric pressure and achieves this cycle with two pistons, one for compression and the other for expan-sion. Engines constructed in accordance with the present invention do not expand pressure to atmospheric pressure after combustion but also utilize a dwell at the end of the compression stroke to result in combustion at a temperature of less than 4000R and preferably in the range of about 3300R to 3500R. The novel thermodynamic cyle is conseq-uently applicable to a variety of engine designs which not only replace the conventional four strokes of an engine into a 360 cycle but which also do not evenly split the ; four strokes into four 90 cycles.
An object of the invention is therefore to provide an improved operating cycle for internal combustion engines ~0 in which the exp~nsion and exhaust strokes are longer than the intake and compression strokes thereby converting more work to useful energy than in conventional operating cycles.
The improved engine and thermodynamic cycle contemplates the ratio of the combination of power and exhaust strokes (PE) to the combination of the intake and compression strokes .

~2(~25 (IC) should not be greater than 4.5 (PE) to (IC) or less than 1 (PE) to 1 (IC).
Another ob~ect of the invention is to provide a ro~ary engine which can be used with a conventional cycle of operation or, can be used with the improved cycle of operation as desired.
Yet another object of the invention is to provide a device of the character herewith described which, when constructed to operate conventionally, can be used as a two or four-stroke engine.
Yet another object of the invention is to pro-vide a device of the character herewithin described which eliminates many of the moving parts normally associated with reciprocating piston type engines.
A further object of the invention is to provide an engine with a mini combustion chamber common to a plur-ality of pistons and cylinders so that continuous combus-tion can take place as the pistons and cylinders rotate past the combustion chamber.
In accGrdance with the invention there is provid-ed an internal combustion engine operating upon an improved thermodynamic cycle having piston strokes of an uneven length and duration in an engine cycle comprising: a sub-stantially cylindrical stator, said stator having bearing ; plates disposed at each end thereof; a drive shaft, said drive shaft being rotatably mounted in said bearing plates, ~'~V9~2~i a rotor disposed within said cylindrical stator said rotor having a plurality of cylindrical bores wherein each of said bores includes a piston disposed ~herein for reciprocal motion;
a ~am ring, said cam ring disposed intermediate said bearing plates and having a programmed cam profile o~ a 360 asymmetrical geometrical configuration to provide a four stroke engine in the 360 profile o~ said cam ring having intake and compression strokes of one length and expansion and exhaust strokes of a different length; a dwell provided on said cam ring profile between said compression stroke and said expansion stroke which moves said piston away from top dead cen.er to maintain a sub-stantially constant combustion temperature of less than 4000R
before initiating said combustion stroke; and means for oper-atively connecting said pistons disposed in said rotor to the pro~ile of said cam ring wherein said cam ring profile in at least a 360 cycle of said engine provides a compression stroke wherein said compression stroke takes U? about 15% to 25% of the engine cycle~ an expansion stroke wherein said expansion stroke is longer than said compression stroke and takes up about 25% to 35% of the engine cycle said expansion stroke terminating when the pressure inside the cylinder is barely sufficient to overcome the friction of the engine, an exhaust stroke said ~xhaust stroke following said expansion stroke wherein said exhaust stroke takes up from about 25% to 35%
of the engine cycle, and an intake stroke wherein said intake stroke takes up from a~out 15% to 35% o the engine cycle;

f~ ,I
,,,, 1 ~2~ ?25 in accordance with another aspect of the invention there is provided a thermodynamic cycle for the conversion of fuel into energy in accordance with the method comprising: introducing fuel into a four stroke engine having a piston cylinder combin-ation having a plurality of rotating cylindrical bores and at least two pistons disposed in each of said cylindrical bores wherein each of said two pistons in each of said cylindrical bores in said engine in at least at 360 cycle provides a combin-ation of intake and comprPssion strokes of one length and an expansion and exhause stroke of a different length in said cycle; limiting said compression stroke to about 15% to 25%
of said engine cycle; providing an expansion stroke wherein said expansion stroke is longer than said compression stroke and takes up about 25V/o to 35% said cycle; utilizing an exhaust stroke wherein said exhause stroke follows said expansion stroke and said expansion stroke terminates when the pressure inside said cylindrical bores is barely sufficient to overcome the friction of said engine and wherein said expansion stroke takes up from about 25% to 35% of said engine cycle; and providing an intake stroke wherein said intzke strokei takes up ~rom about 15% to 35% of said engine cycle; in accordance with a further aspect of the invention there is provided a thermodynamic engine having piston strokes of an uneven length and duration in an engine cycle comprising: a substantially cylindrical stator, said stator having bearing plates disposed at each end thereof;
a drive shaft, said drive shaft being mounted in said bearing ;.

120~ 5 plates; a rotor mounted to said drive shaft, said rotor having a plurality of cylindrical bores wherein each of said bores including a piston disposed therein for reciprocal motion;
a am ring of an asymmetrical configuration, said cam ring disposed intermediate said bearing plates and adjustably s~cured with respect to said stator wherein said cam ring is adjustably secured by a gear mounted on said stator communicating w;th said cam ring to annularly rotate said cam ring; means for operatively connecting said piston disposed in said rotor to the profile of said cam ring and wherein said profile of said cam ring includes a dwell at the end of the compression stroke to move said pistons to a position away from top dead center to provide a combustion temperature in the range of about 3000R to 4000~; fuel and air inlet means disposed on said stator, said fuel and air inlet means intermittently communicat-ing with said cylindrical bores; and a common combustion chamber disposed intermediate said bearing plates whereion said common combustion chamber is of such an angular extent as to overlap two cylindrical bores of said rotor to provide continuous corn-bustion by flame propagation resulting from the intermittent overlapping and the initi~tion of combustion in each of said plurality of cylindrical bores as said pistons and cylinders rotate past said common combustion chamber; in accordance with a still further aspect of the invention there is provided a method of manufacturing novel cams for a four stroke internal combustion engine in a 360 or greater cycle having piston 1~ 25 strokes of uneven length and duration in an engine cycle com-prising: providing a rotary type engine wherein pistons re-ciprocate in cylinders arranged substantially parallel to a po~er shaft; connecting the ends of the pistons to a cam having a profile of a predetermined configuration; forming the con-figuratior. of the cam profile in accordance with the equation;

E l-k nkk = 1 - 1 + [ 1 - ~ ], ck-l C~-l ' 1 TT~ ~
wherein nkk is the air standard thermal efficiency, k is the ratio of specific heat, C is the compression ration, E is the expansion ratio, Tb is the temperature at the end of compres-sion, Tc is the temperature combustion; and operatively connect-ing said engine to said power shaft; and in accordance with further aspect of the invention there is provided a four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel com-prising: a substantially cylindrical stator having bearing plates disposed at each end thereof; a drive shiaft rotatably mounted in said bearing plates; a rotor disposed within said cylindrical stator said rotor having a plurality of cylindrical bores wherein each of said bores includes a piston disposed therein for reciprocal motion; a cam ring disposed intermediate , said bearing plates, said cam ring having a profile programmed in accordance with the equation:

E l-k nkk = 1 - _ 1 + [ 1 - (~) ]
ck-l Ck 1 (1 ~ T~ ) wherein nkk is the air stand thermal efficiency, k is the ratio of specific heat, C is the compresion ration, E is the expansion ratio, Tb is the temperature at the end of compres-sion; Tc is the temperature at the end of combustion and is the range of from about 3~000 R to 4,000R; and means for operatively connecting said pistons disposed in said rotor to the profile of said cam ring.
With the foregoing objects in view, and other ~209~Z5 such objects and advantages as will become apparent to those skilled in the art to which this invention relates as this specification proceeds, my invention includes a novel thermodynamic cycle which is applicable to the arran-gement and construction of parts all as hereinafter will be more particularly described.

Other advantages of the invention will become apparent to those skilled in the art from the following detailed description of the invention in conjunction with the appended drawings in which:
Figure 1 is an isometric partially schematic view of one embodiment of an engine capable of utilizing the novel thermodynamic cycle in operation.
Figure 2 is an end view of Figure 1 with one cylinder head removed.
Figure 3 is a schematic section along the line 3-3 of Figure 2.
Figure 4 is an enlarged fragmentary plan view substantially along the line 4-4 of Figure 3.
Figure 5 is a fragmentary view showing an alter-native construction of the connection of the piston to the cam ring.
Figure 6 is a schematic view of the cycle of oper-.
~, ~, ~, , , ~20~!~25 ation of the engine utilizing the improved thermodynamic cycle.
Figure 7 is an isometric view of one of the con-necting rods showing two alternative connections of the rod to the cam ring as illustrated in Figures 3 and 5.
Figure 8 is an operating diagram of the improved cycle using a carburetor.
Figure 8A is a pressure-volume graph comparing the Otto and Brayton cycles of a normally aspirated engine with the pressure-volume curve of the novel thermodynamic cycle.
Figure ~ is an entropy temperature graph compar-ing the Otto and Brayton cycles of a normally aspirated engine with the entropy temperature curve of the novel thermodynamic cycle.
Figure 9 is a view similar to Figure 8, but showing the cycle used with a diesel operation.
Figure 10 is an isometric, partially sectioned view of an alternative embodiment of the invention.
Figure 11 is an isometric partially sectioned view of a further embodiment of the invention.
In the drawings like characters of reference in-dicate corresponding parts in the different figures.

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120~25 DETAILED DESCRIPTION
Although the drawings illustrate a novel engine utilizing the new cycle of operation in which the expansion stroke is longer than the compression stroke, nevertheless it will be appreciated that the novel engine can be con-structed to operate on a conventional cycle in which the lengths of stroke are equal, under which circumstances pre-ferably only one rotor may be utilized. Furthermore, when used with a conventional cycle of operation, the engine is readily adapted for use with either a two or four-stroke ,i cycle.
The advantages of the new thermodynamic cycle are accomplished in part by modifying the conventional four stroke engine by not only achieving the four strokes in a 360 degree cycle rather than the traditional 720 degree cycle but also by unevenly dividing the four strokes in other than 90 degree per stroke. As a result each of the power and exhaust strokes can and preferably do occupy more than 90 of the cycle and each o~ the intake and compres-sion stxoke5 occupy less than 90 of the 360 cycle.
More particularly the advantages of the present thermodynamic cycle are achieved by increasing the length of the power and exhaust stroke each over 90 of the rota-tion of the engine while reducing the intake and compression ~209~

strokes. In the preferred embodiment the com~ination of power and exhaust strokes (PE) to the combination of intake and compression strokes (IC) should not be greater than 4-1/2 (P~) to 1 (IC) not less than 1 (PE) to 1 (TC~. In addition, as will be discussed hereinafter in greater de-tail, the advantages of the invention are further realized where the length of the expansion stroke is limited by the friction in the engine. More particularly the expansion of the combustion stroke preferably continues until the pres-sure inside the cylinder is barely sufficient to overcome the friction in the engine. As a result the pressure in the cylinder is not allowed to reach atmospheric pressure be-fore initiating the exhaust stroke.
The present thermodynamic cycle in the preferred embodiment employs a dwell near the top dead center at or near the end of the compression cycle immediately prior to ignition to provide a relatively constant ignition. Con-stant ignition occurs at a relatively constant temperature and volume and results in an ignition temperature of less than 4000R and preferably in the range of about 3300R to 3500R to provide an engine that emits less pollutants. It will be understood that constant combustion provided by the thermodynamic cycle reduces the emission of oxides of nitrogen since formation of these pollutants occur at much 12~ 5 higher temperatures that are typically above 4500R.
The dwell of the Piston near top dead center to provide a relatively constant ignition may be achieved in the preferred embodiment by programming the cam profile to hold or dwell the piston at top dead center for a num~er of degrees of rotation to allow the combustion of the mix-ture and/or the rate of injection of fuel to be made with a controlled upper temperature unit or made isothermally or at a relatively constant temperature and volume.
It will be recognized by those skilled in the ar$ that all heat engines in general operate on a given thermodynamic cycle, and that there are various ran~es of values for the pressures, volumes and temperatures on which the engine can be made to operate. Comparisons of the rela-tive performances between actual engines of a given type can be made by defining ratios between the variables and by using these ratioS as the basis for comparison.
The spark ignition "Otto" cycle engines, for ex-ample, use the well known compression ratio, definea as the ratio of the volume of the gases at the beginning of com-pression to the volume of the gases at the end of compres-sion as the basis of predictiny thermal efficiencies, octane requirements, etc. This can be dcne because the ex-pansion ratio is necessarily the same because of the crank . .~ .
,, ~., ~209~ZS

m~chanism used.
The present thermodynamic cycle and the engine configurations which opexate on this cycle, have provided a novel cycle and engine structures for optimizing thermal efficiencies, maximum combustion temperatures, scavenging, etc.
The compression ratio (C) for the engines con-structed in accordance with the invention is identical to that used for the conventional Otto cycle and Diesel cycle engines, and is the ratio of the volume of working fluid at the beginning of compression (Vi) to the volume of the fluid at the end of compression ~Vc). As for example:

C = vi In further discussions of the novel cycle the compression ratio (c) will have the same meaning regarding cylinder pressures, temperatures at the beginning of combus-tion, fuel requirements, etc. as for conventional internal combustion Pngines.
The expansion ratio (E) is not identical to th~
compression ratio as in conventional engines but is in generally much larger, and may vary considerably in engines constructed in accordance with the thermodynamic cycle with identical compression ratios, depending only on the cam profile chosen for that particular engine. E is defined as ~209~25 the ratio of the volume of the working fluid at the end of the expansion to the volume of the working fluid at the beginning of the expansion stroke. The volume of the fluid at the beginning of the expansion process is identical to the volume at the end of compression above, as fox example:
E = Vi Vc The ratio governs the work output of novel en~ines utilizing the present thermodynamic cycle in determining the pressures and temperatures of the exhaust gases, the efficiencies, etc. These factors cannot be predicted from ! the compression ratio alone as in conventional engines.
The compression ratio, C, and the expansion ratio E, as de~ined above, may be combined into an expansion/
compression ratio ~E/C), i.e.

E = Ve / _ Ve C ~ / Vi One of the parameters of engines constructed in accordance with the thermodynamic cycle which is of prime importance in the optimization of the engine for the maxîmum brake thermal efficiency, is the pressure to which the ex-pansion of the gases should be continued in order to "break even" on the conversion of the gas pressure inta work versus the friction losses associated with the increased length of stroke of the pistons, and internal friction of ~2(~9~2S

the engine~ Since this depends also on the CompresSiQn ratio, the combination of the two variables into one ratio is very useful for analysis, particularly under part throttle conditions.
This E/C ratio is also one of the basic consider-ations on which the thermalefficiencies of the engine pri-marily depend. The other consideration particularly impor-tant is the temperature of the formation of oxides of nitrogen whicn can be set at or below 4000R to reduce the formation of oxides of nitrogen.
The scavenging ratio (S), defined as the ratio of the volume of exhaust gases (Vh) left in the head space at the end of the exhaust stroke to the volume of gases in the cylinder (Vi) at the beginning of compression, as for example:
S = Vh/Vi In conventional Otto cycle and in engines using a crankshaft, this ratio is the inverse of the compression ratio. In an engine where the piston motion is controlled by a cam, this ratio may differ as the head space at the end of the exhaust process may be made very small. One of the many advantages to the present thermodynamic cycle is that the reduction of the residual exhaust gases makes it possible to operate engines constructed in accordance with , 1;~09'~25 the thermodynamic cycle at leaner mixtures of the new charge supplied by the carburetor when the engine is idling~
Less fuel, therefore, is required to ensure ignition of mixtures near the "lean limit of inflammability". Further-more, more oxygen can be inducted on the inlet stroke, thus increasing the maximum power output per cylinder and the volumetric efficiency is improved.
Heat engine cycles have formulas which define their air standard thermal efficiencies (nk) in simple terms based on the ideal gas laws and the parameters governing the pressures and values of the working fluids on which the engine operate. By the use of suitable polytropic coefficients, these formulas may be made to predict attain-able indicated thermal efficiencies, which, when corrected for actual conditions in the engine, give reasonable pre-dicated indicated outputs.
As will be understood by those skilled in the art, the thermal efficiency, nkO of the Otto cycle may be devised very simply as follows:

nkO ~ Heat in - Heat out = QH ~ QL
Heat in QH

~Z0~125 - ~3 ~

The heats are theoretically added and rejected at constant volume, hence:

nkO = 1 - QL = 1 - m CV (T4 - T1) (1) v ( 3 2) where m is the mass o~ working fluid, Cv is the specific heat of the fluid at constant volume and TlT2 etc. are the temperatures attained at the various points each of the four strokes of the Otto cycle corresponding to points 102, 104, 106 and 108 on curve 110 in Figure 8B which are similarly noted in prime in Figure 8A.
From the ideal gas laws:

2 (Vl) k - 1 and T3 = (V4~ k - 1 Tl (v2) T4 ~V3) Since Vl + V4 and V2 V3~ _3 = 2 ~r T3 = T4 T2 Tl Rearranging (1):

n = 1 - T4 - Tl = 1Tl (T4_1) = 1 - T
T - T (Tl T2 T2 (T3 ( 2 ~209~2S

and substituting the volume/temperature relationships:
nkO ~ 1 - (V2) k - 1 (Vl) = 1 - 1 (Vl) k ~ 1 (v2) Where C = Vl = Compression ratio of the engine n = 1 ~
ck - 1 In the pressure limited Diesel cycle, the cycle on which most modern Diesel engines operate, heat is added at constant volume until some arbitrarily assigned pressure is reached, and the remainder of the heat is added at con-stant pressure. The air standard thermal ef~iciency is given by:

nkd k - 1 . ~ C~ 3 ~ p(rCP - 1) + r Where rp = Pc = limited pressure ratio Pb r p = Vd = Volume ratio after which adiabatic Vc expansion occurs and r = compression ratio of the engine as defined.

.

~209~25 The air standard thermal efficiency of the Brayton cycle on which gas turbines operate is defined by the standard equation:

n = 1 - P (l-k) kb r ( k ) Where Pr is the pressure ratio of the pressure after the end of compression divided by the pressure at the beginning of compression. The temperatures of the Brayton cycle are illustrated as points 112, 114, 116, 118 of curve 120 in Figure 8A which are similarly noted in primed numbers in Figure 8B.
The novel thermodynamic cycle is illustrated by line 140 in Flgure 8B with the temperatures illustrated as points 122, 124, 126, 128 and 130 with these points simil-arly referenced with primed numbers in Figure 8A.
The air standard thermal efficiency, nkk f engines constructed in accordance with the invention utiliz-ing the novel thermodynamic cycle may be expressed as:

n = Net work out kk Heat in n~k = Heat in - Heat rejected at D+work of expan-sion D-E
Heat in Heat additi~ns or rejections, Q = m C~ (Tl - T2) ~209~25 and work of expansion, W = PlVl P2V2 k - 1 for closed systems, i.e. pistons acting in cylinders.
mCv (Tc - Tb) - mCv (Td - Ta) + (PdVd - PeVe) n. = J(k - 1) kk mCv (Tc - Tb) Where m is the mass of working fluid Cv is the specific heat at constant volume k is the ratio of specific heat J is the mechanical equivalent of heat The respective terms may be conveniently separated.
(PdVd - PeVe) n k = Tc Tb - Td - Ta + Jtk-l) k Tc - Tb Tc - Tb mCv(Tc - Tb) n = 1 - Ta(Td/Ta - 1) + (PdVd - PeVe) kk Tb(Tc/Tb - 1) J(k - l)mCv(Tc - Tb) Since Tb = (Va)k -lTc = Vd k ~ 1 1 AN~ Vb = Vc, Vd = Va Ta (Vb) Td Vc Td = Tc -Ta Tb nkk = 1 - Ta + (PdVd - PeVe) Tb J(k - 1) mCv (Tc - Tb) nkk = 1 - 1 + (PdVd - PeVe) Va k-l J(k - 1) mCv ~Tc - Tb) (--) I
Vb Va = C, the compression ratio of the engine Vb " . - .

1209~2~i Pe = (Vd)k Pe = Pd (Vd)k Pd (Ve) (Ve) (Vd)k nkk = 1 - lk_l ~ PdVd - PdVe Ve (C) JmCv (k - l)(Tc - Tb) nkk = 1 - 1 k-l + Pd Vd (1 - (Vd)k 1 C (Ve) JmCv (k-l) (Tc-Tb) PdVd = nRTd and dividing numerator and denominator by Ta, (Vd) k-l T = 1 - 1 + nR Td/Ta (1 - Ve k k-1 (C) JmCv(k-l)(Tc - Tb Ta Ta Term in brackets, (Tc-Tb) = (Tc x Tb-Tb) = Tb (Tc 1) Td = Tc (Ta Ta) (Tb Ta Ta) Ta (Tb Ta Tb (Vd)k-n = 1 - 1 + nR (1 - Ve k C)~-l JmCv(k-l)Tb(l _ Tb Ta Tc Cv(k-l) = Cv(CP - 1) = Cp - Cv = R
Cv Consistent with the foregoing a general formula for engines constructed in accordance with the preferred embodiment of the thermodynamic cycle is as follows:
E l-k nk = 1 - 1 + ( 1 - C
~ C ( 1 - T~ ) .....
~,. . .
., ~%O'~Z5 where:
nk is the air standard thermal efficiency;
k is the ratio of specific heat;
C is the compression ratio;
E is the expansion ratio;
Tcis from about 3000 to 4000R; and Tb is the temperature at the end of compression.
It will be recognized by those skilled in the art from the foregoing discussion of the air standard thermal effi-ciency of engines constructed in accordance with the present thermodynamic cycle that when:
E = 1, 1 - (E) K - 1 = 0 C (C) and nk = 1 which is identical to the thermal efficiency of the Otto cycle.
The third term of the equation represents the increase in nk over nO. In addition nk is dependent on the heat added, repre-sented in the formula by the temperature ratio Tc/Tb and as heretofore discussed Tc should be in the range of about 3300R
to 4000R to provide pollution control advantages.
Figures 8 and 8A also illustrate that the usable E ratio is also determined by tne heat added i.e. E cannotC C

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1~09~2S

be greater than F/C without have Pe become lower than Pa (atmos-heric pressure). In the actual throttle settings, a relief port located on the expansion stroke and fitting with a one-way valve, allows air to enter the cylinder if and when the pressure drops to below atmospheric pressure.
E l-k nk = 1 - 1 + ( 1 - C
~ C~ 1(1 _ Tb) As the heat added decreases, the term (1 ~ T~) approaches zero, and the fraction increases. The efficiency of the novel engine, therefore, increases at part load until the limiting value of the usable E/C ratio is reached. In the limit, however, as E/C approaches 1, the numberator also approaches zero, and the expression becomes indeterminate.
In addition the third term 1 of the equation shows that for a constant E/C ratio, the efficiency of the K-Cycle in-creases as the compression ratio C increases.
The advantages of the thermodynamic cycle are opti-mized by employing a constant volume combustion by providing a dwell at the end of the compression stro~e to utilize a fairly consistent combustion temperature of under 4000R and preferably in the range of 3300R to 3500R. As used herein the term dwell refers to a means of maintaining a constant volume or a slight modification of volume to achieve and maintain a constant tem-,.;, ~"-''.

%5 perature. A modification of the basic cycle differs in that only part of the heat is added at constant volume. After the desired temperature of the working fluid is reached, the remain-der of the heat is added isothermally (at the same temperature) while the volume increases. The cycle is illustrated in Figure 8B.
The heat in Qin= MCv (Tc - Tb) + M R Tc log n (Vd) (Vc) The work out = the work of expansion, minus the work of com-pression.
Qout = M R Tc log n (Vd) + J (Pd Vd - Pe Ve) - J (PbVb PaVa) (Vc) k - 1 k - 1 From the relationship, Pv = M R T

~2~9!~25 Qout = M R Tc log n tVd) + M J R ~Td-Te) M J R ~Tb-Ta) The air standard thermal efficiency, nkt is given by:

n = MC (Tc - Tb) - M J R (Td - Te) + MJR (Tb - Ta) kt v k - 1 k - 1 . . . _ _ . _ _ . _ _ . _ . . .
M Cv (Tc - Tb~ - M R Tc log n (Vd) From the foregoing it will be recogni~ed that the temperature limiting formula is as follows.
n~t = C~ (Tc - Tb) - JR (Td - Te~ -(Tb - Ta) Cv (Tc - Tb) + R Tc log n (Vd) Other forms are possible, but in the analysis of the engine process, Volume Vd is-dependent on the limiting temperature Tc which are heretofore described is preferably in the range of 3300R to 3500~. It will also be recognized by those skilled in the art that as the fuel supply is reduced in the throttling process of the actual engine, Tc and Te will vary also, hence no simple formula exists which will give the efficiency in terms of the ~ngine parameters.
In general, engines using this temperature-limited cycle will have greater E/C ratios and will have slightly lower indicated thermal efficiencies than engines where all of the heat is added at constant volume.
Also to be appreciated is the fact that the novel 12~)9'.9;2~

cycle described herein can readily be used with other con ventional rotary type engines whether these engines are rotary piston type or not.
With a conventional opposing piston engine, a cam-type crank shaft can be utilized so that the expansion stroke is longer than the compression stroke and by modifying the lobe of a rotary engine such as the Wankel, a similar effect can be obtained.
The novel cycle described herein may be defined as a cycle based on the Otto, Diesel or Dual Combustion cycle, but having an expa~sion ratio greater than the com-pression ratio, said ratio being less than that required to expand the gases to atmospheric pressure. This cycle is achieved within one cylinder or chamber and is-a geometrical and volumetrical ratio in which a motion of a chamber, or cylinder and piston, produces a geometrical ratio which theoretically is the same as the volumetrical ratio. This eliminates dead motion as in the case of some engines which reduce the volumetrical compres5ion ratio in comparison to the geometrical ratio by either early or late intake valve closing.
Proceeding therefore to describe the invention in detail, reference should first be made to Figures 8 and 9 209~25 - ~3 -which show pressure diagrams for the new cycle using a car-buretor in the case of Figure 8 and a diesel cycle in the case of Figure 9.
The line between positions 1 and 2 illustrates the compression stroke, and between 2, 3 and 4, the conven-tional expansion stroke at which time the exhaust valve normally opens.
In the present cycle, the expansion stroke ex-tends from position 2 through 3, 4 and to position 5 thus utilizing more of the power developed by the fuel than heretofore.
Approximately 30~ more power can be utilized at all power setting thus increasing the efficiency of the ~engine whether using carburetor, diesel or dual combustion type cycles of operation.
Figures 1 to 7 show one embodiment of a novel en-gine which may utilize this cycle although of course, it will be appreciated that a conventional cycle can be used, Proceeding to describe the engine in detail, re-ference to the accompanying drawings will show a substantially cylindrical stator collectively designated 10 which is pro-vided with a cylindrical chamber 11 formed therethrough.
Each end of this cylindrical chamber is closed by by means of an engine end plate 12 secured by bolts 13 with 1209~9~5 a conventional type seal 14 being provided between the cy-linder head 12 and the cylindrical stator 10.
A cylindrical rotor 15 is mounted for rotation within each end of the stator 10 and secured to a cornmon shaft 16 which in turn is bearably supported within bearings 17 provided centrally of each cylinder head 12.
Each rotor 15 is provided with a plurality of piston bores 18 equidistantly spaced around the axis of the rotor in an annular ring as clearly illustrated in Figure 2 and a piston 19 is provided for each bore and is reciprocal therein, conventional piston rings 20 being pro-vided as shown.
The rotors 15 are situated at each end of the stator as hereinbefore described, with the outer ends 21 in bearing contact with the cylinder heads and sealed by means of annular seals 22 which are shown schematically in Figure 3. Howe~er, these seals are preferably labyrinth.
type seals which are well known so that it is therefore not believed necessary to describe same further. However, op-tional radial seals 22A may be incorporated between each cylinder (see Figure 2).
Annularly formed fluid passages ~3 may be provided in each of the rotors and connected to an external source for cooling purposes.

.

lZ0~25 A pair of cam rings 24 are provided intermediate the ends of the sta~or 10 and are secured around the wall of the cylindrical chamber 11 by means of bolts 25 screw threadably engaging the cam rings 24 through elongated slots 26 in the wall of the stator and these stators may be rotated within limits, for purposes hereinafter to be described, by any convenient means. In the present embodiment, such means includes gear teeth 27 formed around part of the outer periphery of the two cam rings 24 engageable by a gear 28 mounted upon a shaft 29 which in turn may be rotated through gear 30 from any convenient location so that rotation of shaft 29 will move the cam rings 24 annularly within limits.
Each cam ring is U-shaped when ~iewed in cross section and reference should be made to Figure 3. Each cam ring includes a base 31 with a pair of upstanding legs 32 extending at right angles and each of these legs is provided with an annular channel 33 formed on the inner wall thereof as clearly shown. This annular channel rises and falls axially through 360, the profile of the channel being shown specifically in Figure 6 and this channel forms the profile of the two cam rings r one being a mirror image of the other as clearly illustrated in Figure 6.
Means are provided to connect the pistons 19 with the cam rings, said means taking the form of a connecting .

)9~2~

rod or link shown specifically in Figure 7.
The inner end 34 of the connecting rod means 35 is provided with a wrist pin 36 which pivotally connects the inner end to the piston 19 through bosses p.rovided ~not illustrated) in the usual manner.
The lower end 37 of the link or connecting rod means 3S is wider tha.n the inner end 34 and is provided with rollers upon either side thereof.
Figure 3 shows one embodiment of these rollers in which a relatively large roller 38 is journalled for rota-tion upon a pin 39 extending from either end of the portion 37 and these relatively large rollers are in rolling con-tact with one wall 40 of the annular channel 33.
A smaller roller 41 is also journalled for rota-tion upon the end of pin 39 and is in contact with the other wall 42 forming the annular channel 33 so that these two rollers anchor the connecting rod means 35 within the profile of the cam ring defined by the annular channel 33, Alternatively, a single relatively wide roller 43 can be journalled for rotation upon the end of pin 39 and engage within a modified channel shown in Figure S.
In either embodiment, rotation of the rotor within the stator will cause the rollers on eac~ piston to roll arouna the cam profile and the shape of this profile causes these pistons to reciprocate within the piston bores 18 in 1%~9~25 a clearly defined sequence.
If the profile of the cam ring is symmetrical then the length of the stroke of the pistons 19 will be the same, but if the profile is modified as illustrated in Figure 6, then the length of the stroke of the pistons while travelling around one portion of the cam profile will be different from the length of the stroke of the piston tra-velling around the remainder portion of the profile.
For example, by programming the profile, intake and compression strokes can be initiated by each piston to-gether with expansion and exhaust strokes through one rota-tion of one piston (360) and the shape of the profile will cause the intake and compression strokes to be shorter than the expansion and exhaust strokes.
By providing two rotors with two sets of pistons and two cam rings, one being a mirror image of the other, a balance is achieved and vibration is reduced. Two such systems are shown in Figure 6 schematically.
A relatively small arcuately curved firing chamber 44 is formed in each of the cylinder heads and at this loca-tion either a spark plug 45 or a fuel injection nozzle 46 (Figure 1) is provided depending upon the system being used.
An intake port 47 extends through the cylinder head together with an exhaust port 48 and these are shown ... .. .

~20~25 schematically in Figure 2 and Figure 6.
Figure 6 shows one cycle of two opposed pistons specifically designated l9A in Figure 6. The cycle extends through 360~ or one revolution of the two rotors 15.
At the first position (0) the pistons are at approximately top dead center and are moving inwardly as they pass the intake conduit or port 47. Assuming an engine using a carburetor, the pistons will draw in a gasoline/air mixture during the intake stroke through approximately 90.
As the pistons l9A pass or close off the intake ports 47, they commence moving outwardly and compress the gases between the pistons head and the cylinder head for a stroke having the same length as the intake stroke. As they reach one end of the small combustion chamber 44, spark plug 45 is fired thus igniting the mixture which causes an expansion stroke and causes the pistons to move inwardly through approximately a further 90. However, due to the shape of the cam profiles, this expansion stroke is longer than the compression stroke by an amount indicated in Figure 6 by reference character 49 and as the pistons approach the innermost position at apprQximately 470, the exhaust port 48 is reached and the spent gases are expelled through the exhaust port during the exhaust stroke which is the same length as the expansion stroke. Due to the longer ex-......... ~, .. . . . ~ . ... ~ .. . ... . .

~2~9~Z5 pansion and exhaust strokes, these strokes will exceed 90 rotary travel and the intake and compression strokes will be less than 90~ rotary travel.
Each succeeding pistons follows the same path so that there is a relatively continuous firing and the combustion chamber 44 spans or overlaps two or more cylinders.
This serves two purposes. First, to stratify the charge with a carburation type engine and secondly, to assist in continuous burning with the injection type engine. Both systems serve to assist in burning a lean mixture and there-by assist further in meeting present day emission standards.
The thermal efficiency is increased considerably with the no~el cycle and the-thermal efficiency of conven-tional cycles is expressed in the formula:
e = 1 _T4 Tl for Otto and Brayton cycle e = 1 - 1 (T4-Tl) for the Diesel cycle (T4-Tl) for the Dual Combustion e = 1 (T2'-T2~ + 1 (T3-T2 ) cycle~

In the above formula, Tl = Temperature of the intake air.
T2 - Temperature of compressed air.
T3 = Temperature of the combusted mixture T4 = Temperature of the Exhaus~ gases 12(~ S

T5 - Temperature of the Exhaust gases in the new cycle.
It is believed that the exhaust temperature with improved cycle is approximately 1000R (Rankin~ lower than the Otto or Diesel cycle so that by substituting this new figure (T5) for T4 in the above formula, the thermal effic-iency in the new cycle is approximately 30% higher or more.
As mentioned previously, Figure 6 shows a new cycle in which the expansion exhaust strokes are consider-ably longer than the intake and compression strokes, but of course, it will be appreciated that by shaping the cam pro-files, the strokes can be made the same length.
The aforementioned gear 28 may-be used to rotate the cams slightly with reference to the position of the combustion chambers 44 thus varying the timing of the ignition or fuel injection.
By the same token, by changing the relationship between the two cams axially, as by inserting or witharawing shims 50 ~etween the two cams, the compression ratio may be varied within limits. This also may be done while the engine is running by means of simple linkage and movable shims (not illustrated).
The preceeding description covers the embodiment shown in Figures 1 through 5 and Figure 7, in which the cam ~2~9~25 rings 24 are situated centrally of the stator 10 and the cylinder heads 12 are situated at each end thereof.
However, it will of course be appreciated that the positions of these parts may be reversed with the cam rings situated adjacent the ends of the stator and a common cylinder head being situated intermediate the ends of the stator and such as design is shown in Figure 10. Where applicable, similar reference characters have been given, but with a prime distinguishing them from the reference characters of the other views.
The cylindrical stator 10' is provided with a cylindrical chamber 11' formed thereby.
End plates 51 are secured to each end of this cylindrical stator by means of bolts 13' and bearing assem-blies 52 are provided centrally within the end plates 51 to support for rotation, a common shaft 16'.
A cylindrical rotor collectively designatea 53 is secured to shaft 16' as by splines or the like (not illustrated), there being a pair of rotors provided in this embodiment one upon each side of a common cylinder head 54.
This cylinder head is secured within the stator 10' centrally thereof and spans the cylindrical chamber 11' as clearly il-lustrated. It is provided with a spark plug such as indi-cated by reference character 55 although/ if desired, fuel injection means may be provided at this point. The spark plug or fuel i~ection means communicates with a small com-bustion chamber 56 formed through the cylinder head 5~ so that it communicates with the rotor 53 on either side vf the cylinder head.
Exhaus~ means 57 communicate with a common exhaust exhaust port 5~ within the cylinder head 54 and an air in-take or air/fuel intake is provided on the opposite side of the cylinder head (not illustrated) similar to the intake 47 of the previous embodiment.
From the foregoing it will be appreciatea that the combustion chamber 56 is common to both rotors 53, together with the exhaust and intake port means.
Each rotor 53 is mounted upon shaft 16' as herein-before described and includes a centrally located support plate 58.
Cylinder bores 59 are formed annularly within a block 59A extending upon one side of plate 58 and being secured thereto and these bores are parallel to the axis of th~ shaft 16'.
A support cylinder 60 is secured to the other side of plate 58 on each in alignment with the bores 59 and these support cylinders are provided with longitudinally extending channels 61 upon each side thereo within which is supported ~2~9~5 or reciprocation, a bifurcated end 52 of a connecting rod 63 which extends from the underside of each piston 64 which in turn reciprocates in each of the bores 59. The rod 63 and the piston 64 are preferably formed from one piece and the support cylinders 6Q are bifurcated as illu5-trated at 65 for reasons which will hereinafter be des-cribed.
Alternatively, the cylinders and support cylinders 60 may be formed of hollow cylindrical shells clamped bet-ween upper and lower plates, but as this is considered to be an obvious alternative, it is not believe~ necessary to describe same further.
However, it should be observed that the claims are intended to cover both structures.
A cam ring 66 is secured within each end of the stator 10' adjacent the end plates ~1 and this cam ring includes the support base portion 67 and a T-shaped portion collectively designated 68. This T-shaped portion includes a web 69 and a flange 70 extending upon each side of the edge of the web as clearly shown in Figure lO and the bifurcated formation of the support cylinders 60 permit rotation of the rotor around the cam ring with a portion of the web and the flange 70 being situated within the bifurcated slots 65.
Means are provided to connect the ends 62 of the , .

:~209~25 connecting rod to the cam rings and in this connection these ends 62 are bifurca~ed to form two side portions 71 between which is situated a first roller means 72 support~d for rotation upon a pin 73 extending between the bifurcated por-tion 71.
This roller rides upon the outer surface 74 of the flanges 70 as clearly shown.
A pair of smaller rollers 75 are journalled for rotation upon a pin 76 also extending between the sides of the bifurcated ends 71, but being spaced from the main or first roller 72 and each of these small rollers 75 engages the inner surface of the flange 70, one upon each side of the web 69 thus ensuring that the roller 72 follows the con-tour of the T-shaped portion of the cam in a manner similar to that described for the previous embodiment.
The operation of the embodiment of Figure 2 is similar with the exception that the common combustion cham-ber 56 communicates with the cylinders of each rotor, it being understood that a pair of pistons 64 are in opposition at the time of firing so that the expansion stroke reacts upon both pistons which are at the com~ustion chamber location.
By the same token, exhaust and intake feed the cylinders of each rotor as they pass thereby.
Once again the cam rings 66 may be rotated slightly l20s~2s to alter the timing and the position of the cam rings re-lative to the end plates 51 that may be varied by shims ~not illustrated) or other similar means so that the compression ratio may be varied within limits.
Although both embodiments illustrate and describe opposed rotor assemblies within a common stator, neverthe-less it will be appreciated that a single rotor can be used with a single cam ring and single cylinder head although, under these circumstances, it is desirable that the length of all strokes are equal in order to reduce vibration.
Figure 11 illustrates an engine constructed in accordance with the invention utilizing a single rotor 53 which like embodiments heretofore described may operate in accordance with the preferred thermodynamic cycle. As will be recognized by those skilled in the art common com~ustion chamber 56 may be of sufficient radial extent such as to overlap two bores of lateral pistons to provide continuous flame propagation as the cylinders rotate past the eombustion chamber such that the spark plug is necessary only to initiate combustion.
In Figure 11 reference number 80 is illustrated one disposition of a relief port having a one-way valve which is designed to admit air into the cylinder if pressure within the cylinder drops below atmospheric pressure. As has here-~209~5 tofore been described, relief port 80 may be disposed in the various engine embodiments as is illustrated in Figures 1, 10 and 11 to assure that the efficiency of the engine is not impeded by allowing the piston to attain a lower than atmos-pheric pressure in the piston bores 18 after combustion.
As w.ill be recognized the precise disposition and number of relief ports along with the programming of the cam to allow the power stroke to continue until pressure inside the cy-linder is sufficient to overcome friction should be disposed at or near the end of the combustion cycle.
Engines constructed in accordance with the inven-tion utilize a programmed cam profile wherein the compression stroke takes up from about 1$% to 25% of the engine cycle, the expansion stroke that is longer than the compression stroke and takes up about 25% to 35% of the engine cycle, an exhaust stroke that takes up from about 25% to 35% of the engine cycle and an intake stroke from about 15% to 35~ of the engine cycle. Optimization of the engine generally results when the compression stroke takes up about 19% of the engine cycle and the expansion stroke takes up about 33% of the engine cycle, the exhaust stroke takes up about 33% or the engine cycle and the intake stroke takes up about.19% of the engine cycle.

~2~2S

Since various modifications can be made in my invention as hereinabove described, and many apparently widely different embodiments of same made within the spirit and scope of the claims without departing from such spirit and scope, it is intended that all matter contained in the accompanying specification shall be interpreted as illus-trative only and not in a limiting sense.

Claims

CLAIMS:
(1) An internal combustion engine operating upon an improved thermodynamic cycle having piston strokes of an uneven length and duration in an engine cycle comprising:
(a) a substantially cylindrical stator, said stator having bearing plates disposed at each end thereof;
(b) a drive shaft, said drive shaft being rotat-ably mounted in said bearing plates;
(c) a rotor disposed within said cylindrical stator said rotor having a plurality of cylindrical bores wherein each of said bores includes a piston disposed there-in for reciprocal motion;
(d) a cam ring, said cam ring disposed intermed-iate said bearing plates and having a programmed cam pro-file of a 360° asymmetrical geometrical configuration to pro-vide a four stroke engine in the 360° profile of said cam ring having intake and compression strokes of one length and expansion and exhaust strokes of a different length;
(e) a dwell provided on said cam ring profile be-tween said compression stroke and said expansion stroke which moves said piston away from top dead center to main-tain a substantially constant combustion temperature of less than 4000°R before initiating said combustion stroke;
and (f) means for operatively connecting said pistons disposed in said rotor to the profile of said cam ring where-in said cam ring profile in at least a 360° cycle of said engine provides a compression stroke wherein said compres-sion stroke takes up about 15% to 25% of the engine cycle, an expansion stroke wherein said expansion stroke is longer than said compression stroke and takes up about 25% to 35%
of the engine cycle said expansion stroke terminating when the pressure inside the cylinder is barely sufficient to overcome the friction of the engine, an exhaust stroke said exhaust stroke following said expansion stroke wherein said exhaust stroke takes up from about 25% to 35% of the engine cycle, and an intake stroke wherein said intake stroke takes up from about 15% to 35% of the engine cycle.
(2) The internal combustion engine operating upon an improved thermodynamic cycle having piston strokes of an uneven length and duration in an engine cycle of Claim 1 wherein said rotor is mounted to said drive shaft and said cam ring is disposed intermediate said bearing plates and adjustably secured with respect to said stator wherein said cam ring is adjustably secured by a gear mounted on said stator communicating with said cam ring to annularly rotate said cam ring.
(3) The internal combustion engine operating upon an improved thermodynamic cycle having piston strokes of an uneven length and duration in an engine cycle of Claim 2 wherein said compression stroke takes up about 18% of the engine cycle, said expansion stroke takes up about 32% of the engine cycle, said exhaust stroke takes up about 32% of the engine cycle and said intake stroke takes up about 18%
of said engine cycle.
(4) The internal combustion engine operating upon an improved thermodynamic cycle having piston stroke of an uneven length and duration in an engine cycle of Claim 1 wherein said profile of said cam ring provides a compression stroke having a constant volume combustion by employing a dwell on said cam profile at the end of said compression stroke to maintain said pistons at a position away from top dead center to provide a combustion temperature in the range of about 3000° to 4000°R.
(5) The internal combustion engine operating upon an improved thermodynamic cycle having piston strokes of an uneven length and duration in an engine cycle of Claim 4 wherein said profile of said cam ring is further modified to provide a dwell at the end of the compression stroke to vary combustion volume to achieve a combustion temperature of less than 4000°R.
(6) The internal combustion engine operating upon an improved thermodynamic cycle having piston stroke of an uneven length and duration in an engine cycle of Claim 5 wherein said profile of said cam ring is further modified to provide a dwell at the end of the compression stroke to further vary combustion volume to achieve a combustion temperature in the range of 3300°R to 3500°R.

(7) The internal combustion engine operating upon an improved thermodynamic cycle having piston strokes of an uneven length and duration in an engine cycle of Claim 2 further comprising a relief port disposed on said cylindrical stator having a one way valve to allow ambient air in said cylindrical bores if the pressure inside said cylindrical bores drops below atmospheric during said ex-pansion stroke.
(8) The internal combustion engine operating upon an improved thermodynamic cycle having piston strokes of an uneven length and duration in an engine cycle of Claim 7 wherein said relief port is disposed on said stator at a position corresponding to about the maximum length of said expansion stroke.
(9) A thermodynamic cycle for the conversion of fuel into energy in an internal combustion engine having piston strokes of an uneven length and duration in an en-gine cycle wherein fuel is burned in accordance with the method comprising:
(a) introducing fuel into a four stroke engine having a piston cylinder combination having a plurality of rotating cylindrical bores and at least two pistons dis-posed in each of said cylindrical bores wherein each of said two pistons in each of said cylindrical bores in said en-gine in at least at 360° cycle provides a combination of intake and compression strokes of one length and an expan-sion and exhaust stroke of a different length in said cycle;
(b) limiting said compression stroke to about 15%
to 25% of said engine cycle;
(c) providing an expansion stroke wherein said expansion stroke is longer than said compression stroke and takes up about 25% to 35% said cycle;
(d) utilizing an exhaust stroke wherein said ex-haust stroke follows said expansion stroke and said expan-sion stroke terminates when the pressure inside said cy-lindrical bores is barely sufficient to overcome the fric-tion of said engine and wherein said expansion stroke takes up from about 25% to 35% of said engine cycle; and (e) providing an intake stroke wherein said in-take stroke takes up from about 15% of said 35% of said engine cycle.
(10) The thermodynamic cycle for the conversion of fuel into energy in an internal combustion engine hav-ing piston strokes of an uneven length and duration in an engine cycle of Claim 9 further comprising the step of mov-ing the pistons to a position away from top dead center by means of a dwell to provide a combustion temperature in the range of about 3000° to 4000°R.
(11) The thermodynamic cycle for the conversion of fuel into energy in an internal combustion engine hav-ing piston strokes of an uneven length and duration in an engine cycle of Claim 9 wherein the ratio of the sum of the percentages of power and exhaust strokes to the sum of the percentages of intake and compression strokes is not greater than from about 4.5 to 1 and not less than 1 to 1.
(12) The thermodynamic cycle for the conversion of fuel into energy in an internal combustion engine having piston strokes of an uneven length and duration in an en-gine cycle of Claim 10 wherein said step of moving said pistons to a position away from top dead center maintains said pistons at a further position away from top dead center to provide a combustion temperature in the range of about 3300°R to 3500°R.
(13) An internal combustion engine having piston strokes of an uneven length and duration in an engine cycle comprising:
(a) a substantially cylindrical stator, said stator having bearing plates disposed at each end thereof;
(b) a drive shaft, said drive shaft being mounted in said bearing plates;
(c) a rotor mounted to said drive shaft, said rotor having a plurality of cylindrical bores wherein each of said bores including a piston disposed therein for re-ciprocal motion;
(d) a cam ring of an asymmetrical configuration, said cam ring disposed intermediate said bearing plates and adjustably secured with respect to said stator wherein said cam ring is adjustably secured by a gear mounted on said stator communicating with said cam ring to annularly rotate said cam ring;
(e) means for operatively connecting said pistons disposed in said rotor to the profile of said cam ring and wherein said profile of said cam ring includes a dwell at the end of the compression stroke to move said pistons to a position away from top dead center to provide a combus-tion temperature in the range of about 3,000°R to 4000°R;
(f) fuel and air inlet means disposed on said stator, said fuel and air inlet means intermittently com-municating with said cylindrical bores; and (g) a common combustion chamber disposed inter-mediate said bearing plates wherein said common combustion chamber is of such an angular extent as to overlap two cy-lindrical bores of said rotor to provide continuous com-bustion by flame propagation resulting from the intermit-tent overlapping and the initiation of combustion in each of said plurality of cylindrical bores as said pistons and cylinders rotate past said common combustion chamber.
(14) The internal combustion engine having strokes of an uneven length and duration in an engine cycle of Claim 13 further comprising exhaust means located on said stator.
(15) The internal combustion engine having pis-ton strokes of an uneven length and duration in an engine cycle of Claim 13 wherein said cylindrical bores have their axes substantially parallel to the axis of said rotor and are disposed annularly around the axis of said rotor.
(16) The internal combustion engine having pis-ton strokes of an uneven length and duration in an engine cycle of Claim 13 further comprising a relief port having a one way valve to allow ambient air to enter the cylinder if the pressure inside the cylinder drops below atmospheric dur-ing the expansion stroke.
(17) The internal combustion engine having pis-ton strokes of an uneven length and duration in an engine cycle of Claim 16 wherein said relief port is disposed on said stator corresponding to the maximum length of said ex-pansion stroke.
(18) The internal combustion engine having piston strokes of an uneven length and duration in an engine cycle of Claim 13 wherein said profile of said cam ring further provides:
a compression stroke wherein said compression stroke takes up from about 15% to 25% of the engine cycle, an expansion stroke wherein said expansion stroke is longer than said compression stroke and takes up from about 25% to 35% of the engine cycle, an exhaust stroke wherein said exhaust stroke follows said expansion stroke and said expansion stroke terminates when the pressure inside the cylinder is barely sufficient to overcome the friction of the engine and where-in said exhaust stroke takes up from about 25% to 35% of the engine, and an intake stroke wherein said intake stroke takes up from about 15% to 35% of the engine cycle.
(19) The internal combustion engine having pis-ton strokes of an uneven length and duration in an engine cycle of Claim 18 wherein said compression, expansion, ex-haust and intake strokes are provided in a 360° cycle.
(20) The internal combustion engine having pis-ton strokes of an uneven length and duration in an engine cycle of Claim 19 wherein said compression stroke takes up about 18% of the engine cycle, said expansion stroke takes up about 32% of the engine cycle, said exhaust stroke takes up about 32% of the engine cycle and said intake stroke takes up about 18% of said engine cycle.
(21) The internal combustion engine having pis-ton strokes of an uneven length and duration in an engine cycle of Claim 13 wherein the ratio of the percentages of power and exhaust strokes to the sum of the percentages of intake and compression strokes is not greater than from about 4.5 to 1 and not less than 1 to 1.
(22) The internal combustion engine having pis-ton strokes of an uneven length and duration in an engine cycle of Claim 21 wherein said compression stroke employs constant volume combustion by employing a dwell on said cam profile at the end of the compression stroke to move and maintain said pistons at a position away from top dead cen-ter to provide a combustion temperature of less than 4000°R.

(23) A four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel comprising:
(a) a substantially cylindrical stator having bearing plates disposed at each end thereof;
(b) a drive shaft rotatably mounted in said bear-ing plates;
(c) a rotor disposed within said cylindrical stator said rotor having a plurality of cylindrical bores wherein each of said bores includes a piston disposed there-in for reciprocal motion;
(d) a cam ring disposed intermediate said bear-ing plates, said cam ring having a profile programmed in accordance with the equation:

wherein nkk is the air standard thermal efficiency, k is the ratio of specific heat, C is the compression ratio, E
is the expansion ratio, Tb is the temperature at the end of compression, Tc is the temperature at the end of combustion and is the range of from about 3,000°R to 4,000°R; and (e) means for operatively connecting said pistons disposed in said rotor to the profile of said cam ring.
(24) The four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 23 wherein Tc is achieved by employing a dwell on the cam profile.

(25) The four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 24 wherein said dwell on said cam profile achieves a Tc in the range of about 3300°R to 3500°R.
(26) The four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 24 wherein said pistons have a common combustion chamber of suffi-cient radial extent as to overlap two bores of radially ad-jacent pistons to provide continuous flame propagation as the cylinders rotate past said common combustion chamber.
(27) The four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 23 further comprising:
(f) a second piston disposed in each of said cy-lindrical bores for reciprocal motion;
(g) a second cam ring disposed intermediate said bearing plates, said second cam ring having a profile form-ed in accordance with the equation:

wherein nkk is the air standard thermal efficiency, k is the ratio of specific heat, C is the compression ratio, E is the expansion ratio, Tb is the temperature at the end of compression, Tc is the temperature at the end of combustion and is in the range of from about 3,000°R to 4,000°R; and (h) a second means for operatively connecting said second piston disposed in each of said cylindrical bores in said rotor to the profile of said second cam ring.
(28) The four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 27 wherein said pistons have a common combustion chamber of such a radial angular extent as to overlap two cylindrical bores of said rotor to provide continuous combustion by inter-mittently overlapping cylindrical bores as said pistons and cylinders rotate past said common combustion chamber.
(29) The four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 28 wherein flame propagation is provided in a serpentine type combus-tion between axially adjacent cylindrical bores.
(30) The four stroke internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 28 further comprising a relief port having a one way valve to allow ambient air to enter the cylindrical bores if the pressure inside said cylindrical bores drops below atmospheric pres-sure during said expansion stroke.

(31) The internal combustion engine having a thermodynamic cycle capable of producing reduced levels of pollutants in burning fuel of Claim 30 wherein said relief port is disposed on said stator at a position correspond-ing to about the maximum length of said expansion stroke.