1

Experimental study of a kerosene fuelled internal combustion engine

Tomás Formosinho Sanchez

Instituto Superior Técnico, Technical University of Lisbon, Av. Rovisco Pais, 1049-001

Lisboa, Portugal; Email: tomasformozinho@gmail.com

Abstract

Mistral Engines, a Swiss Aeronautical Company wishes to be able to have operating Wankel

engines on kerosene fuel because of its market availability.

Kerosene is a fuel with lower cetane number than diesel fuel, thus it should give a longer ignition

delay. This makes it viable for lower emissions since the longer ignition delay means longer time for

the fuel to mix with the in-cylinder gas prior to combustion onset.

With the cooperation of the Industrial Energy Systems Laboratory of the Swiss Federal Institute of

Technology in Lausanne, a study on the combustion of kerosene in an internal combustion engine

(ICE) is here presented.

This study is performed in a mono cylinder ICE with variable Compression Ratio (CR) and

Variable Valve Timing (VVT).

Different strategies, such as Homogeneous Charge Compression Ignition (HCCI), Homogeneous

Charge Spark Ignition (HCSI), Controlled Auto Ignition (CAI) and a combustion pre-chamber concept

are considered.

1. Introduction

The transportation sector, including aviation, an essential part of our modern society, represents

the largest part of the petroleum based fuels consumption. Its importance has been growing

continuously during the last decades. The aviation sector normally uses kerosene based fuels in jet

engines. The application of this type of fuel in internal combustion engines (Wankel engine and

reciprocating piston engines) is a big challenge as a result of their lower fuel tolerance than jet

engines.

Nevertheless, Mistral Company wants to adapt its Wankel engines to that fuel since the market

availability is bigger than the used 100LL avgas gasoline. Therefore, a study on the kerosene

combustion in a single cylinder engine with variable CR and VVT is undertaken. The engine is

naturally aspirated and the kerosene is direct injected into the combustion chamber.

To understand the behaviour of kerosene combustion in an ICE several parameters were

investigated. The influence of the fuel temperature, the timing of the pre-injection and the richness of

the mixture. Furthermore, SI and CI were investigated and HCCI was attempted with a new concept of

combustion pre-chamber.

Controlled Auto Ignition (CAI) and HCCI combustion are radically different from the conventional

SI combustion in a gasoline engine and CI diffusion combustion in a diesel engine. The combination of

diluted and premixed fuel and air mixture with multiple ignition sites throughout the combustion

chamber eliminates the high combustion temperature zones and prevents the production of soot

particles, hence producing ultra low NOX and particulate emissions. The use of lean, or more often

diluted, air-fuel mixture with recycled burned gases permits unthrottled operation of a CAI/HCCI

2

gasoline engine, thus yielding higher engine efficiency and better fuel economy than SI combustion.

Therefore, CAI/HCCI combustion represents for the first time a combustion technology that can

simultaneously reduce both and NOX particulate emissions from a diesel engine and has the capability

of achieving simultaneous reduction in fuel consumption and NOX emissions from a gasoline engine.

Based on these promises, an alternative solution is intended to auto ignite the mixture in several

points as in HCCI combustion. An auto ignition pre-chamber jet ignition is here studied and presented.

A glow plug inside the pre-chamber should ignite virtually simultaneously an important air-fuel mixture

that will form jets flowing out of the nozzles igniting, in several sites, the remaining of the air-fuel in the

main chamber. The jet ignition will develop significantly faster than the spark ignition type flame front

propagation.

2. Combustion pre-chamber

Occasionally, in CI engines, fuel is injected into a pre-chamber where a Glow Plug helps to ignite

the mixture on cold starts.

However, a new concept was developed where fuel is injected into the main combustion chamber.

A pre-chamber is located where the spark plug was. The air-fuel mixture is forced into the pre-

chamber through six nozzles at high velocity during the compression stroke. Fuel is injected into a

very turbulent flow which leads to desired high mixing rates.

The rich mixture with very high swirl in the pre-chamber will ignite readily and combust very

quickly, finally igniting the main chamber gases.[1] Flame jets are intended to burn quickly the whole

mixture in a sort of HCCI combustion.

a) Cylinder head and pre-chamber b) Combustion pre-chamber concept

Figure 1 Cylinder head and combustion pre-chamber images

Figure 1 a) shows the images of the pre-chamber designed for this engine. The Bosch glow plug

(yellow) selected was the GPM902 from a Bosch catalogue.

The pre-chamber concept is shown in Figure 1 b). The nozzle orifices can be observed as well.

Given that the glow plug reaches 1000ºC of temperature, the pre-chamber was made of heat resistant

austenitic stainless steel X15CrNiSi25-20 (Böhler H525).

Combustion

pre-chamber

Bosch glow

plug

3

Consequently, the erosion process of the nozzle orifices is also slowed. The real pictures of the

pre-chamber are shown in Figure 2.

a) Bosch glow plug and pre-chamber b) side view of pre-chamber

Figure 2 Real photos of pre-chamber and glow plug

3. Lean combustion mode and HCCI

The use HCCI combustion in internal combustion engines is of interest because it has the

potential to produce low NOx and Particle Mater (PM) emissions while providing diesel-like efficiency.

In HCCI combustion, a premixed charge of fuel and air auto-ignites at multiple points in the cylinder

near TDC, resulting in rapid combustion with very little flame propagation. In order to prevent

excessive knocking during HCCI combustion, it must take place in a dilute environment, resulting from

either operating fuel lean or providing high levels of either internal or external Exhaust Gas

Recirculation (EGR). Operating the engine in a dilute environment can substantially reduce the

pumping losses, thus providing the main efficiency advantage compared to SI engines.[2]

Lean combustion burning is a solution that allows relatively high compression ratios combined with

significantly NOX emissions without after-treatment.

HCCI is a form of internal combustion in which well mixed fuel and oxidizer are compressed to the

point of auto-ignition. It combines characteristics from SI engines (HCSI: homogeneous charge spark

ignition) and CI (SCCI: stratified charge compression ignition) engines. The pressure and temperature

of the mixture are raised by compression until the entire mixture reaches the point of auto ignition.[1]

4. Results and Discussion

4.1. Pre-Injection timing variation

Several measures were taken in order to study the effects of the position/timing variation of the

pre-injection on the combustion process in the engine. This test consists of varying the timing of the

pre-injection and keeping all the other parameters constant. Therefore, earlier injections should give a

longer mixing period of time conducting to more homogeneous mixtures. If properly chosen, sufficient

time is available between the end of injection and the start of ignition, assuring a relatively good

homogeneous mixture, which will result in fully premixed combustion. The use of direct injection

compared to port injection should be an advantage since the fuel is injected during the compression

stroke, the gas temperature and density are higher than at intake conditions, enhancing the

vaporization process and thus reducing the time to prepare the mixture and/or avoiding the need to

heat up the intake air.

4

0

10

20

30

40

50

60

-10 -5 0 5 10 15 20

Crank Angle

Pressure[bar]

θ=-36ºθ=-63ºθ=-90ºθ=-117ºθ=-144º

-50

0

50

100

150

200

250

300

350

-10 -5 0 5 10 15 20

Crank Angle

Instantaneous Heat Release [J]

θ=-36ºθ=-63ºθ=-90ºθ=-117ºθ=-144º

The test was performed at 1500rpm and the duration of injection was 0.4ms for the pre-injection

and 0.85ms for the post-injection. Thus the fuel quantity injected was 7.66x10-5 kg per cycle. Although

the fuel injected per cycle was constant, the lambda sensor revealed values for λ within [0.96; 1.03].

The pre-injection timings were the following: -36º, -63º,-90º,-117º, and -144ºCA BTDC. The post-

injection timing was kept at a constant -23ºCA BTDC. There was no spark ignition. Therefore the

ignition mode was auto-ignition.

As a result, homogeneous compression ignition combustion should be achieved without the

adaptation of fuel supply in the intake port. Direct injection was made at 950 bars of pressure with a

CR of 13.

a) Pressure vs crank angle b) IHR vs crank angle

Figure 3 Pressure and IHR as function of crank angle for pre-injection timing variation

Figure 3 a) shows the evolution of the pressure as function of crank angle. There is a clear

tendency to shift the pressure curve at the vicinity of TDC as function of the timing of injection. The

later the pre-injection, the closer the curve is to TDC. The same can be observed in Figure 3 b), the

Instantaneous Heat Release (IHR) shifts the same way as the pressure curve. Furthermore, it is

noticeable that too late pre-injections have oscillating behaviour at the end of High Temperature Heat

Release (HTHR). At pre-injection timing of º36−=θ big oscillations of IHR are visible, especially

between 3 to 5ºCA (at the end of HTHR). These oscillations can be a sign of fuel stratification, small

clouds of rich mixture that could auto-ignite locally or knock (as some authors suggest [3]) and

consequently some peaks on the IHR are found. This is visible when Figure 3 a) and Figure 4 b) are

observed. They both show high oscillations between 3 to 5ºCA with a pre-injection timing of º36−=θ .

As a matter of fact, during the experiments, the later the pre-injection, the heavier the sound of the

engine was perceptible.

Concerning the Cumulative Heat Release (CHR), Figure 4 a) shows the specific performance of

each of the tests carried out. It can be observed that too late pre-injection ( º36−=θ ) noticeably

affects the CHR, shifting the curve to the vicinity of the TDC and delaying the combustion process.

There seems to be an ideal pre-injection timing at º90−=θ where the cumulative heat release is

delivered more rapidly than the other pre-injection timings.

5

0

10

20

30

40

50

60

70

80

90

100

-20 -10 0 10 20 30 40 50 60

Crank Angle

Cumulative Heat Release [%]

θ=-36ºθ=-63ºθ=-90ºθ=-117ºθ=-144º

0

0.5

1

1.5

2

2.5

3

3.5

-160 -140 -120 -100 -80 -60 -40 -20 0

Pre-Injection [degrees before TDC]

Q50%[CAD]

40

42

44

46

48

50

52

54

56

58

60

-160 -140 -120 -100 -80 -60 -40 -20 0

Pre-Injection [degrees before TDC]

Combustion Duration [CAD]

0

200

400

600

800

1000

1200

1400

1600

1800

-20 -10 0 10 20 30 40 50 60

Crank Angle

Cylinder Gas Temperature [K]

θ=-36ºθ=-63ºθ=-90ºθ=-117ºθ=-144º

a) CHR vs crank angle b) Cylinder gas temperature vs crank angle

Figure 4 CHR and Cylinder gas temperature as function of crank angle for pre-injection timing variation

Figure 4 b) shows the cylinder gas temperature profile of each pre-injection timing strategy. There

is a clear tendency to shift up the gas temperature as the pre-injection delay increases. Thus, pre-

injection strategies closer to TDC appear to have gas temperatures higher than earlier pre-injection.

Consequently, later pre-injections have hotter exhaust temperatures as well. It is also possible to

observe temperature oscillations in the later pre-injection. As said before, this can occur due to fuel

stratification that can locally ignite after the main combustion.

a) Q50% vs pre-injection timing b) Combustion duration vs pre-injection timing

Figure 5 Q50% and Combustion duration as function of crank angle for pre-injection timing variation

In Figure 5 a) the curve shifting observed in Figure 3 is emphasised. In fact, it shows the position

of 50% of the CHR as function of the various pre-injection strategies. Again, a tendency can be

observed. As the pre-injection comes later, the 50% of CHR approximates the TDC.

Figure 5 b) shows the combustion duration in CAD as function of the various pre-injection

strategies. It can be noticed that the combustion duration increases with the increase of the pre-

injection delay. Moreover, an optimal strategy seems to occur, where the combustion duration is the

fastest. That strategy is, when the pre-injection is done at º90−=θ .

4.2. Excess of air ratio λ variation with pre-chamber, SI and CI comparison

The test was performed at 1500rpm and the timing of the injections was -126ºCA for the pre-

injection and -26ºCA for the post-injection. As a result, a very homogeneous mixture is intended to be

achieved. There was no spark ignition; therefore, there is no direct control on start of combustion. The

Bosch Glow Plug shown in Figure 2 a) was constantly on. Hence, HCCI should be achieved with the

help of jet flames that leave the pre-chamber. The duration of injection was 0.4ms for the pre-injection

6

0

10

20

30

40

50

-60 -40 -20 0 20 40 60

Crank Angle

Pressure[bar]

λ=0.65λ=0.70λ=0.85λ=1.00λ=1.15λ=1.30

0

10

20

30

40

50

-60 -40 -20 0 20 40 60

Crank Angle

Pressure[bar]

λ=0.70λ=0.85λ=1.00λ=1.15λ=1.30

-50

0

50

100

150

200

250

300

-5 0 5 10 15 20 25

Crank Angle

Instantaneous Heat Release [J]

λ=0.65λ=0.70λ=0.85λ=1.00λ=1.15λ=1.30

0

10

20

30

40

50

60

70

80

90

100

-20 -10 0 10 20 30 40 50 60

Crank Angle

Cumulative Heat Release [%]

λ=0.65λ=0.70λ=0.85λ=1.00λ=1.15λ=1.30

and [1.9; 1.63; 1.06; 0.71; 0.58; 0.52; 0.48] ms for the post-injection. Thus the fuel quantity injected

was [14.09, 12.44; 8.95; 6.80; 6.01; 5.64; 5.39] x10-5kg per cycle. The lambda sensor revealed the

following values: =λ [0.65; 0.7; 0.85; 1.00; 1.15; 1.30; 1.45]. Direct injection was made at 950 bars of

pressure with a CR of 13.

a) Pressure vs crank angle (Pre-Chamber) b) Pressure vs crank angle (SI)

Figure 6 Pressure (Pre-chamber and SI) as function of crank angle for various λ

Figure 6 compares the cylinder pressure for both ignition methods. As mentioned, the pre-

chamber tremendously affects the start of combustion. There is a clear propensity that the pre-

chamber delays the start of combustion (see also Figure 8 a)). The maximum pressure peak of the

pre-chamber always appears for higher crank angle degrees than in SI tests. It is also noticeable that

the peak of pressure in the pre-chamber tests is always lower than the corresponding SI ones. Since

the start of combustion is delayed, the piston is already in the expansion stroke on its way to BDC;

therefore lower peaks of pressure during combustion are expected. If the pressure curves are shifted

to the right and have lower peaks of pressure in the pre-chamber tests, thus all the other curves

(cylinder gas temperature, IHR and CHR) should behave similarly as well. In fact delaying the start of

combustion has an important effect on the IHR curve as well; this can be shown in Figure 7.

a) IHR vs crank angle (Pre-chamber) b) CHR vs crank angle (Pre-chamber)

Figure 7 IHR and CHR (Pre-chamber) as function of crank angle for various λ

Figure 8 compares 5% of CHR (Q5%) and the combustion duration between the pre-chamber, SI

and CI tests. As said before, Figure 8 a) shows that there is a clear propensity of delaying the start of

combustion when the pre-chamber is used.

7

-4

-2

0

2

4

6

8

10

0.60 0.70 0.80 0.90 1.00 1.10 1.20 1.30 1.40 1.50

Excess of air ratio λ

Q5%[CAD]

CI with Pre-Chamber

SI

CI

0

10

20

30

40

50

60

0.60 0.70 0.80 0.90 1.00 1.10 1.20 1.30 1.40 1.50

Excess of air ratio λ

Combustion Duration [CAD]

CI with Pre-Chamber

SI

CI

0.470

0.480

0.490

0.500

0.510

0.520

0.530

0.540

0.60 0.70 0.80 0.90 1.00 1.10 1.20 1.30 1.40 1.50

Excess of air ratio λ

Therm

odynamic Efficiency [%]

CI with Pre-Chamber

SI

CI

0.900

0.920

0.940

0.960

0.980

1.000

1.020

0.60 0.70 0.80 0.90 1.00 1.10 1.20 1.30 1.40 1.50

Excess of air ratio λ

Combustion Efficiency/Combustion

Efficiency(Pre-chamber) CI with Pre-Chamber

SI

CI

In Figure 8 b) it is clear that the pre-chamber severely affects the combustion duration. When used

it can burn the fuel faster then flame propagation on SI mode (10ºCA faster, maximum). Thus, the

flame jets flowing out of the pre-chamber should burn better and quicker than the mixture in the

cylinder. If the combustion duration is reduced, its efficiency should be increased as predicted in HCCI

theory.

a) Q5% vs λ b) Combustion duration vs λ

Figure 8 Q5% and Combustion duration (Pre-chamber, SI and CI) as function of λ

Figure 9 a) proves that combustion efficiency is in fact higher when the pre-chamber is used.

Figure 9 Combustion and Thermodynamic efficiencies (Pre-chamber, SI and CI) as function of λ

Although the combustion efficiency is higher to the pre-chamber tests, the thermodynamic

efficiency (see Figure 9 b)) presents the inverse results. This is due to the definition of the

thermodynamic efficiency. In fact, the IMEPgross is severely affected by the big delay of start of

combustion.

In Figure 10 it is possible to observe that the CHR (QMEP) is higher, for leaner and richer mixture,

in the pre-chamber tests than in SI mode. Comparing with CI mode, pre-chamber present always

higher QMEP values. This also helps to conclude that with pre-chamber the fuel burns better,

delivering more heat, thus more work is produced and consequently more power.

a) Combustion Efficiency vs λ b) Thermodynamic Efficiency vs λ

8

6

7

8

9

10

11

12

0.60 0.70 0.80 0.90 1.00 1.10 1.20 1.30 1.40 1.50

Excess of air ratio λ

QMEP [bar]

CI with Pre-Chamber

SI

CI

Figure 10 QMEP (Pre-chamber, SI and CI) as function of λ

The combustion and thermodynamics efficiencies should give an idea how kerosene behaves

inside the combustion chamber in different conditions. Some conclusions are now outlined.

5. Conclusions

To understand the behaviour of kerosene combustion in an ICE several parameters were

investigated: influence of fuel temperature, timing of the pre-injection and richness of the mixture.

Furthermore, SI and CI were investigated and HCCI was attempted with a new concept of combustion

pre-chamber. The experiments allow to drawn the following conclusions.

1. Start of combustion is sensitive to excess of air ratio in SI and CI mode, presenting a large delay

when the mixture is very rich (λ<0.70).

2. Engine start is difficult on CI mode when engine is cold.

3. Better efficiencies are obtained for leaner mixtures, presenting a maximum when λ=1.15.

4. Too late pre-injections lead to IHR instability at the end of HTHR. Fuel stratification may take place

and auto-ignition of small clouds of rich mixture happens.

5. Curves shifting are similar with variation of λ for both SI and pre-chamber mode.

6. The pre-chamber leads to larger ignition delay. In fact, the ignition is not controlled as in SI mode.

7. The use of pre-chamber leads to overall better efficiencies (combustion and indicated), thus better

engine performance.

8. The use pre-chamber and HCCI approach leads to improved combustion stability.

9. The long ignition delay, which can partly lead to low soot emissions, could be beneficial for

expanding the HCCI operation area. In combination with the market availability of kerosene, it has

the possibility to be a suitable HCCI fuel.

6. Perspectives

In this preliminary study, the pre-chamber results seem to be encouraging since overall better

efficiencies were achieved when compared with SI mode. However, the ignition delay seems to be

very high. For this reason, even better results are expected if the ignition delay could be properly

controlled.

9

V

W

V

dVp

V

dVpIMEP i

net =⋅

=⋅

=∫∫ −

360

360

V

dVpIMEPgross

∫− ⋅=

180

180

V

LHVmFuelMEP

f ⋅=

V

QQMEP =

QMEP

IMEPgrossth =η

FuelMEP

QMEPcomb =η

A comparison between CI and CI with pre-chamber should complete the final conclusions. It

should also help to conclude if there is a positive influence in the use of the combustion pre-chamber

when compared with normal CI and SI mode. Therefore, further tests should be carried out with and

without pre-chamber using CI mode.

EGR can be used to increase the mixture dilution and augment its temperature so that the ignition

delay can be reduced. Thus, a Controlled Auto Ignition (CAI) of the HCCI mixture should be achieved.

Air assist DI combustion systems can be used for better fuel vaporization. This methodology can

also reduce the poor start problems once it enhances homogenization.

Pollutants measurement should be executed once the aviation market will have its own restrictions

as in automobile market.

As the present work intends to study the kerosene combustion behaviour inside of an ICE for

aviation applications, the air admission pressure and its temperature should be fitted to the altitude

conditions.

Since Mistral Company intends to use Wankel engines, the results presented in this paper should not

be conclusive. Tests on a Wankel engine must be performed so that the study can be validated on the

application field.

Formulae Instantaneous heat release (IHR) Indicated Mean Effective Pressure (IMEP) Fuel Mean Effective Pressure (FuelMEP) The fuel mean effective pressure is defined as:

where fm [kg] is the mass of fuel supplied per

cycle, LHV [J/kg] is the lower heating value

for the fuel and V [m3] is the displacement of the engine. Heat Release Mean Effective Pressure (QMEP) Thermodynamic Efficiency Combustion Efficiency

Symbols and acronyms

ABDC After bottom dead centre

ATDC After top dead centre

BBDC Before bottom dead centre

BDC Bottom dead centre

BMEP Break mean effective pressure [bar]

BTDC Before top dead centre CA or CAD Craft shaft angle [º] CAI Controlled auto igniton CI Compression ignition CR Compression ratio DI Direct Injection ECU Engine unit control

EGR Exhaust gas recirculation

EVC Exhaust valve closure EVO Exhaust valve opening

FuelMEP Fuel mean effective pressure [bar]

HCCI Homogeneous Charge Compression Ignition

HTHR High temperature heat release

IHR Instantaneous heat release [J]

IMEP Indicated mean effective pressure [bar]

IVC Intake valve closure

⋅+⋅⋅−

=θ

θθ

θγγθ σσ d

dpV

d

dVp

d

dQ)()(

1

1

10

IVO Intake valve opening LHV Low heat value [MJ/kg]

N Engine speed [RPM, RPS, Hz]

NOX Oxides of nitrogen P Cylinder pressure [bar]

Q Cylinder energy release [J,kJ]

QMEP Heat release mean effective pressure

RPM Rotation per minute [rpm]

SI Spark ignition TDC Top dead centre

UAV Unmanned Aeronautical Vehicles

V Total cylinder volume [cm3]

VVT Variable valve timing

γ

Ratio between specific heats, adiabatic polytrophic index

σγ Evaluated polytrophic index

combη Combustion efficiency

gasη Gas exchange efficiency

thη Thermodynamic efficiency

θ Crank shaft angle [º]

λ Air-to-fuel ratio or excess of air ratio

References

[1] MEIER, M : Etude expérimentale de l’auto-inflammation d’une préchambre de moteur monocylindre, EPFL Master Thesis 2006.

[2] SZYBIST, J, & BUNTING, B: Chemistry Impacts in Gasoline HCCI, Oak Ridge National Laboratory, 2006.

[3] ZHAO, H: HCCI and CAI engines for the automotive industry, Woodhead Publishing Limited, Cambridge, 2007.

[4] TÄSCHLER, C: Caractérisation expérimentale de l’inflammation d’un “spray” de kérosène, dans les conditions d’un moteur, EPFL Master Thesis 2008.