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1. INTRODUCTION
A diesel engine is an internal combustion engine that uses the heat of
compression to initiate ignition to burn the fuel, which is injected into the combustion
chamber during the final stage of compression. Diesel engines have wide range of
utilization for automobiles, locomotives & marines and co-generation systems.
However, large problem is still related to undesirable emission.
The six-stroke engine is a type of internal combustion engine based on the
four-stroke engine but with additional complexity to make it more efficient and
reduce emissions. Two different types of six-stroke engine have been developed:
In the first approach, the engine captures the heat lost from the four-stroke
Otto cycle or Diesel cycle and uses it to power an additional power and exhaust stroke
of the piston in the same cylinder. Designs use either steam or air as the working fluid
for the additional power stroke. The pistons in this type of six-stroke engine go up
and down three times for each injection of fuel. There are two power strokes: one
with fuel, the other with steam or air. The currently notable designs in this class are
the Crower Six-stroke engine invented by Bruce Crower of the U.S. ; the Bajulaz
engine by the Bajulaz S.A. company of Switzerland; and the Velozeta Six-stroke
engine built by the College of Engineering, at Trivandrum in India.
The second approach to the six-stroke engine uses a second opposed piston in
each cylinder that moves at half the cyclical rate of the main piston, thus giving six
piston movements per cycle. Functionally, the second piston replaces the valve
mechanism of a conventional engine but also increases the compression ratio . The
currently notable designs in this class include two designs developed independently:
the Beare Head engine, invented by Australian Malcolm Beare, and the German
Charge pump , invented by Helmut Kottmann.
http://en.wikipedia.org/w/index.php?title=Bajulaz_engine&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=Bajulaz_engine&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=Velozeta_Six-stroke_engine&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=Velozeta_Six-stroke_engine&action=edit&redlink=1http://en.wikipedia.org/wiki/Opposed_piston_enginehttp://en.wikipedia.org/wiki/Compression_ratiohttp://en.wikipedia.org/wiki/Compression_ratiohttp://en.wikipedia.org/wiki/Beare_Headhttp://en.wikipedia.org/w/index.php?title=German_Charge_pump&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=German_Charge_pump&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=Velozeta_Six-stroke_engine&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=Velozeta_Six-stroke_engine&action=edit&redlink=1http://en.wikipedia.org/wiki/Opposed_piston_enginehttp://en.wikipedia.org/wiki/Compression_ratiohttp://en.wikipedia.org/wiki/Beare_Headhttp://en.wikipedia.org/w/index.php?title=German_Charge_pump&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=German_Charge_pump&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=Bajulaz_engine&action=edit&redlink=1http://en.wikipedia.org/w/index.php?title=Bajulaz_engine&action=edit&redlink=1 -
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To improve exhaust emissions from diesel engines, a new concept of Six Stroke
Engine has been proposed. This engine has a second compression and combustion
processes before exhaust process.
Fig 1 Diesel engine sectional view Fig 2 Ideal Otto cycle
Fig 3 Pressure- Volume diagrams for dual cycle
As the fuel in one cycle was divided into two combustion processes and the
EGR (Exhaust Gas Recirculation) effect appeared in the second combustion process,
the decreased maximum cylinder temperature reduced Nitrous Oxide (NO)
concentration in the exhaust gas. It was further confirmed that soot formed in the first
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combustion process was oxidized in the second combustion process .Therefore, a six
stroke diesel engine has significant possibilities to improve combustion process
because of its more controllable factors relative to a conventional four-stroke engine.
Since the cylinder temperature before the second combustion process is high
because of an increased temperature in the first combustion process, ignition delay in
the second combustion process should be shortened. In addition, typically less
desirable low cetane number fuels might also be suitable for use in the second
combustion process, because the long ignition delays of these fuels might be
improved by increased cylinder temperatures from the first combustion process.
Methanol was chosen as the fuel of the second combustion. The cetane
number of methanol is low and it shows low ignitability. However, since methanol
will form an oxidizing radical (OH) during combustion, it has the potential to reduce
the soot produced in the first combustion process.
Fig 4 Comparison of 4 stroke and 6 stroke cycle
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2. BAJULAZ SIX STROKE ENGINE
The majority of the actual internal combustion engines, operating on different
cycles have one common feature, combustion occurring in the cylinder after each
compression, resulting in gas expansion that acts directly on the piston (work) and
limited to 180 degrees of crankshaft angle.
According to its mechanical design, the six-stroke engine with external and
internal combustion and double flow is similar to the actual internal reciprocating
combustion engine. However, it differentiates itself entirely, due to its
thermodynamic cycle and a modified cylinder head with two supplementary
chambers: Combustion, does not occur within the cylinder within the cylinder but in
the supplementary combustion chamber, does not act immediately on the piston, and
its duration is independent from the 180 degrees of crankshaft rotation that occurs
during the expansion of the combustion gases (work).
The combustion chamber is totally enclosed within the air-heating chamber.
By heat exchange through the glowing combustion chamber walls, air pressure in the
heating chamber increases and generate power for an a supplementary work stroke.
Several advantages result from this, one very important being the increase in thermal
efficiency. IN the contemporary internal combustion engine, the necessary cooling of
the combustion chamber walls generates important calorific losses.
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2.1 Analysis:
Six-stroke engine is mainly due to the radical hybridization of two- and four-
stroke technology. The six-stroke engine is supplemented with two chambers, which
allow parallel function and results a full eight-event cycle: two four-event-each
cycles, an external combustion cycle and an internal combustion cycle. In the internal
combustion there is direct contact between air and the working fluid, whereas there is
no direct contact between air and the working fluid in the external combustion
process. Those events that affect the motion of the crankshaft are called dynamic
events and those, which do not effect are called static events.
Fig 5 Prototype of Six stroke engine internal view
1. Intake valve, 2.Heating chamber valve,
3.Combustion chamber valve, 4. Exhaust valve,
5.Cylinder, 6.Combustion chamber,
7. Air heating chamber, 8.Wall of combustion chamber,
9.Fuel injector and 10.Heater plug.
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2.1.1 Analysis of events
Fig 6 Event 1: Pure air intake in the cylinder (dynamic event)
1. Intake valve.
2. Heating chamber valve
3. Combustion chamber valve.
4. Exhaust valve
5. Cylinder
6. Combustion chamber.
7. Air heating chamber.
8. Wall of combustion chamber.
9. Fuel injector.
10. Heater plug.
Fig 7 Event 2: Pure air compression in the heating chamber.
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Event 3: Keeping pure air pressure in closed chamber where a maximum heat
exchange occurs with the combustion chambers walls, without direct action on the
crankshaft (static event).
Fig 8 Event 4: Expansion of the Super heat air in the cylinder work (dynamic Event).
Fig 9 Event 5: Re-compressions of pure heated air in the combustion chamber (dynamic event).
Events 6: fuel injection and combustion in closed combustion chamber, without direct
action on the crankshaft (static event).
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Fig 10 Events 7: Combustion gases expanding in the cylinder, work (dynamic event).
Fig 11 Events 8: Combustion gases exhaust (dynamic event).
Fig 12 Six-stroke engine cycle diagram:
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2.1.2 External combustion cycle: (divided in 4 events):
No direct contact between the air and the heating source.
e1. (Event 1) Pure air intake in the cylinder (dynamic event).
e2. (Event 2) Compression of pure air in the heating chamber (dynamic event).
e3. (Event 3) Keeping pure air pressure in closed chamber where a maximum heat
exchange occurs with the combustion chambers walls, without direct action on the
crankshaft (static event).
e4. (Event 4) Expansion of the super heated air in the cylinder, work (dynamic event).
2.1.3 Internal combustion cycle: (divided in 4 events)
Direct contact between the air and the heating source.
I1. (Event 5) Re-compression of pure heated air in the combustion chamber (dynamic
event)
I2. (Event 6) Fuel injection and combustion in closed combustion chamber, without
direct action on the crankshaft (static event).
I3. (Event 7) Combustion gases expanding in the cylinder, work (dynamic event).
I4. (Event 8) Combustion gases exhaust (dynamic event).
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2.2 Constructional details:
The sketches shows the cylinder head equipped with both chambers and four
valves of which two are conventional (intake and exhaust). The two others are made
of heavy-duty heat-resisting material. During the combustion and the air heating
processes, the valves could open under the pressure within the chambers. To avoid
this, a piston is installed on both valve shafts which compensate this pressure. Being a
six-stroke cycle, the camshaft speed in one third of the crankshaft speed.
The combustion chambers walls are glowing when the engine is running.
Their small thickness allows heat exchange with the air-heating chamber, which is
surrounding the combustion chamber. The air-heating chamber is isolated from the
cylinder head to reduce thermal loss.
Through heat transfer from the combustion chamber to the heating chamber,
the work is distributed over two strokes, which results in less pressure on the piston
and greater smoothness of operation. In addition, since the combustion chamber is
isolated from the cylinder by its valves, the moving parts, especially the piston, are
not subject to any excessive stress from the very high temperatures and pressures.
They are also protected from explosive combustion or auto-ignition, which are
observed on ignition of the air-fuel mixture in conventional gas or diesel engines.
The combustion and air-heating chambers have different compression ratio.
The compression ratio is high for the heating chamber, which operates on an external
cycle and is supplied solely with pure air. On the other hand, the compression ratio is
low for the combustion chamber because of effectively increased volume, which
operates on internal combustion cycle.
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The combustion of all injected fuel is insured, first, by the supply of preheated
pure air in the combustion chamber, then, by the glowing walls of the chamber, which
acts as multiple spark plugs. In order to facilitate cold starts, the combustion chamber
is fitted with a heater plug (glow plug). In contrast to a diesel engine, which requires a
heavy construction, this multi-fuel engine, which can also use diesel fuel, may be
built in a much lighter fashion than that of a gas engine, especially in the case of all
moving parts.
Injection and combustion take place in the closed combustion chamber,
therefore at a constant volume, over 360 degrees of crankshaft angle. This feature
gives plenty of time for the fuel to burn ideally, and releases every potential calorie
(first contribution to pollution reduction). The injection may be split up, with dual
fuel using the SNDF system (Single Nozzle, Dual Fuel). The glowing walls of the
combustion chamber will calcite the residues, which are deposited there during fuel
combustion (second contribution to pollution reduction).
As well as regulating the intake and exhaust strokes, the valves of the heating
and the combustion chambers allow significantly additional adjustments for
improving efficiency and reducing noise.
2.3 Factors Contributing To the Increased Thermal Efficiency,
Reduced Fuel Consumption, and Pollutant Emission
1. The heat that is evacuated during the cooling of a conventional engines
cylinder head is recovered in six-stroke engine by air-heating chamber
surrounding the combustion chamber.
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2. After intake, air is compressed in the heating chamber and heated through
720 degrees of crankshaft angle, 360 degrees of which in closed chamber
(external combustion).
3. The transfer of heat from thin walls of the combustion chamber to the air
heating chambers lowers the temperature, pressure of gases on expansion
and exhaust (internal combustion).
4. Better combustion and expansion of gases that take place over 540 degrees
of crankshaft rotation, 360 of which is in closed combustion chamber, and
180 for expansion.
5. Elimination of the exhaust gases crossing with fresh air on intake. In the six
stroke-engines, intake takes place on the first stroke and exhaust on the
fourth stroke.
6. Large reduction in cooling power. The water pump and fan outputs are
reduced. Possibility to suppress the water cooler.
7. Less inertia due to the lightness of the moving parts.
8. Better filling of the cylinders on the intake due to the lower temperature of
the cylinder walls and the piston head.
9. The glowing combustion chamber allows the finest burning of any fuel and
calcinate the residues.
10. Distribution of the work: two expansions (power strokes) over six strokes,
or a third more than the in a four-stroke engine.
Since the six-stroke engine has a third less intake and exhaust than a four stroke
engine, the depression on the piston during intake and the back pressure during
exhaust are reduced by a third. The gain in efficiency balances out the losses due tothe passage of air through the combustion chamber and heating chamber valves,
during compression of fresh and superheated air. Recovered in the six-stroke engine
By the air-heating chamber surrounding the combustion. Friction losses,
theoretically higher in the six-stroke engine, are balanced by a better distribution of
pressure on the moving parts due to the work being spread over two strokes and the
elimination of the direct combustion.
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3. DUAL FUEL SIX STROKE ENGINE
3.1 Working
The cycle of this engine consists of six strokes:
1. Intake stroke
2. First compression stroke
3. First combustion stroke
4. Second compression stroke
5. Second combustion stroke6. Exhaust stroke
Fig 13 Concept of a Six-stroke diesel engine
3.1.1 Intake or Suction stroke
To start with the piston is at or very near to the T.D.C., the inlet valve is open
and the exhaust valve is closed. A rotation is given to the crank by the energy from a
flywheel or by a starter motor when the engine is just being started. As the piston
moves from top to bottom dead centre the rarefaction is formed inside the cylinder i.e.
the pressure in the cylinder is reduced to a value below atmospheric pressure. The
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pressure difference causes the fresh air to rush in and fill the space vacated by the
piston. The admission of air continues until the inlet valve closes at B.D.C.
3.1.2 First Compression stroke
Both the valves are closed and the piston moves from bottom to top dead
centre. The air is compressed up to compression ratio that depends upon type of
engine. For diesel engines the compression ratio is 12-18 and pressure and
temperature towards the end of compression are 35-40 kgf/cm 2 and 600-700 0C
3.1.3 First combustion stroke
This stroke includes combustion of first fuel (most probably diesel) and
expansion of product of combustion. The combustion of the charge commences when
the piston approaches T.D.C.
Here the fuel in the form of fine spray is injected in the combustion space. The
atomization of the fuel is accomplished by air supplied. The air entering the cylinder
with fuel is so regulated that the pressure theoretically remains constant during
burning process.
In airless injection process, the fuel in finely atomized form is injected in
combustion chamber. When fuel vapors raises to self ignition temperature, the
combustion of accumulated oil commences and there is sudden rise in pressure atapproximately constant volume. The combustion of fresh fuel injected into the
cylinder continues and this ignition is due to high temperature developed in engine
cylinder. However this latter combustion occurs at approximately constant pressure.
Due to expansion of gases piston moves downwards. The reciprocating
motion of piston is converted into rotary motion of crankshaft by connecting rod and
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crank. During expansion the pressure drop is due to increase in volume of gases and
absorption of heat by cylinder walls.
3.1.4 Second compression stroke
Both the valves are closed and the piston moves from bottom to top dead
centre. The combustion products from the first compression stroke are recompressed
and utilized in the second combustion process before the exhaust stroke. In typical
diesel engine combustion the combustion products still contains some oxygen.
3.1.5 Second combustion stroke
This stroke includes combustion of second fuel having low cetane (Cetane
number of fuel is defined as percent volume of cetane (C 16H34) in a mixture of cetane
and alpha-methyl-naphthalene that produces the same delay period or ignition lag as
the fuel being tested under same operating conditions on same engine). The
combustion of the charge commences when the piston approaches to TDC.
The second fuel injected into recompressed burnt gas can be burnt in the
second combustion process. In other words combustion process of the second fuel
takes place in an internal full EGR (Exhaust Gas Recirculation) of the first
combustion. This second combustion process was the special feature of the proposed
Six Stroke DI Diesel Engine.
3.1.6 Exhaust stroke
The exhaust valve begins to open when the power stroke is about to complete.
A pressure of 4-5 kgf/cm 2 at this instant forces about 60% of burnt gases into the
exhaust manifold at high speed. Much of the noise associated with automobile engine
is due to high exhaust velocity. The remainder of burnt gases is cleared of the swept
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volume when the piston moves from TDC to BDC. During this stroke pressure inside
the cylinder is slightly above the atmospheric value. Some of the burnt gases are
however left in the clearance space. The exhaust valve closes shortly after TDC.
The inlet valve opens slightly before the end of exhaust stroke and cylinder is
ready to receive the fresh air for new cycle. Since from the beginning of the intake
stroke the piston has made six strokes through the cylinder (Three up And Three
down). In the same period crank shaft has made three revolutions. Thus for six stroke
cycle engine there are two power strokes for every three revolutions of crank shaft.
3.2 Performance analysis
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3.2.1 Modification over four stroke diesel engine
This six-stroke diesel engine was made from a conventional four-stroke diesel
engine with some modification. A sub-shaft was added to the engine, in order to drive
a camshaft and injection pumps. The rotation speed of the sub-shaft was reduced to
1/3 of the rotation of an output shaft. To obtain similar valve timings between a four-
stroke and a six-stroke diesel engine, the cam profile of the six-stroke diesel engine
was modified. In order to separate the fuels, to control each of the injection timings
and to control each injection flow rate in the first and the second combustion
processes, the six-stroke diesel engine was equipped with two injection pumps and
two injection nozzles. The injection pumps were of the same type as is used in the
four-stroke diesel engine.
The nozzle is located near the center of a piston cavity, and has four injection
holes. For the six-stroke diesel engine, one extra nozzle was added on the cylinder head. This extra nozzle was of the same design as that of the four-stroke engine.
Fig 14 Volume Angle diagram for six stroke engine
Diesel fuel for the first combustion process was injected through this extra
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nozzle, and methanol for the second combustion process was injected through the
center nozzle. Here, we denoted the injection timing of the four stroke diesel engine
as X i. The injection timings of the first and second combustion strokes for the six-
stroke diesel engine are shown as X i I and X i II, respectively. Crank angle X was
measured from the intake BDC. In the six-stroke engine, crank angle of the first
combustion TDC is 180 degrees. The second combustion TDC is 540 degrees.
Specifications of the test engines are shown in Table 1. The conventional four-
stroke diesel engine that was chosen as the basis for these experiments was a singlecylinder, air cooled engine with 82 mm bore and 78 mm stroke. The six-stroke engine
has the same engine specifications except for the valve timings. However, the
volumetric efficiency of the six-stroke engine showed no significant difference from
that of the four-stroke engine.
Characteristics of the six-stroke diesel engine were compared with the
conventional four-stroke diesel engine. In this paper, the engine speed (Ne) was fixed
at 2,000 rpm. Cylinder and line pressure indicators were equipped on the cylinder
head. NO concentration was measured by a chemiluminescences NO meter, and soot
emission was measured by a Bosch smoke meter.
The physical and combustion properties of diesel fuel and methanol are shown
in Table. 2. Since combustion heats of diesel fuel and methanol are different,
injection flow rates of the first and the second combustion processes are defined bythe amount of combustion heat. Here, the supplied combustion heat for the first
combustion process is denoted by Q I. The second combustion stroke is denoted by
Q II. The ratio of Q II to Q t (Q t = Q I+Q II) supplied combustion heat per cycle) is defined
as the heat allocation ratio H: H = Q II = Q II
Q I +Q II Q t
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Table 1. Specifications of the test engine:
Four stoke Six strokeDiesel Engine Diesel Engine
Engine type DI, Single cylinder, Air cooled, OHVBore x Stroke [mm] 82 x 78Displacement [cc] 412Top Clearance [mm] 0.9Cavity Volume [cc] 16Compression ratio 21
Intake Valve Open 10 0 BTDC 7 0 BTDCIntake valve Close 140 0 BTDC 145 0 BTDCExhaust Valve Open 135 0 ATDC 140 0 ATDCExhaust Valve Close 12 0 ATDC 3 0 ATDCValve Overlap 22 0 10 0
Rated power 5.9 kW /3000rpmBase Engine ----------------
Table 2. Physical and combustion properties of diesel fuel and methanol:
Diesel Fuel Methanol
Combustion heat [MJ/kg] 42.7 19.9
Cetane number 40-55 3.0
Density [kg/m 2] 840 793
Theoretical air-fuel ratio 14.6 6.5
3.3 Performance of six stroke diesel engine
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3.3.1 Comparison with four stroke diesel engine
A four-stroke engine has one intake stroke for every two engine rotations. For
the six-stroke engine, however, the intake stroke took place once for every three
engine rotations. In order to keep the combustion heat per unit time constant, the
combustion heat supplied to one six-stroke cycle should be 3 or 2 times larger than
that of the four-stroke engine.
There are many ways to compare performance between the four-stroke and
six-stroke engines. For this paper, the authors have chosen to compare thermalefficiency or SFC at same output power. If the thermal efficiency was the same in
both engines, the same output power would be produced by the fuels of equivalent
heats of combustion.
Therefore, in order to make valid comparison, fuels supplied per unit time
were controlled at the same value for both engines and engine speeds were kept
constant. In this section, fuel supplied for the engines was only a diesel fuel.Performance of the six-stroke engine was compared with that of the four-stroke
engine under various injection timings.
Detailed conditions for comparison of the four-stroke and six-stroke engines
are listed in Table. 3. The heat allocation ratio of the six-stroke engine was set at H =
0.5. Injection flow rate of fuel was Q t4 = 0.50 KJ/cycle for the four-stroke engine and
Q t6 = 0.68 KJ/cycle for the six stroke engine. For six stroke engine, it meant that the
amount of 0.34KJ was supplied at each combustion process.
At the viewpoint of combustion heat, 0.75 KJ/cycle of heat should be supplied
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for the six stroke engine to make the equivalence heat condition. However diesel fuel
of 0.68 KJ/cycle was supplied here because of difficulties associated with methanol
injection.
Injection timing of the four-stroke engine was changed from 160 degrees
(20 0BTDC) to 180 degrees (TDC). For six -stroke engine, the injection timing of the
first combustion process was fixed to 165 degrees (15BTDC) or 174 degrees
(6BTDC), and the second injection timing was changed from 520 degrees (200 0
BTDC) to 540 degrees (TDC).
Fig 15 Valve timing diagram four stroke engine
Table 3. Detailed conditions of comparison between the four stroke and six stroke diesel enginesand performance of engine
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Engine ParametersFour Stroke
Diesel EngineSix Stroke
Diesel Engine
Engine Speed Ne [rpm] 2007 2016
Supplied combustion heat per cycleQ t [KJ/cycle] 0.50 0.68
Supplied combustion heat per unit timeH t [KJ/s] 8.36 7.62
Intake air flow per cycleMa [mg/cycle] 358.7 371.4
Injection quantity per cycleM f [mg/cycle] 11.8 16
Excess air ratio 2.40 1.83
Intake air flow per unit timeM a [g/cycle] 6.00 4.16
Injection quantity per unit timeM f [g/sec] 0.197 0.179
Brake torque T b [N-m] 15.52 15.28Brake power L b [KW] 3.26 3.24
BSFC. b [ g / KW-h] 217.9 520.3IMEP P i [Kgf / cm 2] 5.94 4.37Indicated torque T i [N-m] 19.10 18.71Indicated power L i [KW] 4.01 3.75
ISFC b i [g / KW-h ] 177.2 163.3
Indicated torque of the six-stroke engine is almost same level with that of
the four-stroke engine under various injection timings. NO concentration in exhaust
gas of the six-stroke engine was lower than that of the four-stroke engine. NOemissions from both engines were reduced by the retard of injection timing. The
effect of retard in the second injection timing of the six-stroke engine was similar to
that of the retard in the four-stroke engine.
For the six-stroke engine, from the comparison between X i I = 165 degrees
(15BTDC) and X i I = 174 degrees (6BTDC), it seemed that the NO reduction effect
appeared with the timing retard in the first combustion process.
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Soot emission in the exhaust gas of the four-stroke engine was low level andit was not affected by the timing retard of injection. However, the level of soot
emission from the six-stroke engine was strongly affected by the timing of the second
injection. When the injection timing was advanced from 528 degrees (12 BTDC), it
was confirmed that the soot emission was lower than that of the four-stroke engine.
From numerical analysis, it was considered that the soot formed in the first
combustion process was oxidized in the second combustion process. On the contrary,
when the injection timing was retarded from 528 degrees (12 BTDC), soot emission
increased with the timing retard. Then, it was considered that the increased part of the
soot was formed in the second combustion process because an available period for
combustion was shortened with the retard of injection timing.
Experimental conditions were X i = X i I = 170 degrees (10 0 BTDC) and
Xi II=530 degrees (100
BTDC). The heat allocation ratio of six stroke engine wasH=0.5.
The cylinder temperature and heat release rate were calculated from the
cylinder pressure. The pattern of heat release rate in the first combustion stroke of
the six-stroke engine was similar to that of the heat release rate of the four-stroke
engine. It was the typical combustion pattern that contained a pre-mixed combustion
and diffusion combustion. On the other hand, since an increase of cylinder temperature in the second combustion process was caused by the compression of the
burned gas formed in the first combustion stroke, a pre-mixed combustion in the
second combustion process was suppressed by a short ignition delay.
The maximum cylinder temperature in the first combustion process was lower
than that in the four-stroke engine. It was caused by smaller amount of fuel which
was injected in the first combustion process. Considering these results, it was proved
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that NO concentration in the exhaust gas was reduced by the decrease of the
maximum cylinder temperature in the first combustion process and EGR effect in the
second combustion process.
The performance of these two engines could be compared by Table. 3. Since
BSFC of the six-stroke engine obtained by the brake power suffered, SFC is
compared with ISFC for the Xi = 163 degree (17 0 BTDC), ISFC of the four-stroke
engine was 177.2 g/KW-h.
On the other hand, for the X i I = 165 degrees (15 BTDC) and X i II = 523
degrees (17 0 BTDC), I.S the six-stroke engine was 163.3 g/KW-h. i.e. ISFC of the
six-stroke engine was slightly lower than that of the four-stroke engine.
It was considered that this advantage in ISFC was caused by a small cut-off
ratio of constant pressure combustion. Because, in the six-stroke engine proposed
here, the fuel divided into two combustion processes resulted in a short combustion
period of each combustion process. Furthermore, in the reduction of NO emission, the
six-stroke engine was superior to the four-stroke engine.
3.3.2 Effect of heat allocation ratio
Injection conditions were X i I = 170 degrees (100 0 BTDC) and X i II = 530
degrees (10 0 BTDC). Both fuels in the first and second combustion processes werediesel fuel. Total fuel at the combustion heat basis was Q t = 0.68 KJ/cycle. It meant a
high load in this engine because the total excess air ratio was 1.83 as previously
shown in Table 3.
The maximum value of the indicated torque appeared around H = 0.5 NO
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concentration in exhaust gas was reduced by an increase of heat allocation ratio. In
other words, NO emission decreased with an increase of the fuel of the second
combustion process.
In the case of H = 0.5, there is a relatively long ignition delay in the first
combustion process and pre-mixed combustion was the main combustion phenomena
in it. NO of high concentration was formed in this pre-mixed combustion process. On
the other hand, in the case of H = 1, diffusion combustion was the main combustion
phenomena and NO emission was low.
Soot emission in exhaust gas increased with an increase of heat allocation
ratio. Since the injection flow rate in the second combustion process increased with
an increase of the heat allocation ratio, the injection period increased with an increase
of the heat allocation ratio. It caused the second combustion process to be long, and
unburnt fuel that was the origin of soot remained after the second combustion
process.
The heat release rates on H = 0.15 and H = 0.85. For H =0.15, since injection
flow rate in the first combustion process was high and injection period in it was long,
the combustion period in the first combustion process became long as compared with
case of H = 0.85. On the other hand, for H = 0.85, the combustion period in the
second combustion process became long as compared with case of H =0.15. It was
also observed that the long combustion periods in both the first and secondcombustion were caused by the long diffusion combustion. Further, diffusion
combustion was the main combustion phenomena of the second combustion process.
When the heat allocation ratio was 0.85, the ratio of heat release rates between
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the first and second combustion should be 15: 85, however the actual ratio obtained
from the figure was 46: 54. This inconsistency was caused from the drift of the base
lines of the heat release diagrams. For H = 0.15, the actual ratio of heat release rates
was 73: 27 with the similar reason.
The cylinder temperature for the H = 0.15 condition was higher than that of
the H = 0.85 condition. This could be explained as follows. In the first combustion
stroke, since the injection flow rate of H = 0.15 was higher than that of H = 0.85, the
combustion temperature for the H = 0.15 condition was higher than that of H = 0.85.
In the second compression stroke, since the high temperature burned gas was re-
compressed, the temperature of H = 0.15 was also higher than that of H = 0.85.
As a result, the temperature at the beginning of the second combustion stroke
was high in H = 0.15 condition as compared with H = 0.85 condition. At the later
stage of the second combustion, however, the opposite relationship between these two
temperatures were observed, because the injection flow rate of the second combustion
process was low in H = 0.15 condition.
The maximum temperatures in the first and second combustion process
decreased with an increase of the heat allocation ratio. Then, it could be concluded
that the reduction of NO concentration with the heat allocation ratio, was caused by
the decrease of the cylinder temperature.
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3.4 Performance of the dual fuel six stroke diesel engine
3.4.1 Comparison with diesel fuel six stroke engine
Operating conditions of comparison between the diesel fuel and the dual fuel
six-stroke engines are shown in Table. 4. Experimental conditions were X i I= 170
degrees (10 0 BTDC), X i II = 530 degrees (10 o BTDC) and H = 0.5.
In dual fuel six-stroke engine, diesel fuel and methanol were supplied into
first and second combustion process, independently. Combustion heats supplied per
one cycle of the diesel fuel and dual fuel six-stroke engines were same. Thecombustion heat supplied per one cycle was selected as Q t = 0.43 KJ/cycle under the
middle load condition. Performance of the dual fuel six-stroke engine was compared
with the diesel fuel six-stroke engine under various injection timings in the second
combustion process. Indicated torques of both engines was revealed constant around
15 N-m. As a result, it could be concluded that states of combustion of the diesel fuel
and the dual fuel six-stroke engines had similar contributions on the engine
performance. NO emissions from the dual fuel six-stroke engine were lower thanthose of the diesel fuel six-stroke engine. This effect appeared prominently at the
advanced injection timing of the second combustion. Further, NO concentrations of
both engines were reduced by the injection timing retard in the second combustion.
Fig 16 Torque- Angle diagram for six stroke engine
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Soot emission in the exhaust gas of the diesel fuel six stroke engines increased
with a retard of the injection timing in the second combustion. For the dual fuel six-
stroke engine, the exhaust level of soot was very low under various injection timings
of the second combustion process. Soot was formed clearly by the combustion of
diesel fuel in the first combustion process and it was oxidized in the second
combustion process. Considering these results, it was possible to estimate that soot
was almost oxidized by methanol combustion in the second combustion process. This
estimation is supported by a dual fuel diesel engine operated with diesel fuel
methanol.
The combustion heat supplied per one cycle was selected as Q t = 0.68
KJ/cycle under the high load condition. Indicated torques of both engines was also
revealed constant around 20 N-m. NO concentration had the same tendency as the
cases of the middle load. Soot emission level of the diesel fuel six-stroke engine was
high in this high load condition. For the dual fuel six-stroke engine, however, soot
was very low under various injection timings of the second combustion process.
The performance of these engines was compared in Table. 4. For the second
combustion process, since combustion heats of diesel fuel and methanol were
different, injection quantities of both engines were different. BSFC and ISFC of the
dual fuel six-stroke engine was sensibly higher than that of the diesel fuel engine. To
compare the performance of these engines, injection quantity of both engines was
defined by an amount of combustion heat, and SFC should be calculated from it. As a
result, indicated specific heat consumption of the diesel fuel six-stroke engine was
5.59 MJ/KW-h, and that of the dual fuel six-stroke engine was 5.43 MJ/KW-h. For
the high load conditions shown in Table. 5, the similar advantage of the dual fuel six-
stroke engine was observed.
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Table 4. Detailed conditions of comparison between the diesel fuel and dual fuel diesel enginesand performance of engines under H = 0.5 and middle load
Diesel Fuel Six
Stroke Diesel
Engine
Dual Fuel Six
Stroke
Diesel EngineEngine Speed Ne [rpm] 2016 2003
Supplied combustion heat per cycle
Q t [KJ/cycle] 0.43
Injection quantity per cycle
(First Combustion Stroke)
M f1 [mg/cycle]
5.0
(Diesel Fuel)
Injection quantity per cycle
(Second Combustion Stroke)
M f2 [mg/cycle]
5.0
(Diesel Fuel)
10.7
(Methanol)
Excess air ratio 2.98 3.15
Brake torque T b [N-m] 3.14 3.14Brake power L b [KW] 0.66 0.66
B.S.F.C. b [ g / KW-h] 610.9 952.9I.M.E.P. P i [Kgf / cm2] 3.43 3.53
Indicated torque T i [N-m] 16.70 15.12Indicated power L i [KW] 3.1 2.77I.S.F.C. b i [g / KW-h ] 130.1 198.4
Indicated specific heat consumption
b i [MJ /KW-h] 5.59 5.43
In order to confirm the advantage of dual fuel six-stroke engine, the
performance of these engines was compared with four-stroke engine as shown in
Table. 6. NO concentrations of the diesel fuel and the dual fuel six-stroke engines
were improved with 85 - 90% as compared with that of the four-stroke engine. Soot
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emission of the diesel fuel six-stroke engine was much higher than that of the four-
stroke engine. However, for the dual fuel six-stroke engine, soot level was very low.
Furthermore, the indicated specific heat consumption of the diesel fuel and
dual fuel six-stroke engine were lower than that of the four-stroke engine. Especially,
for the dual fuel six-stroke engine, the indicated specific heat consumption was
improved with 15% as compared with that of the four stroke engine. From these
results, it could be confirmed that the dual fuel six-stroke engine was superior to the
diesel fuel six-stroke engine, and also it was superior to the four-stroke engine.
Table 6. Percentage improvements of exhaust emission and specific heat consumptionFour Stroke
Diesel Engine
Six Stroke
Diesel Engine
Dual Fuel Six
Stroke Engine NO [ppm]
( % improvement) 768
113
(85.3%)
90.5
(88.2%)Soot [%]
(%improvement) 6.8
28.8
(- 323.5%)
0
(100%)Indicated specific heat
consumption bi [MJ/KW-h]
(% improvement)
7.51 6.61
(12.0%)
6.37
(15.2%)
Table 5. Detailed conditions of comparison between the diesel fuel and dual fuel diesel engineand performance of engines under H =0.5 and high load
Six Stroke Diesel
Engine
Dual Fuel Six
Stroke EngineEngine Speed Ne [rpm] 2016 2006
Supplied combustion heat per cycle
Q t [kJ/cycle] 0.68
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Injection quantity per cycle
(First Combustion Stroke)
M f1 [mg/cycle]
8.0
(Diesel Fuel)
Injection quantity per cycle
(Second Combustion Stroke)
M f2 [mg/cycle]
8.0
(Diesel Fuel)
17.2
(Methanol)
Excess air ratio 1.86 1.93
Brake torque T b [N-m] 6.18 6.08Brake power L b [kW] 1.52 1.5
B.S.F.C. b [ g / kW.h] 504.0 777.7
I.M.E.P. P i [kgf / cm2] 4.56 4.75Indicated torque T i [N-m] 21.68 20.38Indicated power L i [kW] 3.45 2.98
I.S.F.C. b i [g / kW.h ] 155.5 236.2Indicated specific heat consumption
b i [MJ /kW.h] 6.61 6.37
3.4.2 Effect of injection timing
Performance of the dual fuel six-stroke engine under various injection timings
in the second combustion process was investigated on middle and high load.
Experimental conditions were X i I = 170 degrees (10 0 BTDC) and H = 0.5.
Performance of the dual fuel six-stroke engine under both load conditions had
the similar tendency with the timing retard. NO concentrations in the high load
condition were higher than those of the middle load condition. However, soot
emission levels of both load conditions were extremely low under various injection
timings of the second combustion.
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3.4.3 Effect of heat allocation ratio
Performance of the dual fuel six-stroke engine under various heat allocation
ratios was investigated on middle and high load. Injection conditions were X i I = 170
degrees (10 0 BTDC) and X i II = 530 degrees (10 0 BTDC). Since the combustion heat of
methanol was low, experimental range of heat allocation ratio was limited by the
smooth operation of the engine. Only the range from H = 0.25 to 0.75 (on Q t = 0.43
KJ/cycle), and from H = 0 to 0.5 (on Qt = 0.68 KJ/cycle) could be tested. .
Indicated torque increased with an increase of the heat allocation ratio. NO
concentration in exhaust gas was reduced with an increase of the heat allocation ratio.
Soot was very low, irrespective of the methanol flow rate. Even if the load condition
was high, it was concluded that soot was practically eliminated by a small amount of
methanol in the second combustion process (8% of total fuel).
4. ADVANTAGES OF SIX STROKE OVER FOUR
STROKE ENGINES
The six stroke is thermodynamically more efficient because the change in
volume of the power stroke is greater than the intake stroke, the compression stroke
and the Six stroke engine is fundamentally superior to the four stroke because the
head is no longer parasitic but is a net contributor to and an integral part of the
power generation within exhaust stroke. The compression ration can be increased
because of the absent of hot spots and the rate of change in volume during the critical
combustion period is less than in a Four stroke. The absence of valves within the
combustion chamber allows considerable design freedom.
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4.1 Main advantages of the duel fuel six-stroke engine:
4.1.1 Reduction in fuel consumption by at least 40%:
An operating efficiency of approximately 50%, hence the large reduction in
specific consumption. the Operating efficiency of current petrol engine is of the order
of 30%. The specific power of the six-stroke engine will not be less than that of a
four-stroke petrol engine, the increase in thermal efficiency compensating for the
issue due to the two additional strokes.
4.1.2 Two expansions (work) in six strokes:
Since the work cycles occur on two strokes (360 0 out of 1080 0 ) or 8%
more than in a four-stroke engine (180 0 out of 720 ), the torque is much more even.
This lead to very smooth operation at low speed without any significant effects on
consumption and the emission of pollutants, the combustion not being affected by theengine speed. These advantages are very important in improving the performance of
car in town traffic.
4.1.2 Dramatic reduction in pollution:
Chemical, noise and thermal pollution are reduced, on the one hand, in
proportion to the reduction in specific consumption, and on the other, through the
engines own characteristics which will help to considerably lower HC, CO and NOx
emissions. Furthermore, its ability to run with fuels of vegetable origin and weakly
pollutant gases under optimum conditions, gives it qualities which will allow it to
match up to the strictest standards.
4.1.3 Multifuel:
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Multifuel par excellence, it can use the most varied fuels, of any origin (fossil
or vegetable), from diesel to L.P.G. or animal grease. The difference in
inflammability or antiknock rating does not present any problem in combustion. Its
light, standard petrol engine construction, and the low compression ration of the
combustion chamber; do not exclude the use of diesel fuel. Methanol-petrol mixture
is also recommended.
5. CONCLUSIONS
The performance of the dual fuel six-stroke engine was investigated. In this dual
fuel engine, diesel fuel was supplied into the first combustion process and methanol
was supplied into the second combustion process where the burned gas in the first
combustion process was re-compressed. The results are summarized as follows.
1. Indicated specific fuel consumption (ISFC.) of the six-stroke engine proposed
here is slightly lower than that of the four-stroke engine (about 9%
improvement). NO and soot emissions from the six-stroke engine was
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improved as compared with four-stroke engine under advanced injection
timings in the second combustion stroke.
2. For the dual fuel six-stroke engine, the timing retard and an increase of heat
allocation ratio in the second combustion stroke resulted in a decrease of the
maximum temperatures in the combustion processes. It caused the reduction
of NO emission.
3. For the dual fuel six-stroke engine, soot was practically eliminated by a
small amount of methanol in the second combustion process.
4. From the comparison of the performance between the dual fuel six-stroke
and the four-stroke engine, it was concluded that indicated specific heat
consumption of the dual fuel six-stroke engine was improved with 15% as
compared with the four-stroke engine. NO concentration of the dual fuel six-
stroke engine was improved with 90%. Furthermore, soot emission was very
low in the dual fuel six-stroke engine.
5. As the fuel in one cycle was divided into two combustion processes and the
EGR effect appeared in the second combustion process, the decreased
maximum cylinder temperature reduced NO concentration in the exhaust
gas It was further confirmed that soot formed in the first combustion
process was oxidized in the second combustion process .Therefore, a six
stroke DI diesel engine has significant possibilities to improve combustion
process because of its more controllable factors relative to a conventional
four-stroke engine. Considering these results, it was confirmed that the dual
fuel six-stroke engine was superior to the four-stroke engine.
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6. REFERENCES
1. Tsunaki Hayasaki, Yuichirou Okamoto, Kenji Amagai and Masataka Arai
A Six-stroke DI Diesel Engine under Dual Fuel Operation SAE Paper No
1999-01-1500
2. S.Goto and K.Kontani, "A Dual Fuel Injector for Diesel Engines", SAE paper, No.
851584, 1985
3. Internal Combustion Engines A book by Mathur & Sharma.
4. Internal Combustion Engines Tata McGraw-hill publications,
Author V Ganesan
7. NOMENCLATURE
Ne : Engine speed
X : Crank angle
X i : Injection timing of the four-stroke diesel engine
H : Heat allocation ratio
Q : Supplied combustion heat
Q t : Supplied combustion heat per cycle
P : Cylinder pressure
V : Cylinder volume
Vs : Stroke volume
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P i : Indicated mean effective pressure (LM.E.P)
T i : Indicated torque
L i : Indicated power
T b : Brake torque
L b : Brake power
H t : Supplied combustion heat per unit time
M a : Intake air flow per cycle
M a ' : Intake air flow per unit time
M f : Injection quantity per cycle
M i : Injection quantity per unit time
: Excess air ratio
b : Brake specific fuel consumption (B.S.F.C.)
b l : Indicated specific fuel consumption (I.S.F.C.)
b i' : Indicated specific heat consumption
SUBSCRIPTS
I: first combustion stroke
II: second combustion stroke
4: four-stroke diesel engine
6: six-stroke diesel engine