U.S. patent application number 12/485683 was filed with the patent office on 2009-10-08 for split four stroke engine.
Invention is credited to Carmelo J. Scuderi.
Application Number | 20090250046 12/485683 |
Document ID | / |
Family ID | 25427513 |
Filed Date | 2009-10-08 |
United States Patent
Application |
20090250046 |
Kind Code |
A1 |
Scuderi; Carmelo J. |
October 8, 2009 |
SPLIT FOUR STROKE ENGINE
Abstract
An engine includes a crankshaft, rotating about a crankshaft
axis of the engine. A power piston is received within a first
cylinder and operatively connected to the crankshaft such that the
power piston reciprocates through a power stroke and an exhaust
stroke during a single rotation of the crankshaft. A compression
piston is received within a second cylinder and operatively
connected to the crankshaft such that the compression piston
reciprocates through an intake stroke and a compression stroke
during a single rotation of the crankshaft. A gas passage
interconnects the first and second cylinders. The gas passage
includes an inlet valve and an outlet valve defining a pressure
chamber therebetween. The outlet valve permits substantially
one-way flow of compressed gas from the pressure chamber to the
first cylinder. The power piston descends to a firing position from
its top dead center position.
Inventors: |
Scuderi; Carmelo J.;
(Springfield, MA) |
Correspondence
Address: |
Scuderi Group LLC
1111 Elm Street, Suite 33
West Springfield
MA
01089
US
|
Family ID: |
25427513 |
Appl. No.: |
12/485683 |
Filed: |
June 16, 2009 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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11385506 |
Mar 21, 2006 |
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12485683 |
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11070533 |
Mar 2, 2005 |
7017536 |
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11385506 |
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10615550 |
Jul 8, 2003 |
6880502 |
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11070533 |
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10139981 |
May 7, 2002 |
6609371 |
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10615550 |
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09909594 |
Jul 20, 2001 |
6543225 |
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10139981 |
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Current U.S.
Class: |
123/70R |
Current CPC
Class: |
F02B 41/04 20130101;
Y02T 10/14 20130101; F02F 7/0019 20130101; F02B 75/228 20130101;
F02B 75/02 20130101; F02B 2075/027 20130101; Y02T 10/12 20130101;
F02B 75/32 20130101; F02B 33/22 20130101; F02G 1/0535 20130101 |
Class at
Publication: |
123/70.R |
International
Class: |
F02B 33/22 20060101
F02B033/22 |
Claims
1. An engine comprising: a crankshaft, rotating about a crankshaft
axis of the engine; a power piston slidably received within a first
cylinder and operatively connected to the crankshaft such that the
power piston reciprocates through a power stroke and an exhaust
stroke during a single rotation of the crankshaft; a compression
piston slidably received within a second cylinder and operatively
connected to the crankshaft such that the compression piston
reciprocates through an intake stroke and a compression stroke
during a single rotation of the crankshaft; and a gas passage
interconnecting the first and second cylinders, the gas passage
including an inlet valve and an outlet valve defining a pressure
chamber therebetween, the outlet valve permitting substantially one
way flow of compressed gas from the pressure chamber to the first
cylinder; wherein the power piston descends to a firing position
from its top dead center position.
2. The engine of claim 1 wherein the inlet valve and the outlet
valve of the gas passage substantially maintain at least a
predetermined firing condition gas pressure in the pressure chamber
during the entire intake, compression, power and exhaust
strokes.
3. The engine of claim 1 wherein the crankshaft rotates at least 10
degrees past the power piston's top dead center position before the
power piston descends to its firing position.
4. The engine of claim 1 wherein the power piston leads the
compression piston by a phase shift angle that is substantially
greater than zero.
5. The engine of claim 4 wherein the phase shift angle is equal to
or greater than 20 degrees.
6. The engine of claim 1 comprising a first piston-cylinder axis
along which the power piston reciprocates within the first
cylinder, wherein the first piston cylinder axis has an offset from
the crankshaft axis such that the first piston-cylinder axis does
not intersect the crankshaft axis.
7. The engine of claim 1 comprising a second piston-cylinder axis
along which the compression piston reciprocates within the second
cylinder, wherein the second piston cylinder axis has an offset
from the crankshaft axis such that the second piston-cylinder axis
does not intersect the crankshaft axis.
8. The engine of claim 1 wherein the intake, power, compression and
exhaust strokes comprise a four stroke cycle.
9. The engine of claim 1 comprising: a first mechanical linkage
system operatively connecting the power piston to the crankshaft;
and a second mechanical linkage system operatively connecting the
compression piston to the crankshaft; wherein the first and second
linkage systems share no common mechanical link.
10. The engine of claim 1 wherein the inlet valve permits
substantially one way flow of compressed gas from the second
cylinder to the pressure chamber.
11. An engine comprising: a crankshaft, rotating about a crankshaft
axis of the engine; a power piston slidably received within a first
cylinder and operatively connected to the crankshaft such that the
power piston reciprocates through a power stroke and an exhaust
stroke during a single rotation of the crankshaft; a compression
piston slidably received within a second cylinder and operatively
connected to the crankshaft such that the compression piston
reciprocates through an intake stroke and a compression stroke
during a single rotation of the crankshaft; and a gas passage
interconnecting the first and second cylinders, the gas passage
including an inlet valve and an outlet valve defining a pressure
chamber therebetween, the outlet valve permitting substantially one
way flow of compressed gas from the pressure chamber to the first
cylinder; wherein the crankshaft rotates at least 10 degrees past
the power piston's top dead center position before the power piston
descends to its firing position.
12. The engine of claim 11 wherein the inlet valve and the outlet
valve of the gas passage substantially maintain at least a
predetermined firing condition gas pressure in the pressure chamber
during the entire intake, compression, power and exhaust
strokes.
13. The engine of claim 11 comprising a first piston-cylinder axis
along which the power piston reciprocates within the first
cylinder, wherein the first piston cylinder axis has an offset from
the crankshaft axis such that the first piston-cylinder axis does
not intersect the crankshaft axis.
14. The engine of claim 11 comprising a second piston-cylinder axis
along which the compression piston reciprocates within the second
cylinder, wherein the second piston cylinder axis has an offset
from the crankshaft axis such that the second piston-cylinder axis
does not intersect the crankshaft axis.
15. The engine of claim 11 comprising: a first mechanical linkage
system operatively connecting the power piston to the crankshaft;
and a second mechanical linkage system operatively connecting the
compression piston to the crankshaft; wherein the first and second
linkage systems share no common mechanical link and the power
piston leads the compression piston by a phase shift angle that is
substantially greater than zero.
16. The engine of claim 11 wherein the intake, power, compression
and exhaust strokes comprise a four stroke cycle.
17. The engine of claim 11 wherein the inlet valve permits
substantially one way flow of compressed gas from the second
cylinder to the pressure chamber.
Description
CROSS REFERENCE TO RELATED APPLICATIONS
[0001] This patent application is a continuation application of
U.S. application Ser. No. 11/385,506, filed Mar. 21, 2006, entitled
"SPLIT FOUR STROKE ENGINE", which is a continuation application of
U.S. application Ser. No. 11/070,533 (now U.S. Pat. No. 7,017,536),
filed Mar. 2, 2005, entitled "SPLIT FOUR STROKE ENGINE", which is a
continuation application of U.S. application Ser. No. 10/615,550
(now U.S. Pat. No. 6,880,502), filed Jul. 8, 2003, entitled "SPLIT
FOUR STROKE ENGINE", which is a continuation application of U.S.
application Ser. No. 10/139,981 (now U.S. Pat. No. 6,609,371),
filed May 7, 2002, entitled "SPLIT FOUR STROKE ENGINE", which is a
continuation application of U.S. application Ser. No. 09/909,594
(now U.S. Pat. No. 6,543,225), filed Jul. 20, 2001, entitled "SPLIT
FOUR STROKE CYCLE INTERNAL COMBUSTION ENGINE", all of which are
herein incorporated by reference in their entirety.
FIELD OF THE INVENTION
[0002] The present invention relates to internal combustion
engines. More specifically, the present invention relates to a
four-stroke cycle internal combustion engine having a pair of
offset pistons in which one piston of the pair is used for the
intake and compression strokes and another piston of the pair is
used for the power and exhaust strokes, with each four stroke cycle
being completed in one revolution of the crankshaft.
BACKGROUND OF THE INVENTION
[0003] Internal combustion engines are any of a group of devices in
which the reactants of combustion, e.g., oxidizer and fuel, and the
products of combustion serve as the working fluids of the engine.
The basic components of an internal combustion engine are well
known in the art and include the engine block, cylinder head,
cylinders, pistons, valves, crankshaft and camshaft. The cylinder
heads, cylinders and tops of the pistons typically form combustion
chambers into which fuel and oxidizer (e.g., air) is introduced and
combustion takes place. Such an engine gains its energy from the
heat released during the combustion of the non-reacted working
fluids, e.g., the oxidizer-fuel mixture. This process occurs within
the engine and is part of the thermodynamic cycle of the device. In
all internal combustion engines, useful work is generated from the
hot, gaseous products of combustion acting directly on moving
surfaces of the engine, such as the top or crown of a piston.
Generally, reciprocating motion of the pistons is transferred to
rotary motion of a crankshaft via connecting rods.
[0004] Internal combustion (IC) engines can be categorized into
spark ignition (SI) and compression ignition (CI) categories. SI
engines, i.e. typical gasoline engines, use a spark to ignite the
air-fuel mixture, while the heat of compression ignites the air
fuel mixture in CI engines, i.e., typically diesel engines.
[0005] The most common internal-combustion engine is the
four-stroke cycle engine, a conception whose basic design has not
changed for more than 100 years old. This is because of its
outstanding performance as a prime mover in the ground
transportation industry. In a four-stroke cycle engine, power is
recovered from the combustion process in four separate piston
movements (strokes) of a single piston. For purposes herein, a
stroke is defined as a complete movement of a piston from a top
dead center position to a bottom dead center position or vice
versa. Accordingly, a four-stroke cycle engine is defined herein to
be an engine which requires four complete strokes of one or more
pistons for every power stroke, i.e. for every stroke that delivers
power to a crankshaft.
[0006] Referring to FIGS. 1-4, an exemplary embodiment of a prior
art four stroke cycle internal combustion engine is shown at 10.
For purposes of comparison, the following four FIGS. 1-4 describe
what will be termed a prior art "standard engine" 10. As will be
explained in greater detail hereinafter, this standard engine 10 is
an SI engine with a 4 inch diameter piston, a 4 inch stroke and an
8 to 1 compression ratio. The compression ratio is defined herein
as the maximum volume of a predetermined mass of an air-fuel
mixture before a compression stroke, divided by the volume of the
mass of the air-fuel mixture at the point of ignition. For the
standard engine, the compression ratio is substantially the ratio
of the volume in cylinder 14 when piston 16 is at bottom dead
center to the volume in the cylinder 14 when the piston 16 is at
top dead center.
[0007] The engine 10 includes an engine block 12 having the
cylinder 14 extending therethrough. The cylinder 14 is sized to
receive the reciprocating piston 16 therein. Attached to the top of
the cylinder 14 is the cylinder head 18, which includes an inlet
valve 20 and an outlet valve 22. The cylinder head 18 cylinder 14
and top (or crown 24) of the piston 16 form a combustion chamber
26. On the inlet stroke (FIG. 1), a fuel air mixture is introduced
into the combustion chamber 26 through an intake passage 28 and the
inlet valve 20, wherein the mixture is ignited via spark plug 30.
The products of combustion are later exhausted through outlet valve
22 and outlet passage 32 on the exhaust stroke (FIG. 4). A
connecting rod 34 is pivotally attached at its top distal end 36 to
the piston 16. A crankshaft 38 includes a mechanical offset portion
called the crankshaft throw 40, which is pivotally attached to the
bottom distal end 42 of connecting rod 34. The mechanical linkage
of the connecting rod 34 to the piston 16 and crankshaft throw 40
serves to convert the reciprocating motion (as indicated by arrow
44) of the piston 16 to the rotary motion (as indicated by arrow
46) of the crankshaft 38. The crankshaft 38 is mechanically linked
(not shown) to an inlet camshaft 48 and an outlet camshaft 50,
which precisely control the opening and closing of the inlet valve
20 and outlet valve 22 respectively.
[0008] The cylinder 14 has a centerline (piston-cylinder axis) 52,
which is also the centerline of reciprocation of the piston 16. The
crankshaft 38 has a center of rotation (crankshaft axis) 54. For
purposes of this specification, the direction of rotation 46 of the
crankshaft 38 will be in the clockwise direction as viewed by the
reader into the plane of the paper. The centerline 52 of the
cylinder 14 passes directly through the center of rotation 54 of
the crankshaft 38.
[0009] Referring to FIG. 1, with the inlet valve 20 open, the
piston 16 first descends (as indicated by the direction of arrow
44) on the intake stroke. A predetermined mass of an explosive
mixture of fuel (gasoline vapor) and air is drawn into the
combustion chamber 26 by the partial vacuum thus created. The
piston continues to descend until it reaches its bottom dead center
(BDC), the point at which the piston is farthest from the cylinder
head 18.
[0010] Referring to FIG. 2, with both the inlet 20 and outlet 22
valves closed, the mixture is compressed as the piston 16 ascends
(as indicated by the direction of arrow 44) on the compression
stroke. As the end of the stroke approaches top dead center (TDC),
i.e., the point at which the piston 16 is closest to the cylinder
head 18, the volume of the mixture is compressed to one eighth of
its initial volume (due to an 8 to 1 compression ratio). The
mixture is then ignited by an electric spark from spark plug
30.
[0011] Referring to FIG. 3, the power stroke follows with both
valves 20 and 22 still closed. The piston 16 is driven downward (as
indicated by arrow 44) toward bottom dead center (BDC), due to the
expansion of the burned gas pressing on the crown 24 of the piston
16. Since the spark plug 30 is fired when the piston 16 is at or
near TDC, i.e. at its firing position, the combustion pressure
(indicated by arrow 56) exerted by the ignited gas on the piston 16
is at its maximum at this point. This pressure 56 is transmitted
through the connecting rod 34 and results in a tangential force or
torque (as indicated by arrow 58) on the crankshaft 38.
[0012] When the piston 16 is at ifs firing position, there is a
significant clearance distance 60 between the top of the cylinder
14 and the crown 24 of the piston 16. Typically, the clearance
distance is between 0.5 to 0.6 inches. For the standard engine 10
illustrated the clearance distance is substantially 0.571 inches.
When the piston 16 is at its firing position conditions are optimal
for ignition, i.e., optimal firing conditions. For purposes of
comparison, the firing conditions of this engine 10 exemplary
embodiment are: 1) a 4 inch diameter piston, 2) a clearance volume
of 7.181 cubic inches, 3) a pressure before ignition of
approximately 270 pounds per square inch absolute (psia), 4) a
maximum combustion pressure after ignition of approximately 1200
psia and 5) operating at 1400 RPM.
[0013] This clearance distance 60 corresponds typically to the 8 to
1 compression ratio. Typically, SI engines operate optimally with a
fixed compression ratio within a range of about 6.0 to 8.5, while
the compression ratios of CI engines typically range from about 10
to 16: The piston's 16 firing position is generally at or near TDC,
and represents the optimum volume and pressure for the fuel-air
mixture to ignite. If the clearance distance 60 were made smaller,
the pressure would increase rapidly.
[0014] Referring to FIG. 4, during the exhaust stroke the ascending
piston 16 forces the spent products of combustion through the open
outlet (or exhaust) valve 22. The cycle then repeats itself. For
this prior art four stoke cycle engine 10, four stokes of each
piston 16, i.e. inlet, compression, power and exhaust, and two
revolutions of the crankshaft 38 are required to complete a cycle,
i.e. to provide one power stroke. Problematically, the overall
thermodynamic efficiency of the standard four stroke engine 10 is
only about one third (1/3). That is 1/3 of the work is delivered to
the crankshaft, 1/3 is lost in waste heat, and 1/3 is lost out of
the exhaust.
[0015] As illustrated in FIGS. 3 and 5, one of the primary reasons
for this low 20 efficiency is the fact that peak torque and peak
combustion pressure are inherently locked out of phase. FIG. 3
shows the position of the piston 16 at the beginning of a power
stroke, when the piston 16 is in its firing position at or near
TDC. When the spark plug 30 fires, the ignited fuel exerts maximum
combustion pressure 56 on the piston 16, which is transmitted
through the connecting rod 34 to the crankshaft throw 40 of
crankshaft 38. However, in this position, the connecting rod 34 and
the crankshaft throw 40 are both nearly aligned with the centerline
52 of the cylinder 14. Therefore, the torque 58 is almost
perpendicular to the direction of force 56, and is at its minimum
value. The crankshaft 38 must rely on momentum generated from an
attached flywheel (not shown) to rotate it past this position.
[0016] Referring to FIG. 5, as the ignited gas expands in the
combustion chamber 26, the piston 16 descends and the combustion
pressure 56 decreases. However, as the crankshaft throw 40 rotates
past the centerline 52 and TDC, the resulting tangential force or
torque 58 begins to grow. The torque 58 reaches a maximum value
when the crankshaft throw 40 rotates approximately 30 degrees past
the centerline 52. Rotation beyond that point causes the pressure
56 to fall off so much that the torque 58 begins to decrease again,
until both pressure 56 and torque 58 reach a minimum at BDC.
Therefore, the point of maximum torque 58 and the point of maximum
combustion pressure 56 are inherently locked out of phase by
approximately 30 degrees.
[0017] Referring to FIG. 6, this concept can be further
illustrated. Here, a graph of tangential force or torque versus
degrees of rotation from TDC to BDC is shown at 62 for the standard
prior art engine 10. Additionally, a graph of combustion pressure
versus degrees of rotation from TDC to BDC is shown at 64 for
engine 10. The calculations for the graphs 62 and 64 were based on
the standard prior art engine 10 having a four inch stroke, a four
inch diameter piston, and a maximum combustion pressure at ignition
of about 1200 PSIA. As can be seen from the graphs, the point of
maximum combustion pressure 66 occurs at approximately 0 degrees
from TDC and the point of maximum torque 68 occurs approximately 30
degrees later when the pressure 64 has been reduced considerably.
Both graphs 62 and 64 approach their minimum values at BDC, or
substantially 180 degrees of rotation past TDC.
[0018] An alternative way of increasing the thermal dynamic
efficiency of a four stoke cycle engine is to increase the
compression ratio of the engine. However, automotive manufactures
have found that SI engines typically operate optimally with a
compression ratio within a range of about 6.0 to 8.5, while CI
engines typically operate best within a compression ratio range of
about 10 to 16. This is because as the compression ratios of SI or
CI engines increase substantially beyond the above ranges, several
other problems occur which outweigh the advantages gained. For
example, the engine must be made heavier and bulkier in order to
handle the greater pressures involved. Also problems of premature
ignition begin to occur, especially in SI engines.
[0019] Many rather exotic early engine designs were patented.
However, none were able to offer greater efficiencies or other
significant advantages, which would replace the standard engine 10
exemplified above. Some of these early patents included: U.S. Pat.
Nos. 848,029; 939,376; 1,111,841; 1,248,250; 1,301,141; 1,392,359;
1,856,048; 1,969,815; 2,091,410; 2,091,411; 2,091,412; 2,091,413;
2,269,948; 3,895,614; British Patent No. 299,602; British Patent
No. 721,025 and Italian Patent No. 505,576.
[0020] In particular the U.S. Pat. No. 1,111,841 to Koenig
disclosed a prior art split piston/cylinder design in which an
intake and compression stroke was accomplished in a compression
piston 12/cylinder 11 combination, and a power and an exhaust
stroke was accomplished in an engine piston 7/cylinder 8
combination. Each piston 7 and 12 reciprocates along a piston
cylinder axis which intersected the single crankshaft 5 (see FIG. 3
therein). A thermal chamber 24 connects the heads of the
compression and engine cylinders, with one end being open to the
engine cylinder and the other end having a valued discharge port 19
communicating with the compressor cylinder. A water cooled heat
exchanger 15 is disposed at the top of the compressor cylinder 11
to cool the air or air/fuel mixture as it is compressed. A set of
spaced thermal plates 25 are disposed within the thermal chamber 24
to re-heat the previously cooled compressed gas as it passes
through.
[0021] It was thought that the engine would gain efficiency by
making it easier to compress the gas by cooling it. Thereafter, the
gas was re-heated in the thermal chamber in order to increase its
pressure to a point where efficient ignition could take place. Upon
the exhaust stroke, hot exhaust gases were passed back through the
thermal chamber and out of an exhaust port 26 in an effort to
re-heat the thermal chamber.
[0022] Unfortunately, transfer of gas in all prior art engines of a
split piston design always requires work, which reduces efficiency.
Additionally, the added expansion from the thermal chamber to the
engine cylinder of Koenig also reduced compression ratio. The
standard engine 10 requires no such transfer process and associated
additional work. Moreover, the cooling and re-heating of the gas,
back and forth through the thermal chamber did not provide enough
of an advantage to overcome the losses incurred during the gas
transfer process. Therefore, the Koenig patent lost efficiency and
compression ratio relative to the standard engine 10.
[0023] For purposes herein, a crankshaft axis is defined as being
offset from the piston cylinder axis when the crankshaft axis and
the piston-cylinder axis do not intersect. The distance between the
extended crankshaft axis and the extended piston-cylinder axis
taken along a line drawn perpendicular to the piston cylinder axis
is defined as the offset. Typically, offset pistons are connected
to the crankshaft by well-known connecting rods and crankshaft
throws. However, one skilled in the art would recognize that offset
pistons may be operatively connected to a crankshaft by several
other mechanical linkages. For example, a first piston may be
connected to a first crankshaft and a second piston may be
connected to a second crankshaft, and the two crankshafts may be
operatively connected together through a system of gears.
Alternatively, pivoted lever arms or other mechanical linkages may
be used in conjunction with, or in lieu of, the connecting rods and
crankshaft throws to operatively connect the offset pistons to the
crankshaft.
[0024] Certain technology relating to reciprocating piston internal
combustion engines in which the crankshaft axis is offset from,
i.e., does not intersect with, the piston-cylinder axes is
described in U.S. Pat. Nos. 810,347; 2,957,455; 2,974,541;
4,628,876; 4,945,866; and 5,146,884; in Japan patent document
60-256,642; in Soviet Union patent document 1551-880-A; and in
Japanese Society of Automotive Engineers (JSAE) Convention
Proceedings, date 1996, issue 966, pages 129-132. According to
descriptions contained in those publications, the various engine
geometries are motivated by various considerations, including power
and torque improvements and friction and vibration reductions.
Additionally, in-line, or straight engines in which the crankshaft
axis is offset from the piston axes were used in early twentieth
century racing engines.
[0025] However, all of the improvements gained were due to
increasing the torque angles on the power stroke only.
Unfortunately, as will be discussed in greater detail hereinafter,
the greater the advantage an offset was to the power stroke was
also accompanied by an associated increasing disadvantage to the
compression stroke. Therefore, the degree of offset quickly becomes
self limiting, wherein the advantages to torque, power, friction
and vibration to the power stroke do not out weigh the
disadvantages to the same functions on the compression stroke.
Additionally, no advantages were taught or discussed regarding
offsets to optimize the compression stroke.
[0026] By way of example, a recent prior art attempt to increase
efficiency in a standard engine 10 type design through the use of
an offset is disclosed in U.S. Pat. No. 6,058,901 to Lee. Lee
believes that improved efficiency will result by reducing the
frictional forces of the piston rings on the side walls over the
full duration of two revolutions of a four stroke cycle (see Lee,
column 4, lines 1016). Lee attempts to accomplish this by providing
an offset cylinder, wherein the timing of combustion within each
cylinder is controlled to cause maximum combustion pressure to
occur when an imaginary plane that contains both a respective
connection axis of a respective connecting rod to the respective
piston and a respective connection axis of the connecting rod to a
respective throw of the crankshaft is substantially coincident with
the respective cylinder axis along which the piston
reciprocates.
[0027] However, though the offset is an advantage during the power
stroke, it becomes a disadvantage during the compression stroke.
That is, when the piston travels from bottom dead center to top
dead center during the compression stroke, the offset
piston-cylinder axis creates an angle between the crankshaft throw
and connecting rod that reduces the torque applied to the piston.
Additionally, the side forces resulting from the poor torque angles
on the compression stroke actually increase wear on the piston
rings. Accordingly, a greater amount of power must be consumed in
order to compress the gas to complete the compression stroke as the
offset increases. Therefore, the amount of offset is severely
limited by its own disadvantages on the compression side.
Accordingly, large prior art offsets, i.e., offsets in which the
crankshaft must rotate at least 20 degrees past a pistons top dead
center position before the piston can reach a firing position, have
not been utilized, disclosed or taught. As a result, the relatively
large offsets required to substantially align peak torque to peak
combustion pressure cannot be accomplished with Lee's
invention.
[0028] Variable Compression Ratio (VCR) engines are a class of
prior art CI engines designed to take advantage of varying the
compression ratio on an engine to increase efficiency. One such
typical example is disclosed in U.S. Pat. No. 4,955,328 to
Sobotowski. Sobotowski describes an engine in which compression
ratio is varied by altering the phase relation between two pistons
operating in cylinders interconnected through a transfer port that
lets the gas flow in both directions.
[0029] However, altering the phase relation to vary compression
ratios impose design requirements on the engine that greatly
increase its complexity and decrease its utility. For example, each
piston of the pair of pistons must reciprocate through all four
strokes of a complete four stroke cycle, and must be driven by a
pair of crankshafts which rotate through two full revolutions per
four stroke cycle. Additionally, the linkages between the pair of
crankshafts become very complex and heavy. Also the engine is
limited by design to CI engines due to the higher compression
ratios involved.
[0030] Various other relatively recent specialized prior art
engines have also been designed in an attempt to increase engine
efficiency. One such engine is described in U.S. Pat. No. 5,546,897
to Bracken entitled "Internal Combustion Engine with Stroke
Specialized Cylinders". In Brackett, the engine is divided into a
working section and a compressor section. The compressor section
delivers charged air to the working section, which utilizes a
scotch yoke or conjugate drive motion translator design to enhance
efficiency. The specialized engine can be described as a
horizontally opposed engine in which a pair of opposed pistons
reciprocate in opposing directions within one cylinder block.
[0031] However, the compressor is designed essentially as a super
charger which delivers supercharged gas to the working section.
Each piston in the working section must reciprocate through all
four strokes of intake, compression, power and exhaust, as each
crankshaft involved must complete two full revolutions per
four-stroke cycle. Additionally, the design is complex, expensive
and limited to very specialized CI engines.
[0032] Another specialized prior art design is described in U.S.
Pat. No. 5,623,894 to Clarke entitled "Dual Compression and Dual
Expansion Engine". Clarke essentially discloses a specialized
two-stroke engine where opposing pistons are disposed in a single
cylinder to perform a power stroke and a compression stroke. The
single cylinder and the crowns of the opposing pistons define a
combustion chamber, which is located in a reciprocating inner
housing. Intake and exhaust of the gas into and out of the
combustion chamber is performed by specialized conical pistons, and
the reciprocating inner housing.
[0033] However, the engine is a highly specialized two-stroke
system in which the opposing pistons each perform a compression
stroke and a power stroke in the same cylinder. Additionally, the
design is very complex requiring dual crankshafts, four pistons and
a reciprocating inner housing to complete the single revolution
two-stroke cycle. Also, the engine is limited to large CI engine
applications.
[0034] Accordingly, there is a need for an improved four-stroke
internal combustion engine, which can enhance efficiency by more
closely aligning the torque and force curves generated during a
power stroke without increasing compression ratios substantially
beyond normally accepted design limits.
SUMMARY OF THE INVENTION
[0035] The present invention offers advantages and alternatives
over the prior art by providing a four-stroke cycle internal
combustion engine having a pair of pistons in which one piston of
the pair is used for the intake and compression strokes and another
piston of the pair is used for the power and exhaust strokes, with
each four stroke cycle being completed in one revolution of the
crankshaft.
[0036] These and other advantages are accomplished in an exemplary
embodiment of the invention by providing an engine with a
crankshaft rotating about a crankshaft axis of the engine. A power
piston is slidably received within a first cylinder and operatively
connected to the crankshaft such that the power piston reciprocates
through a power stroke and an exhaust stroke of a four stroke cycle
during a single rotation of the crankshaft. A compression piston is
slidably received within a second cylinder and operatively
connected to the crankshaft such that the compression piston
reciprocates through an intake stroke and a compression stroke of
the same four stroke cycle during the same rotation of the
crankshaft. A gas passage interconnects the first and second
cylinders. The gas passage includes an inlet valve and an outlet
valve defining a pressure chamber therebetween. The outlet valve
permits substantially one-way flow of compressed gas from the
pressure chamber to the first cylinder. Combustion is initiated in
the first cylinder between 0 degrees and 40 degrees of rotation of
the crankshaft after the power piston has reached its top dead
center position.
[0037] In an alternative embodiment of the invention the inlet
valve and the outlet valve of the gas passage substantially
maintain at least a predetermined firing condition gas pressure in
the pressure chamber during the entire four stroke cycle.
[0038] In another alternative embodiment of the invention, the
power piston leads the compression piston by a phase shift angle
that is substantially equal to or greater than 20 degrees and equal
to or less than 39 degrees.
BRIEF DESCRIPTION OF THE DRAWINGS
[0039] FIG. 1 is a schematic diagram of a representative prior art
four stroke cycle engine, during the intake stoke;
[0040] FIG. 2 is a schematic diagram of the prior art engine of
FIG. 1 during the compression stoke;
[0041] FIG. 3 is a schematic diagram of the prior art engine of
FIG. 1 during the power stoke;
[0042] FIG. 4 is a schematic diagram of the prior art engine of
FIG. 1 during the exhaust stoke;
[0043] FIG. 5 is a schematic diagram of the prior art engine of
FIG. 1 when the piston is at the position of maximum torque;
[0044] FIG. 6, is a graphical representation of torque and
combustion pressure of the prior art engine of FIG. 1;
[0045] FIG. 7 is a schematic diagram of an engine in accordance
with the present invention during the exhaust and intake
strokes;
[0046] FIG. 8 is a schematic diagram of the engine of FIG. 7 when
the first piston has just reached top dead center (TDC) at the
beginning of a power stroke;
[0047] FIG. 9 is a schematic diagram of the engine of FIG. 7 when
the first piston has reached its firing position;
[0048] FIG. 10, is a graphical representation of torque and
combustion pressure of the engine of FIG. 7; and
[0049] FIG. 11 is a schematic diagram of an alternative embodiment
of an engine in accordance with the present invention having
unequal throws and piston diameters.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0050] Referring to FIG. 7, an exemplary embodiment of a four
stroke internal combustion engine in accordance with the present
invention is shown generally at 100. The engine 100 includes an
engine block 102 having a first cylinder 104 and a second cylinder
106 extending therethrough. A crankshaft 108 is journaled for
rotation about a crankshaft axis 110 (extending perpendicular to
the plane of the paper).
[0051] The engine block 102 is the main structural member of the
engine 100 and extends upward from the crankshaft 108 to the
junction with the cylinder head 112. The engine block 102 serves as
the structural framework of the engine 100 and typically carries
the mounting pad by which the engine is supported in the chassis
(not shown). The engine block 102 is generally a casting with
appropriate machined surfaces and threaded holes for attaching the
cylinder head 112 and other units of the engine 100.
[0052] The cylinders 104 and 106 are openings, typically of
generally circular cross section, that extend through the upper
portion of the engine block 102. Cylinders are defined herein as
the chambers within which pistons of an engine reciprocate, and do
not have to be generally circular in cross section, e.g., they may
have a generally elliptical or half moon shape.
[0053] The internal walls of cylinders 104 and 106 are bored and
polished to form smooth, accurate bearing surfaces sized to receive
a first power piston 114, and a second compression piston 116
respectively. The power piston 114 reciprocates along a first
piston-cylinder axis 113, and the compression piston 116
reciprocates along a second piston-cylinder axis 115. The first and
second cylinders 104 and 106 are disposed in the engine 100 such
that the first and second piston-cylinder axes 113 and 115 pass on
opposing sides of the crankshaft axis 110 without intersecting the
crankshaft axis 110.
[0054] The pistons 114 and 116 are typically cup shaped cylindrical
castings of steel or aluminum alloy. The upper closed ends, i.e.,
tops, of the power and compression pistons 114 and 116 are the
first and second crowns 118 and 120 respectively. The outer
surfaces of the pistons 114,116 are generally machined to fit the
cylinder bore closely and are typically grooved to receive piston
rings (not shown) that seal the gap between the pistons and the
cylinder walls.
[0055] First and second connecting rods 122 and 124 each include an
angle bend 121 and 123 respectively. The connecting rods 122 and
124 are pivotally attached at their top distal ends 126 and 128 to
the power and compression pistons 114 and 116 respectively. The
crankshaft 108 includes a pair of mechanically offset portions
called the first and second throws 130 and 132, which are pivotally
attached to the bottom opposing distal ends 134 and 136 of the
first and second connecting rods 122 and 124 respectively. The
mechanical linkages of the connecting rods 122 and 124 to the
pistons 114, 116 and crankshaft throws 130,132 serve to convert the
reciprocating motion of the pistons (as indicated by directional
arrow 138 for the power piston 114, and directional arrow 140 for
the compression piston 116) to the rotary motion (as indicated by
directional arrow 142) of the crankshaft 108. The first piston
cylinder axis 113 is offset such that it is disposed in the
imaginary half plane through which the first crankshaft throw 130
rotates from its top dead center position to its bottom dead center
position. The second piston cylinder axis 115 is offset in the
opposing imaginary half plane.
[0056] Though this embodiment shows the first and second pistons
114 and 116 connected directly to crankshaft 108 through connecting
rods 122 and 124 respectively, it is within the scope of this
invention that other means may also be employed to operatively
connect the pistons 114 and 116 to the crankshaft 108. For example
a second crankshaft may be used to mechanically link the pistons
114 and 116 to the first crankshaft 108.
[0057] The cylinder head 112 includes a gas passage 144
interconnecting the first and second cylinders 104 and 106. The gas
passage includes an inlet check valve 146 disposed in a distal end
of the gas passage 144 proximate the second cylinder 106. An outlet
poppet valve 150 is also disposed in an opposing distal end of the
gas passage 144 proximate the top of the first cylinder 104. The
inlet check valve 146 and outlet poppet valve 150 define a pressure
chamber 148 there between. The inlet valve 146 permits the one way
flow of compressed gas from the second cylinder 106 to the pressure
chamber 148. The outlet valve 150 permits the one way flow of
compressed gas from the pressure chamber 148 to the first cylinder
104. Though check and poppet type valves are described as the inlet
and the outlet valves 146 and 150 respectively, any valve design
appropriate for the application may be used instead, e.g., the
inlet valve 146 may also be of the poppet type.
[0058] The cylinder head 112 also includes an intake valve 152 of
the poppet type disposed over the top of the second cylinder 106,
and an exhaust valve 154 of the poppet type disposed over the top
to the first cylinder 104. Poppet valves 150, 152 and 154 typically
have a metal shaft 156 with a disk 158 at one end fitted to block
the valve opening. The other end of the shafts 156 of poppet valves
150, 152 and 154 are mechanically linked to camshafts 160, 162 and
164 respectively. The camshafts 160, 162 and 164 are typically a
round rod with generally oval shaped lobes located inside the
engine block 102 or in the cylinder head 112.
[0059] The camshafts 160, 162 and 164 are mechanically connected to
the crankshaft 108, typically through a gear wheel, belt or chain
links (not shown). When the crankshaft 108 forces the camshafts
160, 162 and 164 to turn, the lobes on the camshafts 160, 162 and
164 cause the valves 150, 152 and 154 to open and close at precise
moments in the engine's cycle.
[0060] The crown 120 of compression piston 116, the walls of second
cylinder 106 and the cylinder head 112 form a compression chamber
166 for the second cylinder 106. The crown 118 of power piston 114,
the walls of first cylinder 104 and the cylinder head 112 form a
separate combustion chamber 168 for the first cylinder 104. A spark
plug 170 is disposed in the cylinder head 112 over the first
cylinder 104 and is controlled by a control device (not shown)
which precisely times the ignition of the compressed air gas
mixture in the combustion chamber 168. Though this embodiment
describes a spark ignition (SI) engine, one skilled in the art
would recognize that compression ignition (CI) engines are within
the scope of this invention also.
[0061] During operation, the power piston 114 leads the compression
piston 116 by a phase shift angle 172, defined by the degrees of
rotation the crankshaft 108 must rotate after the power piston 114
has reached its top dead center position in order for the
compression piston 116 to reach its respective top dead center
position. Preferably this phase shift is between 30 to 60 degrees.
For this particular preferred embodiment, the phase shift is fixed
substantially at 50 degrees.
[0062] FIG. 7 illustrates the power piston 114 when it has reached
its bottom dead center (BDC) position and has just started
ascending (as indicated by arrow 138) into its exhaust stroke.
Compression piston 116 is lagging the power piston 114 by 50
degrees and is descending (arrow 140) through its intake stroke.
The inlet valve 156 is open to allow an explosive mixture of fuel
and air to be drawn into the compression chamber 166. The exhaust
valve 154 is also open allowing piston 114 to force spent products
of combustion out of the combustion chamber 168.
[0063] The check valve 146 and poppet valve 150 of the gas passage
144 are closed to prevent the transfer of ignitable fuel and spent
combustion products between the two chambers 166 and 168.
Additionally during the exhaust and intake strokes, the inlet check
valve 146 and outlet poppet valve 150 seal the pressure chamber 148
to substantially maintain the pressure of any gas trapped therein
from the previous compression and power strokes.
[0064] Referring to FIG. 8, the power piston 114 has reached its
top dead center (TDC) position and is about to descend into its
power stroke (indicated by arrow 138), while the compression piston
116 is ascending through its compression stroke (indicated by arrow
140). At this point, inlet check valve 146, outlet valve 150,
intake valve 152 and exhaust valve 154 are all closed.
[0065] At TDC piston 114 has a clearance distance 178 between the
crown 118 of the piston 114 and the top of the cylinder 104. This
clearance distance 178 is very small by comparison to the clearance
distance 60 of standard engine 10 (best seen in FIG. 3). This is
because the power stroke in engine 100 follows a low pressure
exhaust stroke, while the power stroke in standard engine 10
follows a high pressure compression stroke. Therefore, in distinct
contrast to the standard engine 10, there is little penalty to
engine 100 to reduce the clearance distance 178 since there is no
high pressure gas trapped between the crown 118 and the top of the
cylinder 114. Moreover, by reducing the clearance distance 178, a
more thoroughly flushing of nearly all exhaust products is
accomplished.
[0066] In order to substantially align the point of maximum torque
with maximum combustion pressure, the crankshaft 108 must be
rotated approximately 40 degrees past its top dead center position
when the power piston 114 is in its optimal firing position.
Additionally, similar considerations hold true on the compression
piston 116, in order to reduce the amount of torque and power
consumed by the crankshaft 108 during a compression stroke. Both of
these considerations require that the offsets on the
piston-cylinder axes be much larger than any previous prior art
offsets, i.e., offsets in which the crankshaft must rotate at least
20 degrees past a pistons top dead center position before the
piston can reach a firing position. These offsets are in fact so
large that a straight connecting rod linking the pistons 114 and
116 would interfere with the lower distal end of the cylinders 104
and 106 during a stroke.
[0067] Accordingly, the bend 121 in connecting rod 122 must be
disposed intermediate its distal ends and have a magnitude such
that the connecting rod 122 clears the bottom distal end 174 of
cylinder 104 while the power piston 114 reciprocates through an
entire stroke. Additionally, the bend 123 in connecting rod 124
must be disposed intermediate its distal ends and have a magnitude
such that the connecting rod 124 clears the bottom distal end 176
of cylinder 106 while the compression piston 116 reciprocates
through an entire stroke.
[0068] Referring to FIG. 9, the crankshaft 108 has rotated an
additional 40 degrees (as indicated by arrow 180) past the TDC
position of power piston 114 to reach its firing position, and the
compression piston 116 is just completing its compression stroke.
During this 40 degrees of rotation, the compressed gas within the
second cylinder 116 reaches a threshold pressure which forces the
check valve 146 to open, while cam 162 is timed to also open outlet
valve 150. Therefore, as the power piston 114 descends and the
compression piston 116 ascends, a substantially equal mass of
compressed gas is transferred from the compression chamber 166 of
the second cylinder 106 to the combustion chamber 168 of the first
cylinder 104. When the power piston 114 reaches its firing
position, check valve 146 and outlet valve 150 close to prevent any
further gas transfer through pressure chamber 148. Accordingly, the
mass and pressure of the gas within the pressure chamber 148 remain
relatively constant before and after the gas transfer takes place.
In other words, the gas pressure within the pressure chamber 148 is
maintained at least (at or above) a predetermined firing condition
pressure, e.g., approximately 270 psia, for the entire four stroke
cycle.
[0069] By the time the power piston 114 has descended to its firing
position from TDC, the clearance distance 178 has grown to
substantially equal that of the clearance distance 60 of standard
engine 10 (best seen in FIG. 3), i.e., 0.571. Additionally, the
firing conditions are substantially the same as the firing
conditions of the standard engine 10, which are generally: 1) a 4
inch diameter piston, 2) a clearance volume of 7.181 cubic inches,
3) a pressure before ignition of approximately 270 pounds per
square inch absolute (psia), and 4) a maximum combustion pressure
after ignition of approximately 1200 psia. Moreover, the angle of
the first throw 130 of crankshaft 108 is in its maximum torque
position, i.e., approximately 40 degrees past TDC. Therefore, spark
plug 170 is timed to fire such that maximum combustion pressure
occurs when the power piston 114 substantially reaches its position
of maximum torque.
[0070] During the next 10 degrees of rotation 142 of the crankshaft
108, the compression piston 116 will pass through to its TDC
position and thereafter start another intake stroke to begin the
cycle over again. The compression piston 116 also has a very small
clearance distance 182 relative to the standard engine 10. This is
possible because, as the gas pressure in the compression chamber
166 of the second cylinder 106 reaches the pressure in the pressure
chamber 148, the check valve 146 is forced open to allow gas to
flow through. Therefore, very little high pressure gas is trapped
at the top of the power piston 116 when it reaches its TDC
position.
[0071] The compression ratio of engine 100 can be anything within
the realm of SI or CI engines, but for this exemplary embodiment it
is substantially within the range of 6 to 8.5. As defined earlier,
the compression ratio is the maximum volume of a predetermined mass
of an air-fuel mixture before a compression stroke, divided by the
volume of the mass of the air-fuel mixture at the point of
ignition. For the engine 100, the compression ratio is
substantially the ratio of the displacement volume in second
cylinder 106 when the compression piston 116 travels from SDC to
TDC to the volume in the first cylinder 104 when the power piston
114 is at its firing position.
[0072] In distinct contrast to the standard engine 10 where the
compression stroke and the power stroke are always performed in
sequence by the same piston, the power stroke is performed by the
power piston 114 only, and the compression stroke is performed by
the compression piston 116 only. Therefore, the power piston 116
can be offset to align maximum combustion pressure with maximum
torque applied to the crankshaft 108 without incurring penalty for
being out of alignment on the compression stroke. Vice versa, the
compression piston 114 can be offset to align maximum compression
pressure with maximum torque applied from the crankshaft 108
without incurring penalty for being out of alignment on the power
stroke.
[0073] Referring to FIG. 10, this concept can be further
illustrated. Here, a graph of tangential force or torque versus
degrees of rotation from TDC for power piston 114 is shown at 184
for the engine 100. Additionally, a graph of combustion pressure
versus degrees of rotation from TDC for power piston 114 is shown
at 186 for engine 100. The calculations for the graphs 184 and 186
were based on the engine 100 having firing conditions substantially
equal to that of a standard engine. That is: 1) a 4 inch diameter
piston, 2) a clearance volume of 7.181 cubic inches, 3) a pressure
before ignition of approximately 270 pounds per square inch
absolute (psia), 4) a maximum combustion pressure after ignition of
approximately 1200 psia and 5) substantially equal revolutions per
minute (RPM) of the crankshafts 108 and 38. In distinct contrast
with the graphs of FIG. 6 for the standard prior art engine 10, the
point of maximum combustion pressure 188 is substantially aligned
with the point of maximum torque 190. This alignment of combustion
pressure 186 with torque 184 results in a significant increase in
efficiency.
[0074] Moreover, the compression piston's 116 offset can also be
optimized to substantially align the maximum torque delivered to
the compression piston 116 from the crankshaft 108 with the maximum
compression pressure of the gas. The compression pistons 116 offset
reduces the amount of power exerted in order to complete a
compression stroke and further increases the overall efficiency of
engine 100 relative to the standard engine 10. With the combined
power and compression piston 114, and 116 offsets, the overall
theoretical efficiency of engine 100 can be increased by
approximately 20 to 40 percent relative to the standard engine.
[0075] Referring to FIG. 11, an alternative embodiment of a split
four stroke engine having unequal throws and unequal piston
diameters is shown generally at 200. Because the compression and
power strokes are performed by separate pistons 114, 116, various
enhancements can be made to optimize the efficiency of each stroke
without the associated penalties incurred when the strokes are
performed by a single piston. For example, the compression piston
diameter 204 can be made larger than the power piston diameter 202
to further increase the efficiency of compression. Additionally,
the radius 206 of the first throw 130 for the power piston 114 can
be made larger than the radius 208 of the second throw 132 for the
compression piston 116 to further enhance the total torque applied
to the crankshaft 108.
[0076] While preferred embodiments have been shown and described,
various modifications and substitutions may be made thereto without
departing from the spirit and scope of the invention. Accordingly,
it is to be understood that the present invention has been
described by way of illustration and not limitation.
* * * * *