surrogate fuels for premixed combustion in compression ignition engines

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2011 12: 452 originally published online 25 July 2011International Journal of Engine ResearchG T Kalghatgi, L Hildingsson, A J Harrison and B Johansson

Surrogate fuels for premixed combustion in compression ignition engines

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Surrogate fuels for premixed combustion incompression ignition enginesG T Kalghatgi1,*y, L Hildingsson1,2, A J Harrison1, and B Johansson2

1Shell Global Solutions (UK), Chester, UK2Division of Combustion Engines, Lund University, Sweden

The manuscript was received on 20 September 2010 and was accepted after revision for publication on 13 April 2011.

DOI: 10.1177/1468087411409307

Abstract: Simple surrogate fuels are needed to model practical fuels, which are complex mix-tures of hydrocarbons. The surrogate fuel should match the combustion and emissions beha-viour of the target fuel as much as possible. This paper presents experimental results using awide range of fuels in both the gasoline and diesel auto-ignition range, but of different volatili-ties and compositions, in a single cylinder diesel engine. Premixed combustion in a compres-sion ignition engine is defined, in this paper, to occur when the injection event is clearlyseparated from the combustion and the engine-out smoke is very low – below 0.05 FSN (filtersmoke number). Under such circumstances, if the combustion phasing is matched for twofuels at a given operating condition and injection timing, the emissions are also comparableregardless of the differences in composition and volatility. For the experimental conditionsconsidered, combustion phasing at a given operating condition and injection timing dependsonly on the octane index (OI), OI = (1-K)RON + KMON, where RON and MON are researchand motor octane numbers and K is an empirical constant that depends on operating condi-tions. A mixture of iso-octane, n-heptane and toluene can be found to match the RON andMON of any practical gasoline and will be a very good surrogate for the gasoline since itwill have the same OI. If the compression ratio is greater than 14, practical diesel fuels, withDCN (derived cetane number) between 40 and 60, will have comparable ignition delays ton-heptane, which is an adequate surrogate for such fuels. However, premixed combustion canbe attained only at much lower loads at a given speed with diesel fuels compared to gasolines.

Keywords: compression ignition, diesel engine, premixed combustion, low-NOx low-smoke,

octane index, surrogate fuels

1 INTRODUCTION

Practical compression ignition engines, or diesel

engines, use diesel fuels that are very prone to auto-

ignition. The diesel fuel auto-ignition quality is usu-

ally characterised by the derived cetane number

(DCN) based on ignition delay measurements in the

ignition quality test (IQT). The higher the DCN, the

lower the ignition delay. In contrast, gasoline fuels,

used in spark ignition (SI) engines, are very resistant

to auto-ignition so as to avoid knock, an abnormal

combustion phenomenon. The gasoline auto-ignition

quality is usually specified by the research and motor

octane numbers (RON and MON) and there is an

inverse correlation between the octane numbers and

DCN if these can be measured for the same fuels [1].

However this is not usually possible because practi-

cal diesel fuels are much heavier and less volatile

than practical gasolines and cannot be run in RON

and MON tests. Fuels with RON . 60 can be classi-

fied as gasoline-like fuels [1].

Practical diesel fuels have DCNs ranging between

40 and 60 and ignite very soon after the start of

injection, at or near the end, top dead centre (TDC)

of the compression stroke, before the fuel has had a

chance to mix properly with the oxygen in the cylin-

der. This causes combustion to occur in mixture

packets which are fuel-rich and leads to high smoke

*Corresponding author: ynow at Saudi Aramco. Saudi Aramco,

PO Box 9290, Dhahran 31311, Saudi Arabia.

email: [email protected]

452

Int. J. Engine Res. Vol. 12

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(particulates) and nitrogen oxide (NOx) formation.

Requirements for the control of NOx and smoke are

becoming increasingly stringent. Smoke formation

can be minimized by ensuring that the equivalence

ratio of the mixture packets where combustion

occurs is not greater than u .~2 or l\~0.5 [2] where

u is the equivalence ratio and l = (1/u) is the normal-

ized air/fuel ratio. Even if smoke is formed, if there is

sufficient oxygen and the temperature is high

enough, it will be oxidized inside the cylinder and

engine-out levels will be low. NOx formation can be

minimized if the combustion temperatures are kept

below about 2200 K [2]. This can be achieved by

either running the engine lean, with l much greater

than 1, or by using high levels of exhaust gas recircu-

lation (EGR). Indeed, much of the advanced technol-

ogy used in modern diesel engines, which makes

them expensive and complicated, is aimed at pro-

moting premixed and low-temperature combustion

by overcoming the low ignition delay of diesel fuels.

For instance, very high injection pressures are used

to increase the mixing rate in advanced diesel

engines using diesel fuel.

Kalghatgi and co-workers [3, 4] demonstrated in a

2 L single cylinder engine that, if fuels with high resis-

tance to auto-ignition, such as gasoline, are used in

diesel engines, compared to diesel fuel, auto-ignition

occurs significantly later after the start of injection at

a given operating condition. At each operating condi-

tion, there is a range where the combustion phasing

can be controlled by varying injection timing. If the

same amount of gasoline is injected early at the same

conditions i.e. with fully premixed conditions as in

homogeneous charge compression ignition (HCCI),

ignition might not occur at all. Thus the inhomogene-

ity is essential for combustion to occur but the high

ignition delay makes combustion happen when fuel

and air are better mixed – fuel and air are ‘premixed

enough’ but must not be fully premixed. The gasoline

fuel has to be injected significantly earlier compared

to the diesel fuel to get the same combustion phasing.

With high ignition delays, the mixture packets where

combustion takes place will be nearer the global

equivalence ratio. Hence, at low loads, when the glo-

bal mixture is lean, very low levels of NOx, smoke,

and pressure rise rates result when gasoline fuel is

used but at the cost of high CO and hydrocarbons

(HC) [3]. Even if smoke is formed with diesel fuel in

the fuel-rich zones, it is oxidized because of the

excess oxygen available and engine-out smoke will be

low for all fuels at low loads. However, combustion is

mixing-controlled with diesel fuel at high loads, since

fuel injection cannot be completed before combus-

tion starts, which is very soon after the start of the

fuel injection. In contrast, with gasoline, combustion

occurs after fuel injection is completed even at high

loads so that the probability of smoke formation is

significantly reduced, even at high loads. When high

levels of EGR are used to control NOx, it does not

matter that the oxygen level in the cylinder is reduced

since not much smoke is formed in the first place and

engine-out smoke can remain very low.

Similar studies have been conducted in a smaller

single cylinder engine of 0.537 L displacement and at

engine speeds up to 3000 r/min [5–9]. Groups from

Lund [10, 11, 12], Wisconsin [13, 14], and Cambridge

[15, 16] universities have also demonstrated the bene-

fits of running diesel engines on gasoline-like fuels. In

summary, NOx and smoke can be controlled simulta-

neously even at higher loads, compared to diesel fuels,

if a diesel engine is run on gasoline, because premixed

combustion is facilitated by the high ignition delay.

This can be achieved at extremely high fuel efficiency

[10–16]; indicated thermal efficiencies of over 50 per

cent have been reported [e.g. 11]. At low loads, signifi-

cantly lower pressure rise rates and NOx can be

obtained with gasoline [e.g. 5, 6, 7, 9]. It could be that

high injection pressures are not needed and (NOx)

after-treatment could be replaced by an oxidation cat-

alyst to control CO and HC if gasoline is used in diesel

engines, so that there is scope for reducing the cost

and complexity of diesel engines. Moreover, the

octane number and the volatility of the gasoline for

such combustion systems could be much lower com-

pared to current market gasolines [6, 7, 9]. This might

lead to significant savings in energy and CO2 in fuel

manufacture. Thus there is great incentive to develop

combustion systems using gasoline-like fuels in

advanced diesel engines.

Good computational models would be of great help

in developing such practical combustion systems.

Such models need to incorporate reliable chemical

kinetic models to predict auto-ignition. However,

practical fuels are complex mixtures of hydrocarbons

and the development of models that represent all

these components would be prohibitively complex. In

any case, all the fundamental data needed for devel-

opment of such a model, such as chemical kinetic rate

constants, are not available. Thus, simplified ‘surro-

gate fuels’ are needed for representing practical fuels

[17, 18, 19]. A surrogate fuel is defined as a fuel com-

posed of a small number of pure compounds whose

behaviour matches the target practical fuel in terms of

combustion and emissions characteristics.

We define premixed combustion to occur in die-

sel engines if the fuel injection is completed before

combustion starts. This does not mean that the fuel

and oxygen are fully premixed at the start of com-

bustion as in HCCI engines, where combustion

phasing is determined by the conditions at the start

Surrogate fuels for premixed combustion in compression ignition engines 453





of the compression and the auto-ignition chemistry

of the fuel mixture. The crucial difference from

HCCI is that combustion phasing can be controlled

by injection timing. This requires that the equiva-

lence ratio is not the same everywhere in the com-

bustion chamber as in HCCI and such conditions

can be attained by fuel injection significantly later

in the cycle, compared to HCCI. Premixed combus-

tion is not possible with conventional diesel fuels at

high loads and the classical mixing-controlled com-

bustion [2] takes place.

Both auto-ignition and the resultant emissions are

greatly influenced by the mixing of the fuel with the

gases in the engine, which in turn is determined by

the injection and vaporization processes. It is desir-

able to match the physical properties such as volati-

lity, density, viscosity, surface tension, and diffusion

coefficients of the surrogate and target fuels in order

to match the mixing process and also the molecular

structure and sooting propensity [19]. Clearly, it is

impossible to match all the properties between a

complex practical fuel and a simple surrogate fuel.

However, experimental evidence shows that, for pre-

mixed compression ignition, where heat release starts

after fuel injection is completed, the volatility and

composition are far less important than the auto-igni-

tion behaviour of fuels. If, for a given condition, using

a single injection, the combustion phasing for a given

injection timing is matched for two fuels, the emis-

sions are also very similar for the two fuels regardless

of the differences in volatility and composition [7, 9].

It was also demonstrated in [8] that combustion phas-

ing at a given injection timing depended on the

octane index (OI) of a fuel which was a function of its

RON and MON at a given operating condition. In [8]

the conditions considered were all without EGR. In

this paper we extend such results to cases with EGR,

bring together data from different earlier publications,

and discuss surrogate fuels that could be used to

model the combustion and emission behaviour of

complex practical fuels in premixed compression igni-

tion combustion.

2 EXPERIMENTAL DETAIL

The experiments were performed on a 4-valve single

cylinder research engine with dimensions as pre-

sented in Table 1. All experiments were done with

coolant and oil temperatures at 90 �C and the inlet

air temperature was kept between 55 �C and 60 �C

using a heater. Fuel was injected via a Bosch 7 hole

injector, with injector cone angle of 153 � and hole

diameters of 0.13 mm, fed by an independent fuel

supply rig. An external air compressor was used

to simulate boosted conditions. When EGR was

introduced, the exhaust backpressure was set 0.2

bar higher than the inlet manifold air pressure and

the recirculated gases were cooled using an external

cooling circuit to the same temperature as the inlet

air, i.e. 60 �C. Any water that condensed out was not

drained and was mixed with the intake air and was

expected to vaporize as the mixture passed through

the air heater, which has a labyrinth design and high

wall temperatures. In-cylinder pressure was mea-

sured with a water-cooled pressure transducer

(Kistler 6041A). Emissions and inlet FSN level were

measured using a Horiba MEXA-9500H system and

soot was measured using an AVL 415 smoke meter.

After a stabilization period, the emissions were

logged once per second for 60 s and the averages

of those 60 recordings are presented in this paper.

At the same time, the in-cylinder pressure was

recorded for 100 cycles.

The results from nine different fuels are considered

in this paper. The properties and compositions of

these fuels are listed in Table 2. Four of these fuels,

PRF 84, TRF 82, TRF 84 and n-Heptane, are model

fuels made up of mixtures of ASTM grade iso-octane,

n-heptane and toluene. Four, ULG 73, ULG 78, ULG

84 and ULG 91, are full boiling range gasolines of dif-

ferent octane numbers. The ninth fuel, D1, is a com-

mercial European low sulphur diesel fuel with a DCN

of 56. The volatility characteristics of these fuels are

shown in Fig. 1 where the volume per cent recovered

at a given temperature in the ASTM volatility test is

plotted against the temperature. N-heptane, iso-octane

and toluene have boiling points of 98 �C, 99.2 �C, and

110.6 �C respectively so that the boiling range of the

four model fuels will be between 98 �C and 110.6 �C.

All the full boiling range fuels have sulphur levels lower

than 10 ppm and aromatic content varying between

19 per cent and 30 per cent by volume. All the fuels had

a sufficient amount of lubricity additive (300 ppm of

Paradyne R655 from Infineum) to ensure that the

lubricity scar size was well within the European speci-

fication. It can also be seen from Table 2 that all the

fuels have similar gravimetric heat of combustion.

Three experimental conditions were used and

these are summarized in Table 3. The operating

Table 1 Engine dimensions

Compression ratio 16:1Displacement 0.537 lBore 88 mmStroke 88.3 mmConnection rod length 149 mmIVO 362 CADIVC 595 CADEVO 143 CADEVC 385 CAD

454 G T Kalghatgi, L Hildingsson, A J Harrison, and B Johansson





conditions, such as injection pressures or intake

temperatures, and loads chosen have no special sig-

nificance other than that they are well within nor-

mal diesel operating conditions.

3 RESULTS AND DISCUSSION

All the experiments used a single injection pulse. In

the discussion below, all crank angle degrees (CADs)

are expressed in relation to the TDC of the compres-

sion stroke, which is zero; the TDC on the exhaust

stroke is 360 CAD. Pressure signals are averaged over

100 cycles and heat release rates are calculated from

pressure signals and averaged over 100 cycles.

Figure 2(a) shows the average pressure for three fuels

with the crank angle position of the electrical signal

marking the start of injection (SOI) at 28.0 CAD for

Condition 1. Figure 2(b) shows the heat release rate

(HRR), ignoring heat losses, calculated from the pres-

sure curves and the integrated heat release normal-

ized with respect to the maximum heat release

normalized heart rate (NHR) for each fuel. The injec-

tion duration is also marked on Figs 2a and 2b.

Combustion phasing parameters such as CA50, the

crank angle degree at which 50 per cent of total heat

release has taken place, and CA2, the crank angle

degree at which 2 per cent of total heat release has

taken place, are calculated from the NHR.

We now define different time constants that can be

used to characterize combustion phasing and the

nature of combustion. The combustion delay (CD) is

the time from the start of injection to the 50 per cent

burn time, CD = CA50 2 SOI. It is reasonable to assume

that a larger CD would mean that fuel and air would

be better mixed so that the local equivalence ratio

where combustion actually occurs would be nearer the

global equivalence ratio. Besides CD, the other time

scales of interest are the ignition delay, ID = CA2 2 SOI

and the ignition dwell time, IDW = CA2 2 EOI where

EOI is the crank angle position of the electrical signal

marking the end of injection. We use CA2 rather than

the start of combustion, which occurs earlier, because

CA2 can be estimated more reliably. When the injec-

tion event is not completed before the start of combus-

tion, as in conventional mixing-controlled diesel

combustion, IDW is negative. From Fig. 2(b) it can be

seen that, for Condition 1, even for n-heptane, com-

bustion starts after the end of injection, IDW is positive,

though small.

Table 2 Properties of fuels considered

Fuel Isooct n-hep ToluAromatics Density

C H

LHVRON MON vol% vol% vol% vol% g/cc MJ/kg

PRF 84 84 84 84 16 0.0 0.682 7.83 17.67 44.4TRF 84 84.5 74.5 0 35 65 65.0 0.785 7.00 10.12 41.7TRF 82 82.1 78.1 50 24 26 26.0 0.723 7.43 14.07 43.2n-Hept 0 0 0 100 0 0.0 0.632 7.00 16.00 44.6ULG 73 72.9 68.4 19.0 0.715 6.54 13.06 43.6ULG 78 78.5 73 23.0 0.726 6.61 12.79 43.4ULG 84 84.1 78 26.5 0.736 6.68 12.52 43.2ULG 91 90.7 81.8 29.8 0.731 6.92 12.32 43.2D1* 25.2 0.833 42.9

*DCN = 56. If RON and DCN can be measured for the same fuel, from Equation 6 in [1], DCN = 54.6 2 0.42*RON.

C, carbon; H, hydrogen; Isooct, iso-octane; LHV, lower heating value; n-hep, n-heptane; Tolu, toluene.

0

10

20

30

40

50

60

70

80

90

100

Vol

% R

ecov

ered

Temperature °C

ULG 91

ULG 84

ULG 73

n-Heptane

Toluene

D156 DCN

4003002001000

Fig. 1 Volatility characteristics of the fuels used. ULG78 is a mixture of ULG 84 and ULG73 and is notshown. Iso-octane has a boiling point of 99.2�C

Table 3 Operating conditions

Speed Pin Tin Pexh IMEP Inj. pr lRPM bar, ab deg C bar, ab bar bar w/o EGR

Condition 1 1200 1.1 60 1.0 4 650 2.7Condition 2 2000 2 55 2.2 4 900 4.8Condition 3 2000 2 55 2.2 10 900 2.3

Pin, intake pressure; Tin, intake temperature; Pexh, exhaust pres-

sure; Inj. Pr, injection pressure.






3.1 Combustion phasing and emissions for

different fuels

Figure 3(a) shows CA50 plotted against SOI for dif-

ferent fuels for Condition 1 in Table 3, 1200 r/min, 4

bar IMEP, no EGR, and injection pressure of 650

bar; Figs 3(b) to 3(g) show, respectively, the corre-

sponding CD, NOx, maximum pressure rise rate

(MPRR), CO, THC (total hydrocarbons as C1) and

smoke plotted against CA50. These figures do not

consider all the fuels listed in Table 2; the results for

the missing fuels and some additional fuels can be

found in references [6–8]. It can be seen from

Figs 3(a) and 3(b) that fuels that are very different in

composition and volatility, TRF 84 and ULG 91 (tri-

angles), PRF 84 and ULG 73 (diamonds), and

n-heptane and the diesel fuel, D1, have very similar

CA50 for the same SOI and hence very similar com-

bustion delay for the same CA50. Moreover, for each

of these pairs of fuels of matching combustion

phasing at a given injection timing, NOx (Fig. 3(c),

MPRR (Fig. 3(d)), and CO (Fig. 3(e)) are also similar.

20

30

40

50

60

70

80

Pre

ssu

re p

[b

ar]

Crank Angle Degree, [CAD]

n-heptane

PRF 84

TRF 82

Injection

201510-10 5-5 0

Fig. 2a Average pressure for Condition 1, 1200 RPM, 4bar IMEP, no EGR, for three fuels withSOI = 28.0 CAD

-0.2

0

0.2

0.4

0.6

0.8

1

-50

0

50

100

150

200

250

-10 No

rmalise

d H

eat

rele

ase

HR

R [

J/C

AD

]

Crank Angle Degree, [CAD]

n-heptPRF 84TRF 82Injection

20151050-5

Fig. 2b Average HRR and normalized heat release(NHR) calculated from pressure in Fig. 2(a)

0.0

5.0

10.0

15.0

20.0

25.0

-40 -30 -20 -10

CA50

[CA

D]

SOI [CAD]

1200 RPM, 650 bar inj, 4bar IMEP,1.1 bar Pin

PRF 84TRF 84n-heptaneULG 91ULG 73D1 56 DCN

100

Fig. 3a CA50 vs SOI for Condition 1 for six differentfuels

0.0

5.0

10.0

15.0

20.0

25.0

30.0

35.0

40.0

3020100

CD =

CA

50 -

SOI [

CAD

]

CA50 [CAD]

1200 RPM, 650 bar inj, 4bar IMEP, 1.1 bar Pin

PRF 84 TRF 84 n-heptaneULG 91 ULG 73 D1 56 DCN

Fig. 3b CD vs CA50 for Condition 1 for fuels fromFig. 3(a)

0

200

400

600

800

1000

1200

1400

1600

0

NO

x [p

pm]

CA50 [CAD]


PRF 84

TRF 84

n-heptane

ULG 91

ULG 73

D1 56 DCN

302010

Fig. 3c NOx vs CA50 for Condition 1 for fuels fromFig. 3(a)






In general, if the combustion delay is large as for

ULG 91 and TRF 84, the mixture packets where

combustion takes place are likely to be lean,

because the local equivalence ratio will be nearer

the global equivalence ratio, and NOx and MPRR

will be low and CO and HC will be high compared

to fuels like n-heptane and D1, with low values of

combustion delay. For intermediate combustion

delays as for PRF 84 and ULG 73, NOx is higher than

for D1 or n-heptane for the reasons discussed in

[6, 7]. There is a difference in THC between TRF 84

and ULG 91 and n-heptane and D1 (Fig. 3(f))

although the difference within the matching pairs is

smaller than the difference between the non-match-

ing pairs. All the fuels show very low levels of

smoke, although D1 has clearly more smoke com-

pared to n-heptane (Fig. 3(g)). The IDW for both

n-heptane and D1 is positive but small – the two

fuels are most likely on the boundary between

mixing-controlled and premixed combustion. We

put a further condition that for premixed combus-

tion, the smoke level should be less than 0.05 FSN.

At each of the Conditions 2 and 3 (Table 3), an

EGR sweep was conducted; for each fuel, CA50 is

fixed at 11 CAD ATDC, the fuelling rate is fixed to

get a nominal IMEP without EGR, and then with

fuelling rate fixed, EGR is varied. EGR rate is defined

as the intake CO2 concentration expressed as a per-

centage of the exhaust FSN concentration. As EGR is

varied, SOI changes to keep CA50 fixed. The intake

pressure, at 2.0 bar absolute, was higher than at

Condition 1. As EGR increases, both the oxygen con-

centration in the intake and the normalized air fuel

ratio, l, decrease (Fig. 3 in reference [9]). For a given

fuel, the IMEP variation was 2 to 4 per cent over the

EGR range considered. We now consider the results

for the EGR sweep for Condition 3; the results for

0.0

5.0

10.0

15.0

20.0

25.0

0

MPR

R [b

ar/C

AD

]

CA50 [CAD]


PRF 84

TRF 84

n-heptane

ULG 91

ULG 73

D1 56 DCN

302010

Fig. 3d MPRR vs CA50 for Condition 1 for fuels fromFig. 3(a)

0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

0

CO [p

pm]

CA50 [CAD]



302010

Fig. 3e CO vs CA50 for Condition 1 for fuels fromFig. 3(a)

0

500

1000

1500

2000

2500

3000

3500

4000

0

THC

[ppm

]

CA50 [CAD]



302010

Fig. 3f CO vs CA50 for Condition 1 for fuels fromFig. 3(a)

0


3020100.00

0.01

0.02

0.03

0.04

0.05

0.06

0.07

0.08

0.09

Smok

e [F

SN]

CA50 [CAD]


Fig. 3g CO vs CA50 for Condition 1 for fuels fromFig. 3(a)






Condition 2 are similar. Figures 4(a) to (g) show,

respectively, CD, IDW, NOx, MPRR, CO, THC, and

smoke, plotted against EGR level, for Condition 3.

Again three pairs of fuels, ULG 91 and TRF 84, ULG

73 and PRF 84 and n-heptane and D1, have very

similar CD and IDW as EGR varies (Figs 4(a) and

(b)). Each of these fuel pairs also shows very similar

variation with EGR of NOx (Fig. 4(c)), MPRR

(Fig. 4(d)) and CO (Fig. 4(e)) in spite of the large

0

200

400

600

800

1000

1200

1400

1600

1800

NO

x [p

pm]

EGR %

2000 RPM, 900 bar inj, 10 bar IMEP, 2.0 bar Pin, CA50 = 11 CAD

PRF 84

TRF 84

n-heptane

ULG 91

ULG 73

D1 56 DCN

0 70605040302010

Fig. 4c NOx vs EGR for Condition 3

15

17

19

21

23

25

27

29

31

33

0

CD =

CA

50 -

SO

I [CA

D]

EGR %



70605040302010

Fig. 4a CD vs EGR for Condition 3

-4

-2

0

2

4

6

8

10

12

14

16

18

IDW

= C

A2

- EO

I [CA

D]

EGR %

2000 RPM, 900 bar inj, 10 bar IMEP, 2.0bar Pin, CA50 = 11 CAD

PRF 84

TRF 84

n - heptane

ULG 91

ULG 73

D1 56 DCN

0 70605040302010

Fig. 4b IDW vs EGR for Condition 3

0

2

4

6

8

10

12

14

16

18

0

MPR

R [b

ar/C

AD

]

EGR %


PRF 84TRF 84n-heptane

ULG 91ULG 73D1 56 DCN

70605040302010

Fig. 4d MPRR vs EGR for Condition 3

-100

100

300

500

700

900

1100

1300

1500

0

CO [p

pm]

EGR %

2000 RPM, 900 bar inj, 10 bar IMEP, 2.0bar Pin, CA50 = 11 CAD

PRF 84

TRF 84

n-heptane

ULG 91

ULG 73

D1 56 DCN

70605040302010

Fig. 4e CO vs EGR for Condition 3

0

100

200

300

400

500

600

700

THC

[ppm

]


0

EGR %

70605040302010

PRF 84

TRF 84

n-heptane

ULG 91

ULG 73

D1 56 DCN

Fig. 4f THC vs EGR for Condition 3






differences in volatility and composition between

fuels in each pair.

Figure 4(h) shows the average heat release rate

for ULG 91 and the diesel fuel D1 at Condition 3

with ~40 per cent EGR; the injection events for the

two fuels are also shown. The injection for ULG 91

is over much before combustion starts and combus-

tion for ULG 91 is clearly premixed while it is not

for D1. It is impossible to avoid equivalence ratios

where smoke formation is likely when IDW is nega-

tive. Even if IDW is slightly positive, there are very

likely to be rich regions, presumably from the fuel

injected later in the injection event, where smoke

will be formed. The engine-out smoke also depends

on the oxidation, which, in turn, depends on both

the temperature and excess oxygen available. As

EGR increases, oxygen levels as well as the tempera-

ture decrease, thereby reducing the smoke oxida-

tion. Figure 4(b) shows that, for n-heptane and D1,

IDW is negative or has only a low positive value

while Fig. 4(g) shows that, for EGR . 30 per cent,

smoke is much higher for these fuels than for TRF

84 or ULG 91, which have very much higher positive

values of IDW. This suggests that, when the com-

bustion is firmly in the premixed mode, smoke is

not formed in the first place and the deficiency of

oxygen at high EGR does not matter. However,

when smoke is formed, fuel composition and volati-

lity, which will affect mixing, will matter. Hence, n-

heptane, which does not contain aromatics and is

much more volatile than D1, has lower smoke than

D1 at high EGR. There is significant difference

between n-heptane and D1, especially at low EGR in

hydrocarbon emissions (Fig. 4(f)) where IDW is neg-

ative, although hydrocarbon emissions for TRF 84

and ULG 91, with large values of IDW, are similar.

Thus hydrocarbon emissions also do not quite

match for fuels with similar combustion phasing at

a given injection timing, unlike CO and NOx emis-

sions, when the combustion is in mixing-controlled

mode. A possible, although not definitive, explana-

tion for this is that CO and NOx are dominated by

processes in the bulk gas while hydrocarbon emis-

sions are also affected by the processes in the

quench and crevice layers and might be more

affected by volatility and compositional differences.

In summary, if the combustion is firmly in the pre-

mixed regime i.e. if the injection event is clearly

separated from the combustion event and smoke is

very low, NOx, CO and MPRR will be comparable for

two fuels at a given operating condition, if their com-

bustion phasing is matched for the same injection

timing, regardless of differences in volatility and

composition of the fuels. There might be some differ-

ences between such matching fuels in terms of

hydrocarbons but these differences are small com-

pared to fuel pairs with very different combustion

phasing (Fig. 3(f)). This observation has also been

demonstrated with other pairs of fuels in [7, 9].

Particularly noteworthy is that a fuel in the diesel

boiling range with 75 per cent aromatic content, Fuel

D4, in [7, 9], had similar combustion phasing at the

same injection timing and emissions as ULG 91 at

Condition 3. Thus, at Condition 3, three fuels of very

widely varying composition and volatility, namely

TRF 84, ULG 91, and Fuel D4 from [7, 9], all show

comparable combustion phasing at the same injec-

tion timing and emissions behaviour. In a different

engine, at low load, n-heptane and two European

commercial diesel fuels were found to have the same

combustion phasing at a given injection timing and

comparable emissions behaviour [20]. Such results

are also in line with observations in [21–23] that, in

low-NOx, low-smoke combustion in compression

ignition engines, fuel auto-ignition quality, which

0

1

2

3

4

5

6

Smok

e [F

SN]


0

EGR %

70605040302010

PRF 84

TRF 84

n-heptane

ULG 91

ULG 73

D1 56 DCN

Fig. 4g Smoke vs EGR for Condition 3

-5

15

35

55

75

95

115

135

155

175

-20 -10

HRR

[J/C

AD

]

Crank Angle [CAD]

2000 RPM, 10 bar IMEP, ~40%EGR, CA50 = 11 CAD,900 bar Inj

D1, 56 DCNULG 91D1 InjULG 91 Inj

3020100

Fig. 4h Comparison of heat release rates for ULG 91and D1 for Condition 3 with EGR ~40 per cent






determines combustion phasing, is far more impor-

tant than its volatility and composition.

Hence for all practical purposes, for premixed

compression ignition, it is sufficient for a surrogate

fuel to match the combustion phasing of the target

fuel at the same injection timing; for the same SOI,

the surrogate fuel should have the same ID and CD

as the target fuel.

3.2 Fuel auto-ignition quality and combustion

phasing

The phasing of an auto-ignition event in an engine

depends on the auto-ignition quality of the fuel and

the variation, with time and space, of the pressure,

temperature and equivalence ratio in the engine. The

auto-ignition quality of practical fuels has to be neces-

sarily defined by empirical parameters such as octane

and cetane numbers. The octane scale is based on

primary reference fuels (PRF), mixtures of the two

paraffins, iso-octane and n-heptane. However, the

auto-ignition behaviour of non-paraffinic fuels is very

different from that of PRF. In general, for a given

temperature, T, if the pressure, P, is increased, non-

paraffinic fuels become more resistant to auto-ignition

compared to paraffinic fuels such as PRF. For exam-

ple, measurements in shock tubes [24–26] show that

the ignition delay, t can be expressed as

t = f (T )P�n (1)

The value of the pressure exponent, n, is around 1.7

for paraffins whereas it has been measured to be

unity or less for full boiling range gasolines and

other non-paraffinic fuels [24–26].

A practical gasoline, which contains many non-

paraffinic components, can behave like different

PRF fuels at different pressure and temperature con-

ditions. Thus in an HCCI engine, a gasoline will

match a PRF fuel of a higher octane number at a

higher intake pressure with all other operating con-

ditions fixed [1]. Similarly the RON of a gasoline is

higher than its MON because RON is measured at

an engine condition where the unburnt gas tem-

perature, for a given pressure, is lower compared

to the MON test condition. The sensitivity,

S = RON 2 MON, of a gasoline is a measure of how

different its chemistry is compared to that of a PRF.

The true auto-ignition behaviour of a gasoline is

best defined by its OI, which is the octane number

of the equivalent PRF at the particular pressure and

temperature evolution in the unburnt gas [1].

OI = (1� K ) � RON + K � MON = RON � K � S (2)

K is an empirical constant depending only on the

pressure/temperature evolution with crank angle in

an engine. Hence it depends on engine design and

operating conditions and is a measure of how differ-

ent the test condition is from the RON test condi-

tion. If the temperature in the unburnt gas for a

given pressure is lower or, equivalently, the pressure

at a given temperature is higher than in the RON

test condition, K becomes negative and, for a given

RON, a fuel with lower MON will be more resistant

to auto-ignition and will have a higher OI [1]. This is

the case with modern SI engines, a consequence of

their evolution towards higher efficiency, and in

HCCI engines, which are run with boosted intakes

[1]. Generally such considerations do not come into

play for fuels in the diesel auto-ignition range,

DCN . ~30 or RON\~60. Thus the auto-ignition

quality of even an aromatic diesel fuel component

would be determined by the long hydrocarbon

chain attached to the aromatic ring. A consequence

of this is that different diesel fuels ranked, say, in

the IQT test, will retain the same ranking for auto-

ignition quality, at different engine operating condi-

tions. The DCN of a mixture of different diesel fuels

can be predicted, for all practical purposes, by

simple linear rules whereas this is not possible for

RON and MON of fuel mixtures in the gasoline

auto-ignition range if the components have different

chemical compositions [1].

3.2.1 Combustion phasing and auto-ignition qualityof gasoline fuels in premixed compressionignition combustion

The above insights about the auto-ignition beha-

viour of fuels in the gasoline auto-ignition range

have been gained from experiments where fuel and

air are fully premixed. In SI and HCCI engines, fuel

and air are fully premixed before compression and

the equivalence ratio remains constant while the

pressure and temperature in the unburnt mixture

change with time. In diesel engines, fuel is injected

in the cylinder when the ambient pressure is high –

around 40 bar in Fig. 2(a). Moreover this pressure

does not change very much between the time of

injection and the first auto-ignition, which starts the

combustion. The temperature will decrease slightly

as the liquid fuel vaporizes but the equivalence ratio

changes with time and space as the fuel mixes with

air. However, in all these cases the pressure and

temperature effects on auto-ignition chemistry

should be expected to be the same. It was demon-

strated in [8], using data for the conditions set out

in Table 3 but without EGR and CA50 fixed at 10

CAD, that the auto-ignition quality of gasoline fuels






could indeed be defined by an OI at a given condi-

tion; ID and CD varied linearly with OI. The K value

was negative – non-paraffinic fuels were signifi-

cantly more resistant to auto-ignition than indicated

by their RON and MON, as would be expected when

the pressure is high. This can be seen in Fig. 2,

where TRF 82 burns much later than PRF 84, which

has higher RON and MON than TRF 82. We now

illustrate these points using the results in the pres-

ence of EGR.

Tables 4 and 5 list CD and ID for Conditions 2

and 3 from Table 3, respectively, for EGR levels

of 10 per cent, 30 per cent, and 40 per cent; CDxx and

IDxx stand for CD and ID with xx per cent EGR.

Results are shown for fuels in the gasoline auto-igni-

tion range (RON . ~60) and also for n-heptane, for

comparison, although the results for n-heptane are

not used in further analysis below. Figures 5(a) and

(b) show ID10, for Condition 2 for different fuels

plotted against RON and MON respectively. There is

some correlation between ID10 and RON but no

correlation between ID10 and MON. However, there

is an excellent correlation if ID10 is plotted against

the OI = (1-K)RON + KMON, where K is 22.2

(Fig. 5(c)). The value of K is established by multiple

linear regression with ID10 as the independent vari-

able and RON and MON as dependent variables as

discussed in [1]. This approach can be used for all

the different delay times listed in Tables 4 and 5 (for

Condition 3); the K values and the R2 values for each

regression are also listed in each column in Tables 4

and 5. Similarly, it was shown in [8], for all three

conditions but without EGR, that the ignition and

combustion delays were linear functions of OI. The

differences in the K values at different conditions

cannot be explained on the available evidence but K

is negative in all cases as is to be expected at high

pressures.

Thus parameters such as CD and ID, which

describe combustion phasing at a given operating

condition and a given injection timing, vary linearly

with the appropriate OI. Hence a surrogate fuel of

the same RON and MON as the target gasoline will

have the same OI at any given condition and hence

will have the same combustion phasing for the same

injection timing as the target gasoline. A mixture of

iso-octane, n-heptane and toluene is the simplest

fuel system that can match any RON and MON [27].

With the experimental conditions considered here,

other properties that might affect mixing, such as

R² = 0.5791

10.0

11.0

12.0

13.0

14.0

15.0

16.0

70

ID1

0 =

CA

2 -

SO

I [C

AD

]

RON

2000 RPM, 4 bar IMEP, 900 bar Inj, 10%EGR, CA50 = 11CAD

9080

Fig. 5a ID10 vs RON for Condition 2 for differentgasoline fuels

R² = 0.0635

10.0

11.0

12.0

13.0

14.0

15.0

16.0

65

ID1

0 =

CA

2 -

SO

I [C

AD

]

MON


8575

Fig. 5b ID10 vs MON for Condition 2 for different gas-oline fuels

R² = 0.9848

10.0

11.0

12.0

13.0

14.0

15.0

16.0

80

ID1

0 =

CA

2 -

SO

I [C

AD

]

OI = RON +2.2S


12011010090

Fig. 5c ID10 vs OI = (3.2RON – 2.2MON) for Condition2 for different gasoline fuels






volatility, have little effect on combustion phasing.

Chemical kinetic models, which have been cali-

brated extensively against experimental data from SI

and HCCI engines and shock tubes, have been

developed recently for such model fuels [28].

3.2.2 Combustion phasing and auto-ignition qualityof fuels in the diesel auto-ignition range in pre-mixed compression ignition combustion

As already discussed, diesel-like fuels, with

DCN . 30 or RON\60, will burn in conventional,

mixing-controlled diesel mode at most conditions

except for very low loads because of the low ignition

delays. Increasing the injection pressure will extend

this load limit by shortening the injection event and

increasing the mixing rate. Diesel fuel D1 was com-

parable to n-heptane at Condition 1 as discussed in

section 3.1. In a different, 2L engine, with the com-

pression ratio of 11.4, n-heptane was again compa-

rable to two European diesel fuels in premixed

combustion [20]. In these cases the injection event

is completed before heat release occurs and the

smoke level is less than 0.05 FSN and n-heptane is a

reasonable surrogate for European diesel fuel. Fuel

D1 was compared to n-heptane at Condition 2 in [9]

and the combustion appeared to be on the verge of

being mixing-controlled for these two fuels [9].

Under such conditions, when smoke formation can-

not be avoided, n-heptane will give much lower

smoke compared to diesel, as is also shown for

Condition 3 in Fig. 4(g) and n-heptane is not a good

surrogate for diesel fuel.

Practical diesel fuels have a DCN between 40

and 60 or an OI between 0 and ~40. Ignition delay

varies little for fuels over this DCN range when the

compression ratio is greater than 14 so that they

have very similar combustion phasing for the same

injection timing and, hence in premixed combus-

tion, very similar emissions. For instance, fuels in

the diesel boiling range but with DCNs of ~39 and

~54 behaved very similarly at low load in a differ-

ent, 2L engine with a compression ratio of 14 [3]

and in the current engine with the compression

ratio at 16 [7]. A fuel in the gasoline boiling range

with DCN of 44 DCN was comparable to Fuel D1

and n-heptane at Condition 1 in the engine used in

this study [9]. In general, ignition and combustion

delays vary non-linearly with measures of fuel

autoignition quality. This is illustrated in Fig. 6

where ID10 from Table 4 and ID30 from Table 5

are plotted against OI. As conditions for autoigni-

tion become more difficult, ignition delay for a

given fuel increases. Moreover, the rate of increase

in ignition delay with OI, particularly for low OI,

appears to increase. Thus, in the 2L engine used in

[3], when the compression ratio was reduced to

11.4, the diesel fuel with 39 DCN did show larger

8.0

10.0

12.0

14.0

16.0

18.0

0

ID =

CA

2 - S

OI [

CAD

]

OI = RON - KS

ID10, Condi�on 2

ID30, Condi�on 3

12010080604020

Fig. 6 ID10 from Condition 2 and ID30 fromCondition 3 vs OI. The relevant K values are inTables 4 and 5. OI will be zero for n-heptane

Table 4 Combustion delay and ignition delay for Condition 2 at different EGR levels for different

fuels

Fuel 10% EGR 10% EGR 30% EGR 30% EGR 40% EGR 40% EGR

CD10 ID10 CD30 ID30 CD40 ID40

ULG 73 12.7 10.2 13.3 10.4 13.8 10.5PRF 84 13.4 10.1 14.1 11.1 15.2 11.4TRF 82 13.9 11.2 15.0 11.6 15.6 11.8ULG 78 14.0 11.1 14.6 11.8 15.3 12.1ULG 84 15.2 12.0 15.8 12.3 16.7 12.6ULG 91 18.8 14.4 19.8 15.3 21.3 16.2TRF 84 19.1 13.3n-hept 11.4 8.4 11.8 8.6 12.0 8.8K –2.28 –2.20 –1.60 –1.67 –1.32 –1.60Rsq 0.940 0.985 0.966 0.962 0.928 0.95

2000 r/min, 4 bar IMEP, CA50 = 11 CAD, 2 bar absolute intake pressure, 900 bar injection pressure.






ignition delays compared to a fuel with 54 DCN

[20] unlike at a compression ratio of 14.

In summary, in premixed mode, when the injec-

tion event is clearly separated from the combustion

event and smoke levels are very low, for compres-

sion ratios higher than 11.4, n-heptane will have

very similar combustion phasing at the same injec-

tion timing and will be comparable in emissions to

European diesel fuels of around 54 DCN. Hence, in

the premixed mode, n-heptane is a good surrogate

for practical European diesel fuels. If the compres-

sion ratio is greater than 14, n-heptane is also a

good surrogate for fuels with DCNs as low as 40,

for example US diesel fuels, in premixed CI

combustion.

CONCLUSION

Premixed combustion in compression ignition is

defined, in this paper, to occur when the injection

event is clearly separated from the combustion, and

engine-out smoke levels are very low, below

0.05 FSN. The operating conditions and the combus-

tion chamber geometry as well as the fuel properties

will determine whether premixed combustion is

attained. For instance, all else being equal, reducing

the injection pressure reduces IDW, and moves

combustion away from the premixed mode towards

the mixing-controlled mode. With practical diesel

fuels, with 40\DCN\60, premixed compression

ignition can occur only at very low loads, at a given

speed, compared to gasoline fuels; increasing the

injection pressure will extend this load limit. For

premixed combustion in compression ignition

engines, the following apply.

1. If two fuels have comparable combustion phas-

ing at a given operating condition and injection

timing, their emissions, particularly NOx and CO

and to a lesser extent, HC, will also be compara-

ble regardless of the differences in volatility and

composition between the fuels.

2. For gasoline fuels, with RON . 60, with the

experimental conditions considered, the combus-

tion phasing at a given injection timing and

operating condition depends only on the OI =

(1-K)RON + KMON where K is an empirical con-

stant depending only on the engine operating

conditions. Hence two fuels of the same RON

and MON will have the same OI and the same

combustion phasing for a given injection timing.

3. A mixture of iso-octane, n-heptane, and toluene

of the same RON and MON as the target fuel is a

very good surrogate for gasoline.

4. For premixed CI combustion, n-heptane is a

good surrogate for European diesel fuels with

DCN around 54 if the compression ratio is

greater than 11.4. If the compression ratio is

greater than 14, n-heptane is a good surrogate

for fuels with DCN . 40.

FUNDING

This work was supported by the EU contract, MTKI-

CT-2006-042242 Marie Curie Engine Efficiency,

under the Marie Curie Programme.

ACKNOWLEDGEMENTS

The contribution of many of our colleagues in Shell

Technology Centre Thornton (STCT), most notably H.

Jones and B. Head, made the engine experiments pos-

sible. The collaboration with Lund University, particu-

larly L. Hildingsson’s stay at STCT, was made possible

by the EU contract detailed under ‘funding’ above.

� Shell Research Limited and Authors 2011

Table 5 Combustion delay and ignition delay for Condition 3 at different EGR levels for different

fuels (Fig. 4(a) for CD)

Fuel 10% EGR 10% EGR 30% EGR 30% EGR 40% EGR 40% EGR

CD10 ID10 CD30 ID30 CD40 ID40

ULG 73 15.0 9.6 16.1 10.7 17.4 12.4PRF 84 15.0 10.1 16.7 11.9 18.5 13.8TRF 82 15.4 11.1 17.3 13.2 19.8 15.3ULG 78 15.5 10.5 16.8 11.6 18.4 13.6ULG 84 15.8 11.4 17.3 13.5 19.6 15.2ULG 91 17.1 15.0 21.1 17.1 24.5 19.6TRF 84 17.2 15.1 20.8 16.7 24.6 19.3n-hept 15.8 7.6 16.7 8.3 17.6 8.9K –2.95 –2.19 –1.95 –1.51 –1.84 –1.55Rsq 0.959 0.928 0.895 0.945 0.899 0.93

2000 r/min, 10 bar IMEP, CA50 = 11 CAD, 2 bar absolute intake pressure, 900 bar injection pressure.






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28 Andrae, J., Brinck, T., and Kalghatgi, G. T. HCCIexperiments with toluene reference fuels modeledby a semi-detailed chemical kinetic model. Com-bust. Flame, 2008, 155, 696–712.

APPENDIX

Notation

ATDC after top dead centre

CAD crank angle degree

CAx crank angle degree when x per cent of total

heat release occurs

CD combustion delay, CA50-SOI

CDxx CD at xx level of EGR

CN cetane number

CO carbon monoxide

DCN derived cetane number

EGR exhaust gas recirculation

EOI crank angle position of electrical signal at

end of injection

FSN filter smoke number

HC hydrocarbons

HCCI homogeneous charge compression

ignition

HRR heat release rate

ID ignition delay, CA2-SOI

IDxx ID at xx level of EGR

IDW ignition dwell, CA2-EOI

IMEP indicated mean Eeffective pressure

K K value, used in OI

MON motor octane number

MPRR maximum pressure rise rate

NHR normalized heat release

NOx nitrogen oxides

OI octane index = (12K)RON + KMON =

RON2KS

PRF primary reference fuels, iso-octane/

n-heptane mixtures

RON research octane number

S sensitivity (RON2MON)

SI spark ignition

SOI crank angle position of electrical signal at

start of injection

TDC top dead centre

THC total hydrocarbons, C1






surrogate fuels for premixed combustion in compression ignition engines

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