Note: Descriptions are shown in the official language in which they were submitted.
~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
-- 2
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
--3--
(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
- 3A -
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
- 3B -
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
- 3C -
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
- 3D -
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:
- 3E -
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.
DESCRIPTION OF THE DRAWING5
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
-- 5
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.
,;
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~
-- 7
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
-- 8
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
-- 10
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
- 12 -
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
T4 T
~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
V2
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
- 15 -
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
-16-
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
-17-
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
-18-
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
~, ,.,.~
1~09~2S
-19-
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
-20-
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
- 21 -
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
- 24 -
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
- 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~
- 26 -
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
- 27 -
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
- 28 -
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
- 29 -
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
- 30 -
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
- 31 -
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
- 32 -
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
- 34 -
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
- 36 -
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
- 36A
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.
