six stroke engine seminar

7/30/2019 Six Stroke Engine Seminar

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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)


(Second Combustion Stroke)

M f2 [mg/cycle]

5.0

(Diesel Fuel)

10.7

(Methanol)


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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(First Combustion Stroke)

M f1 [mg/cycle]

8.0

(Diesel Fuel)


(Second Combustion Stroke)

M f2 [mg/cycle]

8.0

(Diesel Fuel)

17.2

(Methanol)


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

six stroke engine seminar

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