INTERNATIONAL JOURNAL OF ENERGY RESEARCH

Published online 28 February 2011 in Wiley Online Library (wileyonlinelibrary.com). DOI: 10.1002/er.1823

TECHNICAL NOTE

An internal combustion engine platform for increased
thermal efficiency, constant-volume combustion,
variable compression ratio, and cold start

Yiding Cao

,y

Department of Mechanical and Materials Engineering, Florida International University, Miami, FL 33174, U.S.A.

SUMMARY

An internal-combustion engine platform, which may operate on a portfolio of cycles for an increased expansion
ratio, combustion under constant volume, variable compression ratio, and cold start, is introduced. Through
unique thermodynamic cycles, the engine may be able to operate on a much greater expansion ratio than the
compression ratio for a significantly improved thermal efficiency. This improvement is attained without involving a
complex mechanical structure or enlarged engine size, and at the same time without reducing the compression
ratio. The engine with these features may serve as an alternative to the Atkinson cycle engine or the Miller cycle
engine. Furthermore, based on the same engine platform, the engine may operate on other cycles according to the
load conditions and environmental considerations. These cycles include those for combustion under constant
volume, variable compression ratio under part load conditions, and cold start for alternative fuels.

It is believed that the introduced thermodynamic cycles associated with the engine platform may enable a future

internal combustion engine that could generally increase the thermal efficiency by about 20% under full and part
load conditions and overcome the cold start problem associated with diesel fuels or alternative fuels such as ethanol
and methanol. Copyright r 2011 John Wiley & Sons, Ltd.

KEY WORDS

increased expansion ratio; constant-volume combustion; variable compression ratio; cold start

Correspondence

*Yiding Cao, Department of Mechanical and Materials Engineering, Florida International University, Miami, FL 33174, U.S.A.

y

E-mail: [e-mail]

Received 23 July 2010; Revised 15 December 2010; Accepted 27 December 2010

1. INTRODUCTION

It is well known that an internal combustion engine
having an expansion ratio greater than the compres-
sion ratio will have a higher thermal efficiency. An
engine with such a feature was first invented by James
Atkinson and was termed as the Atkinson cycle engine.
The concept of the Atkinson cycle may be employed
for both two-stroke and four-stroke engines. In an
Atkinson cycle engine, a complex linkage mechanism is
used to allow different stroke lengths for intake/
compression and power/exhaust strokes, so that the
expansion ratio may be greater than the compression
ratio to increase engine’s efficiency. However, this
complex linkage mechanism may substantially increase
the engine size and result in mechanical weakness of
the engine [1,2]. As a result, although the Atkinson
cycle engine may have been a dream engine for many

engine designers over the past 100 years, engine design
based on the original Atkinson cycle has not reached
the mass production level.

An alternative to the Atkinson cycle is the Miller

cycle, through early or late closing of the intake valve
to decrease the compression ratio. The Miller cycle
differs from the Atkinson cycle in that the engine’s
structure remains the same as that of an engine oper-
ating on a conventional four-stroke cycle. When an
engine is operating under part load conditions, the
Miller cycle may have the benefit of reducing the
pumping losses by eliminating the charge throttling.
One common practice of evaluating the performance
of a Miller cycle is to treat it as an Atkinson cycle.
In this evaluation, the Miller cycle’s reduced com-
pression ratio due to the early or late closing of the
intake valve was used against the engine’s expansion
ratio, and it was concluded that a Miller cycle would

Copyright r 2011 John Wiley & Sons, Ltd.

Int. J. Energy Res. 2012; 36:682–690

682

have a thermal efficiency higher than that of a con-
ventional cycle, such as an Otto cycle or a diesel cycle.
This evaluation may be confusing, however, because
the calculation is based on the decreased compression
ratio for both Miller and conventional cycles. It is well
known that a greater compression ratio will produce a
higher thermal efficiency. As a result, a conventional
cycle without throttling, with its compression ratio
being equal to the engine’s full compression ratio, may
in fact have a higher thermal efficiency than that of the
Miller cycle.

In this paper, an internal combustion engine plat-

form having a substantially increased expansion ratio
and subsequently a significantly improved thermal
efficiency is described. This improvement is attained
without involving a complex mechanical structure or
an enlarged engine size, and at the same time without
reducing the compression ratio. The engine with these
features may serve as an alternative to the Atkinson
cycle engine or the Miller cycle engine. Additionally,
operation of the engine is not limited to the cycle for
increasing the expansion ratio. Based on the same
engine platform, the engine may operate on a portfolio
of cycles according to the load conditions and environ-
mental considerations. The other significant cycles in
the portfolio include those for combustion under con-
stant volume, variable compression ratio under part
load conditions, and cold start mechanism for alter-
native fuels.

2. CYCLES FOR INCREASED
EXPANSION RATIO

Figure 1 illustrates schematically the configuration of an
internal combustion engine in accordance with the
objective of increasing the expansion ratio. The engine
block contains at least a cylinder and a piston that is
disposed within the cylinder. Associated with each
cylinder, the cylinder head contains two combustion

chambers that are designated as chamber 1 and
chamber 2, respectively. When the piston reaches the
top dead center (TC), the cylinder space, as enclosed by
the bottom face of the cylinder head, the top face of the
piston, and the sidewall of the cylinder, is minimized.
As in a conventional engine, combustion chamber 1
may be provided with an opening to an intake port and
an opening to an exhaust port. The intake port is
provided with an intake valve, and the exhaust port is
provided with an exhaust valve, as shown in Figure 2(a),
which is a top-sectional view of the internal combustion
engine, illustrating the arrangements of the combustion
chambers and valves. It should be mentioned that the
arrangement as shown in Figure 2(a) is one of the many
design options depending upon conditions such as the
size of the engine as well as thermal and structural
considerations. Figure 2(b) illustrates another option in
connection with the arrangement of the combustion
chambers as well as intake and exhaust valves.

In this option, the exhaust valves are installed out-

side of the combustion chambers and shared by two
combustion chambers. It should be pointed out that
Figure 2(a, b) shows only two possible options of valve
arrangements. In an alternative arrangement, both
intake and exhaust valves may be installed outside of the
two combustion chambers and shared by the two com-
bustion chambers. In this case, the combustion chamber
described here may be similar in shape or performance
characteristics to the well-known prechamber in a

Figure 1. Schematic of an IC engine with two combustion

chambers per cylinder, showing the positions of valves and

piston during an intake stroke for both combustion chambers.

Figure 2. (a) Schematic top-sectional view of the engine, illustrat-

ing the arrangements of the combustion chambers and valves and

(b) Schematic top-sectional view of the engine, illustrating another

arrangement of the combustion chambers and valves with exhaust

valves being shared by the two combustion chambers.

An internal combustion engine platform

Y. Cao

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683

compression–ignition engine or a divided chamber in a
homogeneous charge combustion engine. Also, com-
bustion chamber 1 is provided with an opening port to
the cylinder. The opening port may be opened or closed
by a chamber valve (Figure 1), which is designated as
chamber valve 1. The chamber valve as illustrated here is
schematic in nature; it may be a puppet valve, slide
valve, rotary valve, butterfly valve, switch valve, gate
valve, or ball valve. For a spark-ignition or homogenous
charge combustion engine, combustion chamber 1 is
also provided with an ignition device. Or for a com-
pression–ignition combustion engine, chamber 1 is
provided with a fuel injector device. Similarly, combus-
tion chamber 2 is provided with an intake valve, an
exhaust valve, chamber valve 2, and an ignition device
or a fuel injector, as shown in Figures 1 and 2(a, b).
Unlike a conventional internal combustion engine that
normally works on a four-stroke cycle or a two-stroke
cycle, the engine described here may be adapted to work
on a six-stroke cycle. The operation of a spark-ignition
engine or a compression–ignition engine is described in
detail with reference to Figures 1–7.

Figure 1 illustrates representative conditions for the

first stroke of the cycle, the intake stroke for both
combustion chambers. In this case, the intake valves
and chamber valves associated with both combustion
chambers are all opened, while both exhaust valves are
closed. The piston moves downwardly, admitting an
amount of charge into the cylinder.

Figure 3 illustrates representative conditions for the

second stroke, the compression stroke for both com-
bustion chambers. In this case, intake valves as well
exhaust valves for both combustion chambers are all
closed while both combustion chamber valves remain
open. The piston moves upwardly, compressing the
charge into both combustion chambers to a higher
pressure and a higher temperature.

Figure 4 illustrates representative conditions for

the third stroke, the power stroke for combustion

chamber 1. In this case, all the intake and exhaust valves
as well as the chamber valve associated with combustion
chamber 2 are closed. The charge previously entering
combustion chamber 2 in the compression stroke is
being held within the same chamber. For combustion
chamber 1, chamber valve 1 remains open. For a
homogenous charge engine the combustible mixture in
combustion chamber 1 was ignited in a predetermined
timing by a spark plug and explosive combustion
occurred. Or for a compression–ignition engine, an
amount of fuel was injected into the compressed air
in chamber 1 through a fuel injector in a suitable
timing, and combustion occurred due to auto ignition.
The high pressure, high temperature gases expand into
the cylinder space, delivering work to the piston.

Figure 5 illustrates representative conditions for

the fourth stroke, the exhaust stroke for combustion
chamber 1. All the valves associated with combustion
chambers 2 remain closed, and the charge previously
entering chamber 2 in the compression stroke is still
being enclosed within the same chamber. For com-
bustion chamber 1, chamber valve 1 remains open.
The exhaust valve associated with chamber 1 is opened
for the arrangement as shown in Figure 2(a), or both
exhaust valves are opened for the arrangement as
shown in Figure 2(b). Exhaust gases are being dis-
charged out of the cylinder.

Figure 6 illustrates representative conditions for the

fifth stroke, the power stroke for combustion chamber 2.
In this case, chamber valve 2 is opened while all other
valves are closed, with the exception that exhaust valve
or intake valve associated with chamber 1 in connection
with the arrangement as shown in Figure 2(a) may not
necessarily be subject to this requirement. For a homo-
geneous

charge

engine,

the

combustible

mixture

previously being held in chamber 2 was ignited in a
predetermined timing by a spark plug and combustion
occurred. Or for a compression–ignition engine, an
amount of fuel was injected into the compressed air

Figure 3. Schematic sectional view of the engine, illustrating

the positions of the valves and piston during a compression

stroke for both combustion chambers.

Figure 4. Schematic sectional view of the engine, illustrating

the positions of the valves and piston during the power stroke

for the first combustion chamber.

An internal combustion engine platform

Y. Cao

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684

previously being held in chamber 2 through a fuel
injector in a suitable timing, and combustion occurred
due to auto ignition. The high pressure, high temperature

gases expand into the cylinder, delivering work to the
piston.

Figure 7 illustrates representative conditions for the

sixth stroke, the exhaust stroke for combustion
chamber 2. For the arrangement as shown in Figure
2(a), the exhaust valve associated with chamber 2 is
opened, and exhaust gases are being discharged from
the cylinder into the exhaust port through chamber 2.
The chamber valve and exhaust valve associated with
chamber 1 may optionally be opened to aid the dis-
charge of the exhaust gases from the cylinder, as
shown in Figure 7. For the arrangement as shown in
Figure 2(b), however, both exhaust valves are opened
to discharge the exhaust gas out of the cylinder space
while both first and second combustion chambers may
be closed when the exhaust gas pressure in the cylinder
space has dropped to a value close to the ambient
pressure to reduce the exhaust gas heating on the
combustion chamber walls.

The operational benefit of the above-described cycle

is significant; the expansion ratios in the two power
strokes are almost doubled while the compression ratio
remains the same. Assuming that the two combustion
chambers have the same interior volume and contain
the same amount of charge at the end of compression
stroke, the gas leakage through the enclosed combus-
tion chamber is negligible, and the charge left in the
cylinder space is very small when the piston reaches the
TC, the expansion ratio of the first power stroke, r

e

, is

r

e

¼ ðV

d

1V

tdc;1

Þ=V

tdc;1

¼ ðV

d

1

0:5V

c

Þ=ð0:5V

c

Þ

¼ ðV

d

=V

c

1

0:5Þ=0:5 ¼ ððV

d

1V

c

V

c

Þ=V

c

1

0:5Þ=0:5

¼ ðr

c

110:5Þ=0:5

¼ ðr

c

0:5Þ=0:5

ð1Þ

where V

tdc,1

is the gas volume in combustion chamber 1

when the piston is at its TC, V

d

is the piston displaced

volume or swept volume, V

c

is the clearance volume

associated with both combustion chambers when the
piston is at its TC, and r

c

is the engine’s compression

ratio, which is given by (V

d

1V

c

)

/V

c

. Consider an engine

that has a compression ratio of 9, the expansion ration
from the above equation would be

r

e

¼ ð9 0:5Þ=0:5 ¼ 17

ð2Þ

The result indicates that the engine’s expansion ratio

is almost doubled. This is also true for the second
power stroke related to combustion chamber 2.

It should be pointed out that the common practices

of variable valve timing and lift, early/late opening
or early/late closing, and valve overlap periods will
still be applicable to the operation of the combustion–
chamber valves. It is understandable that the timings
of opening or closing combustion chambers, as well as
the timings of combustion–ignition/fuel injection will
be important to the operation of an engine. Although
the timings of the combustion-chamber valves and the
combustion–ignition/fuel injection may be set based on

Figure 5. Schematic sectional view of the engine, illustrating

the positions of the valves and piston during the exhaust stroke

for the first combustion chamber.

Figure 6. Schematic sectional view of the engine, illustrating

the positions of the valves and piston during the power stroke

for the second combustion chamber.

Figure 7. Schematic sectional view of the engine, illustrating

the positions of the valves and piston during the exhaust stroke

for the second combustion chamber.

An internal combustion engine platform

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any desired operational considerations; two of the
most important operational considerations are the
structural/thermal consideration and complete burning/
higher power output consideration. In the structural/
thermal consideration, the engine’s operation is limited
by the predetermined maximum combustion gas pres-
sure or temperature. In this case, early closing of
chamber valve 1 and early opening of chamber valve 2
during the exhaust stroke for combustion chamber 1
(the fourth stroke) may be selected, so that combustion
chamber 2 may be largely open to the cylinder space
during the combustion period of chamber 2. Moreover,
the charge ignition or fuel injection timing for the
charge in combustion chamber 2 may be set close to the
end of the first exhaust stroke related to combustion
chamber 1, so that an excessive pressure rise in com-
bustion chamber 2 or a prolonged high temperature
heating period on the combustion chamber wall
may be avoided. In other words, similar to the open-
chamber combustion in chamber 1, the combustion
in chamber 2 is also largely under an open-chamber
condition. It should be recognized that due to a lower
charge

temperature

compared

to

the

maximum

cylinder head temperature when the charge is being
enclosed within the second combustion chamber before
the charge is ignited, this structural/thermal considera-
tion may also provide a cooling effect on the combus-
tion chamber wall to reduce its peak temperature.
Similar to the concept of a pre-chamber or divided
chamber, this cooling effect as well as the mixing
between the fuel and air may be enhanced through
shaping the passage between the combustion chamber
and cylinder space, so that the charge would rotate
rapid within the combustion chamber toward the end of
the compression stroke.

For the second consideration, complete burning

and higher power output are the priority of the
operation. The ignition or fuel injection for combus-
tion chamber 2 may be set at an earlier time, so that the
combustion in combustion chamber 2 is substantially
complete and the gas pressure rose to the maximum
before chamber valve 2 is opened. Because of a longer
combustion time period and the closed-chamber
combustion, combustion in chamber 2 may be more
complete and peak gas pressure in the cycle would be
higher compared to the operation of a conventional
four-stroke cycle engine.

The spirit of the present cycle to increase the

expansion ratio is not limited to the situation of
two combustion chambers per cylinder as illustrated
above. The same principle is applicable to three or
more combustion chambers per cylinder. Suppose
that

the

number

of

combustion

chambers

per

cylinder is n, where n is an integer, the relation bet-
ween the expansion ratio and the compression ratio
would be

r

e

¼ nðr

c

1Þ11

ð3Þ

and the ratio of the expansion ratio to the compression
ratio is therefore

r

e

=r

c

¼ nð1 1=r

c

Þ11=r

c

ð4Þ

For r

c

5

9, n 5 3, the above two relations give

r

e

5

25, and r

e

/r

c

5

2.78. In general, for n combustion

chambers per cylinder, the engine would operate on a
cycle having (2n12) essential strokes according to the
present engine platform. While theoretically an engine
may be equipped with any number of combustion
chambers per cylinder, it is believed that the number of
combustion chambers per cylinder may be limited to 2
or 3 for most practical applications. Also, once the
exhaust is expanded to a pressure close to the ambient
pressure, a further expansion through the increase
in the number of combustion chambers is no longer
necessary.

The cycle described above bears similarity to the

Atkinson cycle in that both expansion and exhaust
strokes are increased compared to the intake and
compression strokes. In an Atkinson cycle, the increase
in the expansion/exhaust strokes is realized through a
complex linkage mechanism and an enlarged engine
size. In the present cycle, however, the increase of
expansion/exhaust strokes is realized through an
increase in the required crank angle to complete the
cycle. The additional two strokes for the present two-
chamber configuration are equivalent to the increased
portions of the expansion and exhaust strokes in a
four-stroke Atkinson engine. In the present engine, the
engine size remains the same as that of a conventional
four-stroke engine without involving a complex link-
age mechanism. Despite these physical differences, the
performance of the present cycle can be analyzed using
a similar thermodynamic cycle for an over expanded/
Atkinson cycle [1], as shown in Figure 8. It should be
pointed out that the P-V diagram in Figure 8 is for the
working fluid associated with an individual chamber
(either chamber 1 or chamber 2, for instance). Since
more than one chamber shares a compression stroke,
the working fluid volume at the beginning of the
compression stroke is less than the cylinder volume
when the piston is at the bottom dead center (BC). The

Figure 8. An ideal gas cycle for an over expanded engine.

An internal combustion engine platform

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686

expansion stroke 3–4 as shown in the figure is clearly
greater than the compression stroke 1–2. The maxi-
mum useful expansion stroke is 3–5

0

until the exhaust

gas pressure reaches the ambient pressure. One may
notice that in this figure, the expansion/exhaust
volume has been replaced by r

e

V

c

5

(n(r

c

1)11)V

c

, in

accordance with the Equation (3) of the present cycle.

The similar relation for the calculation of the

thermal efficiency of an over expanded gas cycle from
Heywood [1] and Ferguson and Kirkpatrick [2] can
also be used for the present cycle for the thermal
efficiency Z:

Z ¼ 1 ðlr

c

Þ

1g

l

1g

lð1 gÞ g

ðg 1Þ

P

1

V

1

Q

in

ð5Þ

where g 5 c

p

/c

v

, and l 5 r

e

/r

c

5 n

(11/r

c

)11/r

c

for the

present cycle.

Thermal efficiencies of the present cycle with different

compression ratio (r

c

) are calculated and the results are

compared with the corresponding results from the
Otto cycle, as shown in Figure 9. In agreement with the
results for over-expanded cycles by Heywood [1] and
Ferguson and Kirkpatrick [3], the improvement in
thermal efficiency is generally about 20%.

3. CYCLES FOR COMBUSTION
UNDER CONSTANT VOLUME

The description of the operational benefits of an
internal combustion engine according to the present
cycles so far focuses mainly on the increase of
expansion ratio to increase the thermal efficiency.
Additional description is needed for the advantages
and implementation of the closed-chamber combustion
that is made possible by the present engine platform.
The closed-chamber combustion may have many
potential operational benefits such as avoiding power
loss related to early fuel ignition or fuel injection
before the piston reaches the TC in the compression
stroke and having a shorter warm-up time and a longer

combustion time for the fuel, which would result
in a more complete combustion. These characteris-
tics may be particularly important to a diesel engine
or an internal combustion engine operating on
alternative fuels such as ethanol and methanol.
Additional benefits include providing a higher com-
bustion gas pressure at the start of the power stroke
and avoiding fuel contact with cooler piston and
cylinder wall in the power stroke for a compression–
ignition engine.

The six essential stroke cycle in connection with the

illustrations in Figures 1–7 may enable closed-chamber
combustion for the second combustion chamber but
not for the first combustion chamber. To enable
closed-chamber combustion for both the second and
first chambers, the engine may operate on a cycle
including eight essential strokes based on the two
combustion chamber configuration. In the following,
the eight-stroke cycle is described based on the engine
platform as shown in Figure 1, but not necessarily
the same condition of charge flow and valve positions.
The eight strokes may include (1) a first intake stroke,
where an amount of charge is admitted into the
cylinder space, (2) a first compression stroke for
combustion chamber 1 following the first intake
stroke, where the charge is compressed to an elevated
pressure within the first combustion chamber, (3) a
second intake stroke for combustion chamber 2
following the compression stroke for combustion
chamber 1, allowing an amount of charge into the
cylinder space, while combustion chamber 1 is closed
and encloses the charge entering combustion chamber
1 in the first compression stroke, (4) a second com-
pression stroke for combustion chamber 2 following
the second intake stroke, in which the charge is com-
pressed to an elevated pressure within combustion
chamber 2, while combustion chamber 1 is closed.
During the time period of the second intake and sec-
ond compression strokes, combustion takes place in
combustion chamber 1, (5) a first power stroke for
combustion chamber 1 following the second compres-
sion stroke, where high pressure, high temperature
gases from combustion chamber 1 expand into the
cylinder space and deliver work to the piston, while
combustion chamber 2 is closed and encloses the
charge entering combustion chamber 2 in the second
compression stroke, (6) a first exhaust stroke for
combustion chamber 1 following the first power
stroke, allowing the exhaust gases to be discharged out
of the cylinder while combustion chamber 2 is closed.
During the time period of the first power and first
exhaust strokes, combustion takes place in combustion
chamber 2, (7) a second power stroke for combustion
chamber 2 following the first exhaust stroke, where
high pressure, high temperature combustion gases
from combustion chamber 2 expand into the cylinder
space and deliver work to the piston, while combustion
chamber 1 is closed, and (8) a second exhaust stroke

Figure 9. Improvement of thermal efficiency with different

compression ratios.

An internal combustion engine platform

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687

for combustion chamber 2 following the second power
stroke, allowing exhaust gases to be discharged out of
the cylinder space.

Figure 10 shows schematically the pressure variation

with crank angle for both combustion chambers 1 and 2.
Process 1–2 is the intake stroke related to combustion
chamber 1, followed by a compression stroke 2–3. From
3–4, the charge is enclosed within the chamber, and the
timing of the combustion may determine the peak pres-
sure and temperature, hence it is preferably near the end
of this period if the heat loss is a serious concern. Process
4–5 is related to the gas expansion stroke including
exhaust gas blowdown when the piston is at BC. Process
5–6 is the exhaust stroke associated with chamber 1, and
afterwards in 6–7, chamber 1 may be closed, maintaining
a pressure close to the exhaust pressure. It should be
pointed out that the complete closure of chamber 1 from
6–7 is an option; it may be opened from 5

0

to 6

0

to aid the

gas exhausting for chamber 2. The curve representing the
pressure variation for chamber 2 is similar to that of
chamber 1 except that it is shifted to the left or to the
right by a crank angle of 3601. Similarly, the complete
closure of chamber 2 from 6

0

to 7

0

is an option; it may

remain open from 1–2 to aid the intake associated with
chamber 1. Also, the roles of chambers 1 and 2 may be
exchanged in the next cycle. In this eight-stroke cycle,
each combustion chamber is given a maximum 3601
crank angle for combustion under a closed-chamber
condition, which may provide a sufficiently long time
period for the completion of the combustion with a
relatively short charge holding time period.

Compared to the six essential stroke cycle discussed

earlier, the eight-stroke cycle includes two power
strokes, resulting in one power stroke per four strokes as
is the case for a conventional four-stroke cycle. There-
fore, its power output could be above or equivalent to a
conventional four-stroke cycle without scarifying the
mean effective pressure. However, it does not have the
benefit of an increased expansion ratio.

4. CYCLES FOR VARIABLE
COMPRESSION RATIO

The description of the engine cycles so far are related
mainly to increasing expansion ratio or closed-chamber

combustion. With equal importance, engine cycles that
enable variable compression ratios are described. It is
well known that the compression ratio in an engine
design is determined primarily based on knock thresh-
olds at the wide-open throttle condition. However, this
knock threshold may not be applicable to the part
throttle condition under part load conditions, possibly
allowing for a higher compression ratio to increase the
engine’s thermal efficiency [4]. Since the majority of an
engine’s operating time occurs under part load condi-
tions, the incentive to enable a variable compression
ratio functionality may be significant.

In the following discussion regarding the variable

compression ratio mechanism, the engine is based on a
two-chamber configuration although the principle may
be equally applicable to the configuration having more
than two combustion chambers per cylinder. First, when
the engine operates under part load conditions, one of
the combustion chambers may be closed and deacti-
vated. Consider again the example discussed earlier
having a compression ratio of 9 when both combustion
chambers are active. If one of the combustion chambers
is closed and deactivated, the active combustion
chamber would operate on a four-stroke cycle with a
compression ratio of 17. As a result, the engine may
operate on a much higher compression ratio without
encountering the knock condition for a spark-ignition
engine. To reduce the combustion chamber wall tem-
perature, the two combustion chambers may be alter-
nately activated and deactivated under the control of an
electronic control unit. With this arrangement, each
combustion chamber is given a fairly long period of
‘resting’ time between active duties and the combustion
wall temperature may be substantially reduced. The
increased compression ratio discussed here would
represent an upper limit under the condition of a single-
active combustion chamber. Under many other working
conditions with a variety of power requirements, how-
ever, the active combustion chamber may be required to
operate at an intermediate compression ratio that is
lower than the upper limit and varies constantly
under an engine’s real-time operating conditions. This
requirement may be met through a suitable control
algorithm of an electronic control unit in connection
with the charge sharing between the two combustion
chambers as well as alternating activating and deacti-
vating of the two chambers.

5. CYCLES FOR COLD START

It is well known that ignition delay is an issue for the
cold start of an engine. This issue is especially
important to a diesel engine or an engine burning
alternative fuels such as ethanol and methanol. It is
also known that the ignition delay is a strong function
of cetane number of a fuel [5,6]. Cold start problems

Figure 10. Pressure variations in the two combustion cham-

bers at different crank angles.

An internal combustion engine platform

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682 690 © 2011 John Wiley & Sons, Ltd.

688

with alternative fuels are well documented because of
their low cetane number. The engine platform presented
in this paper could potentially solve this cold start
problem. During the cold start process, the engine based
on the present platform may be operated on a cycle that
enables combustion under constant volume as discussed
earlier, so that the cycle could provide a sufficiently large
crank angle for the ignition of the fuel.

Another problem associated with the cold start is

the ignition temperature for a compression–ignition
engine, which is especially severe for ethanol or
methanol that has a high ignition temperature. This
problem could also be solved through the present
engine platform. During the cold start process, only
one combustion chamber may remain active with the
other chamber or chambers associated with the same
cylinder being closed. With this arrangement, the
compression ratio of the active combustion chamber
could be nearly doubled or tripled. This doubled or
tripled compression ratio could generate a sufficiently
high charge temperature at the end of the compression
stroke for a successful cold start of the engine. An
additional cold start mechanism is also discussed by
Cao [7], which is beyond the scope of this paper.

Finally, the present engine platform that enables

closed-chamber combustion may solve the corrosion
problem of an alternative fuel, such as ethanol or
methanol. It is known that high alcohol concentrations
may inhibit upper engine lubrication and corrode
engine components such as the cylinder wall and crank
bearings, which may lead to premature engine failure.
In some situations, it may require a lubricant different
from that used in a gasoline or diesel engine. For an
internal combustion engine having the feature of closed-
chamber combustion according to the present engine
platform, particularly for a compression–ignition engine
with air as the intake charge, the fuel is directly injected
into the closed combustion chamber and combustion is
substantially complete (almost no alcohol left) before
the gases resulting from combustion are released into the
cylinder. As a result, the problems associated with the
lubrication and corrosion due to a high concentration of
an alcohol are substantially eliminated. In addition,
because of the ample charge holding time, the fuel
injection rate into the combustion chamber may be
adjusted to avoid so-called ‘diesel knock’ for a fuel
having a low cetane number [1].

6. DISCUSSION OF THE PRESENT
CYCLES

It is understandable that one of the most important
components to enable the present cycles is the
combustion chamber valve. Although the puppet valve
is the most established type of valve in the current
internal combustion engine industry, it may be subject

to damage by the piston as the valve is required to
move into the cylinder space. Furthermore, a large
pressure difference between the enclosed combustion
chamber and the cylinder space may cause the
difficulty of opening up the valve when needed. As a
result, a preferred type of valve for the present
application may be a slide valve. To avoid excessive
combustion gas heating, the slide valve may be
installed in a peripheral location of the piston and
cylinder assembly [7]. Moreover, the localized cooling
of the reciprocating slide valve member could also be
done using the technique of reciprocating heat pipes
developed by Cao and Wang [8,9].

In addition to the unique operational benefits asso-

ciated with each individual cycle, the engine introduced
in this paper may have a distinct advantage of pro-
viding very flexible operating modes on the basis of a
portfolio of thermodynamic cycles. For example, based
on the same engine platform/hardware and controlled
by an electronic engine control unit, the engine may
operate on the cycle for cold start, and the eight stroke
cycle with two power strokes per cycle when the vehicle
accelerates, which is especially important for a diesel
powered engine to avoid excessive smoke emission.
During the majority of the driving time on a highway
or a road in a rural area, the engine may operate on the
six essential stroke cycle to provide sufficiently high
power output and at the same time to save energy.
Also, during the quick acceleration or under a light
load condition, the engine may switch, respectively, to
the conventional four-stroke cycle or the variable
compression ratio cycle introduced in this paper.

It should be noted that although the discussions of

the present cycles are based largely on an Otto-type
spark-ignition internal combustion engine or a diesel-
type compression–ignition engine, the spirit of the
cycles may also be applicable to other types of engines,
such as gas-burning engines including natural-gas
burning engines and hydrogen-burning engines, two-
stroke type internal combustion engines, and engines
with alternative structures or fuel ignition means,
such as paired piston engines, free piston engines, and
homogenous

charge

compression–ignition

engines

where fuel ignition or the fuel injection device,
as shown in Figures 1–7, may not be necessary.
Additionally, an internal combustion engine according
to the present cycles may be employed in a hybrid
electric vehicle, which incorporates an internal com-
bustion engine with an electric motor and storage
batteries. With a hybrid electric vehicle platform, any
reduction in the mean effective pressure during the
vehicle acceleration may be more than compensated by
the electric power from the combination of the motor
and batteries.

Finally, for the engine platform as presented in this

paper with a two-chamber configuration, each com-
bustion chamber is provided with a chamber valve.
However, the configuration of a single chamber valve

An internal combustion engine platform

Y. Cao

Int. J. Energy Res. 2012; 36:

–

DOI: 10.1002/er

682 690 © 2011 John Wiley & Sons, Ltd.

689

may also be possible. For example, combustion
chamber 2 as shown in Figure 1 may be provided with
a chamber valve, but chamber 1 may be directly open
to the cylinder space without a chamber valve. Under
this configuration, the engine may also operate on
a six-stroke cycle with the penalty of a reduced
improvement in thermal efficiency. In this six-stroke
cycle, the power stroke of the combustion gas from
chamber 1 will still have an increased expansion ratio
similar to the two-chamber valve configuration.
However, the possibility of an increased expansion
ratio for the power stroke of the gas in chamber 2 may
be decreased due to the open space of chamber 1.

7. CONCLUSIONS

It is believed that introduced thermodynamic cycles
associated with the engine platform in this paper may
enable a future internal combustion engine that could
potentially increase the thermal efficiency by about
20% under full and part load conditions and overcome
the cold start problem associated with diesel fuels
or alternative fuels such as ethanol and methanol.
In addition to the above-mentioned benefits, the engine
platform would enable flexible operating cycles to
optimize the engine performance under given load
conditions and environmental considerations. Finally,
the cycles introduced in this paper may represent some
new concepts of engine operation from a thermo-
dynamics point of view for various types of internal
combustion engines.

NOMENCLATURE

c

p

5

specific heat at constant pressure

c

v

5

specific heat at constant volume

n

5

the number of combustion chambers
per cylinder

P

1

5

intake pressure

Q

in

5

heat transfer per unit mass of working
fluid

r

c

5

compression ratio

r

e

5

expansion ratio

T

1

5

intake temperature

V

c

5

clearance volume

V

d

5

piston displacement volume

V

tdc

5

charge volume at top dead center

Subscripts

1

5

chamber 1

2

5

chamber 2

TC

5

top dead center

BC

5

bottom dead center

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An internal combustion engine platform

Y. Cao

Int. J. Energy Res. 2012; 36:

–

DOI: 10.1002/er

682 690 © 2011 John Wiley & Sons, Ltd.

690