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Experimental comparison of DI and PFI

in terms of emissions and efficiency running Ethanol-85

DANIEL OTTOSSON KONSTANTINOS ZIORIS

Master of Science Thesis

Stockholm, Sweden 2014

Experimentell jämförelse av DI och PFI med avseende på emissioner och verkningsgrad

med Etanol-85

av

Daniel Ottosson Konstantinos Zioris

Examensarbete MMK 2014:78 MFM 156

KTH Industriell teknik och management

Maskinkonstruktion

SE-100 44 STOCKHOLM

Experimental comparison of DI and PFI in terms of emissions and efficiency running

Ethanol-85

Daniel Ottosson Konstantinos Zioris

Master of Science Thesis MMK 2014:78 MFM 156

KTH Industrial Engineering and Management

Machine Design

SE-100 44 STOCKHOLM

Examensarbete MMK 2014:78 MFM 156

Experimentell jämförelse av DI och PFI med avseende på emissioner och verkningsgrad med

Etanol-85

Daniel Ottosson

Konstantinos Zioris

Godkänt

2014-10-06

Examinator

Andreas Cronhjort

Handledare

Andreas Cronhjort

Uppdragsgivare

Scania AB

Kontaktperson

Eric Olofsson

Sammanfattning

Det har på senare tid blivit allt viktigare att hitta ett alternativ till fosila bränslen i våra fordon på

grund av, dels det höga bränlsepriset men, framför allt för att reducera deras negativa effect på

klimatet. Ett sådant alternative finns redan idag tillgängligt i stor skala, nämligen etanol. Etanol

har , förutom låg klimatpåverkan, egenskaper som gör det till ett fördelaktigt bränsle i

förbränningsmotorer. Det höga oktantalet tillsammans med det högre förångingsvärmet gör att

etanol blir väldigt knackbeständigt vilket I sin tur möjligör en motor med högre

kompressionsförhållande och verkningsgrad. Traditionella Otto motorer har använt

portinsprutning medans moderna motorer mer och mer gått över till direktinsprutning. Det finns

manga fördelar med direktinsprutning och då framförallt högre verkningsgrad på grund av den

högra volumetriska verkningsgraden och knackbeständigheten. Nackdelar med direktinsprutning

kan vara sämre A/F blanding och ökad komplexitet. De positiva effekterna av direktinsprutning

tycks ytterligare förstärkas när det används i kombination med etanols bättre charge cooling

effekt och högre oktantal.

För att undersöka om en etanol driven otto motor är lämplig för HD undersöks både DI och PFI

med avseende på verkningsgrad och emissioner på en Scania D12 HD motor. Motorn modifieras

för att möjligöra otto drift med tändsstift. Scanias XPI system används som DI med endast några

mindre modifikaktioner för att möjliggöra etanol drift.

DI systemet utvärderas vid to insprutningsvinklar, en stratifierad och en homogen. Ett SOI svep

görs för att identifiera de optimala SOIs. Homogen DI och PFI produceras liknande resultat

medans stratifierad DI sticker ut på grund av sin; knackbeständighet, mycket snabbare

förbränning, lägre HC emissioner och lägre CoV. Railtryck har ingen eller lite effekt på

homogen DI medan den ses öka förbränningshastigheten, HC och CO emissioner, verkningsgrad

samt sänkt CoV för stratifierad DI.

Inga slutsater kunde dras gällande verkningsgraden för de olika insprutningssystemen. Detta på

grund av problematiska bränsleflödes mätningar.

Master of Science Thesis MMK 2014:78 MFM 156

Experimental comparison of DI and PFI in terms of emissions and efficiency running Ethanol-85

Daniel Ottosson

Konstantinos Zioris

Approved

2014-10-06

Examiner

Andreas Cronhjort

Supervisor

Andreas Cronhjort

Commissioner

Scania AB

Contact person

Eric Olofsson

Abstract

It has in recent year become more and more important to find an alternative to fossil fuel in our

vehicles due to the increasing fuel price and to reduce their negative impact on the environment.

One alternative is already in widespread use around the world, namely ethanol. Ethanol has,

besides its environmental qualities, properties that makes it a favorable fuel to use in Internal

Combustion Engines (ICE). Its high octane rating combined with its high heat of evaporation

makes it resilient against knock which allows for an engine with higher compression ratios and

overall increased efficiency. The traditional SI engines use Port Fuel Injection (PFI) while

modern engines are moving towards Direct Injection (DI). There are many advantages of the DI

system, most notably increased efficiency and performance by increased volumetric efficiency

and knock suppression while poorer air/fuel mixing and added complexity are the negatives. The

positive effects of DI seem to be further increased when utilizing ethanol's improved charge

cooling effect and its higher octane number.

In order to investigate if an ethanol fueled SI engine is suitable for HD application both DI and

PFI are evaluated in terms of efficiency and emissions on a Scania D12 HD engine. The engine

is modified to accomedate a sparkplug. The Scania XPI is used as DI with some light

modifications in order to run ethanol.

The DI system is evaluated at two SOIs, stratified and homogenous, and a SOI sweep is

performed for both DI and PFI in order to find the optimum SOIs. DI homogeneous and PFI are

found to produce similar results while DI stratified stands out with its; low knock propensity,

much faster combustion, lower HC emissions and lower CoV of IMEP. Railpressure is found to

have little or no effect on homogeneous DI while it slightly increases the combustion speed, HC

and CO emissions and efficiency as well as lowers the CoV of IMEP for stratified DI.

No conclusions can be drawn about efficiency in this study due to a lack of reliable fuel flow

measuraments.

Contents

1 Introduction 81.1 Background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8

1.2 Purpose and definitions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8

1.2.1 Objectives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9

1.3 Contributions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9

1.4 Delimitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9

1.5 Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9

2 Literature review 102.1 Ethanol . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10

2.1.1 Properties of ethanol . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10

2.1.2 Emissions from ethanol combustion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11

2.2 PFI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12

2.3 Direct Injection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12

2.3.1 Spray, Air and Wall guided injection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12

2.3.2 Injectors and fuel jets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13

2.3.2.1 Injection pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14

2.3.3 Charge cooling, Volumetric efficiency & Knock tendencies . . . . . . . . . . . . . . . . . . 14

2.3.4 Charge mixture and motion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15

2.3.5 Injection strategies and timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15

2.3.5.1 Split injections . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15

2.3.5.2 Effects of injection timing on efficiency, IMEP and heat release . . . . . . . . . . 16

2.3.6 Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17

3 Experimental Methodology 183.1 Hardware . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18

3.2 Test plan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18

3.3 Calculation and evaluation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19

3.3.1 Heat release rate . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19

3.3.2 Mass fraction burned . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19

3.3.3 Efficiency and fuel consumption . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20

3.3.4 Volumetric efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20

4 Results 214.1 Start of injection sweeps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21

4.1.1 PFI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21

6

4.1.2 DI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23

4.2 DI vs PFI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24

4.2.1 Cylinder pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25


4.2.3 Mass fraction burned . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27

4.2.4 Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28

4.2.5 Efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29

4.2.6 Coefficient of Variation in IMEP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29


4.3 Rail pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30

4.3.1 Cylinder pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31


4.3.3 Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33

4.3.4 Efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33

4.3.5 Coefficient of Variation in IMEP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34


5 Summary & Conclusions 355.1 DI vs PFI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

5.1.1 Load potential . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

5.1.2 Combustion speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

5.1.3 emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

5.1.4 Efficiency and combustion stability . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

5.2 DI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 36

5.2.1 Rail pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 36

6 Discussions and future work 376.1 Discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37

6.1.1 Restrictions of Hardware and test bed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37

6.2 Future work . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38

6.2.1 Test plan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38

6.2.2 Hardware and test bed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38

Nomenclature 39

References 40

7

Introduction

1.1 Background

It has in recent year become more and more im-portant to find an alternative to fossil fuel in ourvehicles due to the increasing fuel price and to re-duce their negative impact on the environment. Onealternative is already in widespread use around theworld, namely ethanol. Ethanol can be producedby feedstock like sugar canes or biomass and issubsequently a renewable alternative to the finiteand polluting fossil fuels.

Ethanol has, besides its environmental qualities,properties that makes it a favorable fuel to use inInternal Combustion Engines (ICE). Its high octanerating combined with its high heat of evaporationmakes it resilient against knock which allows for anengine with higher compression ratios and overallincreased efficiency. Furthermore it has been shownthat ethanol reduces tailpipe emissions such as;NOx due to the cooler charge; CO, CO2, PM and HCdue to the chemical composition containing oxygenatoms. A drawback of ethanol is the reduced lowerheating value (LHV) and therefore a lower energydensity as well as a reduced stoichiometric A/Fratio. This leads to a higher fuel consumption andreduced mileage for the same fuel tank and enginespecifications. The volumetric energy content ofthe A/F mixture is however increased with ethanol,leading to a higher power output during similarconditions.

All in all; ethanol has beneficial qualities bothin terms of performance and environmental impact

making it an interesting fuel to use in Spark Ignited(SI) engines.

The traditional SI engines use Port Fuel Injec-tion while modern engines are moving towardsDirect Injection (DI). There are many advantagesof the DI system, most notably increased efficiencyand performance by increased volumetric efficiencyand knock suppression while poorer air/fuel mix-ing and added complexity are the negatives. Adrawback related to the poorer air/fuel mixing isan increased CO emittance due to locally richermixtures while the HC is expected to decreasedue to the absence of wetting of the crevices. Thepositive effects of DI seems to be further increasedwhen utilizing ethanol’s improved charge coolingeffect and its higher octane number. Volumetricefficiency is expected to increase for the samereason with DI, it will however decrease with PFIwhere the improved cooling is counteracted by theincreased mass needed for a stoichiometric mixture

A more detailed explanation and discussion aboutethanol and DI is presented in Chapter 2.

1.2 Purpose and definitions

Ethanol is an attractive fuel for the future and it isalready in widespread use in passenger cars. Com-pression ignited engines running ethanol (ED95)exists but requires complex and expensive additivesto make ethanol suitable for CI combustion. NoHeavy Duty (HD)s trucks have adopted an SIethanol engine even though it seems like a promis-

8

This section will present the background, purpose and contributions of the thesis. It is meant toanswer the questions what, why and how the thesis topic is researched.

ing solution. The question then arises; Is an ethanolfueled SI engine suitable for HD applications ?

Two different injection concepts exist for SI en-gines, namely PFI and DI. The latter has beenshown to increase volumetric efficiency and knocksuppression but suffer from poorer air/fuel mixingand additional complexity. It might be difficultto produce a homogeneous stoichiometric mix-ture with DI while the opposite is true for PFI.An inhomogeneous mixture will increase theafter-treatment systems workload and might thusincrease the tailpipe emissions. This leads to thequestions; What performance and efficiency gainscan be achieved with DI, how will it effect theemissions and which one of these systems are moresuitable for an ethanol fueled HD SI engine ?

1.2.1 Objectives

Based on the background and the questions askedin this section the following tasks will be pursued:

• Study the literature on both ethanol combustionand PFI/DI injection systems. Find the param-eters that effects engine performance and emis-sions. Present the findings in a literature review.

• Evaluate the two different injection systems interms of consumption, emissions, combustionstability and load potential (knock resistance) ina single cylinder HD SI engine fueled by E85 (85

vol% ethanol and 15 vol% gasoline) fuel.

• Evaluate the DI system by studying the interest-ing parameters found in the literature review.

1.3 Contributions

Contributions of academic novelty in this thesis isDI operations with high injection pressures (up to1600 bar) by using the Scania XPI system. Previ-ous research of DI has been performed with injec-tion pressures of at most 300 bar.

1.4 Delimitations

The cylinder head used will be a slightly modifieddiesel head. This means that there is no naturalplace to fit the spark plug needed for Otto op-erations. A pressure sensor hole at around 3/4

radius will be used instead. This might not beideal for PFI systems where a centrally mountedspark plug is preferable. An optimum solutionwould be to manufacture one or two new cylinderheads, one with just a centrally mounted sparkplug and one with both a central injector and sparkplug. This is however more difficult than it soundsand would require extensive design work beforemanufacturing. Secondly, It would draw fundsfrom the thesis budget, money that could otherwisebe used for engine tests. Thirdly it would take timefrom the tests as the engine would have to be dis-assembled to change the head between comparisons.

An already installed intake manifold will be usedfor PFI. Its injector position is further upstreamthan optimum and will increase the HC emissionsduring the transients, but for the same reasons asabove (time and money) it will still be used. Theeffects of the non-optimum injector placement willbe taken into account when discussing the findingsfrom the tests.

1.5 Methods

To obtain and assess the scope of the tasks presentedin Section 1.2 a pre-study of existing research isneeded. The main sources of such research arepublications from SAE (Society of AutomotiveEngineers) and MTZ (Motor techniche zeitung).Books as sources are used to a lesser extent in thisthesis due too their often to brief or unexisting textsabout this relatively new subject.

Engine tests is the only way to answer the othertwo tasks. A single cylinder Scania engine will bemodified in order to run both DI and PFI. Detailsabout the engine and the engine tests is presentedin Chapter 3.

9

Literature review

2.1 Ethanol

Ethanol is a renewable fuel produced by fermentingsugars from crops such as sugar canes or corn.First generation ethanol uses fossil fuels or biomassas energy source for production. It is howeverestimated that the use of ethanol instead of gasolineleads to a reduction in total life cycle CO2 emissionsby between 20 and 80% depending on the energysource and crop used. Second generation ethanol iscurrently being studied. Here, agriculture residuesare used instead of the crops of the first generation.This evades the problem of "food vs. fuel" andfurther reduces the total CO2 emissions [1, 2].

The total world production of ethanol was in2012 107327.8 millions of liter [3]. The total en-ergetic contribution to the worlds transportationsector in the same year by ethanol was around 2%[4].

2.1.1 Properties of ethanol

The main chemical difference between ethanol andgasoline or diesel is the addition of oxygen. Thisgives ethanol a different set of properties as listed inTable 2.1.

Ethanol has a reduced LHV and therefore a lowerenergy density as well as a reduced stoichiometricair/fuel ratio. This leads to a higher fuel con-sumption and reduced mileage for the same fueltank. The energy content of the air/fuel mixture ishowever increased with ethanol, leading to a higherpower output during similar conditions [6].

Gasoline EthanolLower heating value(MJ/kg)

42.7 26.8

Density (kg/m3) 715-765 790Research octane num-ber [5]

95-98* 110

Boiling temperature(◦C)

25-215 78

Latent heat of vaporiza-tion (kj/kg)

380-500 904

Self-ignition tempera-ture (◦C)

300 420

Stoichiometric air/fuel-ratio

14.7 9

Laminar flame speed(cm/s)

33 39

Mixture calorific value(MJ/m3)

3.75 3.85

Ignition limits in air:Lower limit 0.6 3.5Upper limit 8 15(Vol %)

Table 2.1: Comparrison of gasoline and ethanolproperties [6, 5]. *Typical european gasoline

The higher octane rating and autoignition tem-perature for ethanol means that it is more resilientagainst knock. A higher compression ratio cantherefore be utilized and increase the power output.Another advantage is the higher heat of evaporationmeaning that more energy is taken from the hot airin order to evaporate the fuel leading to a coolercharge. This further increases the power as well asknock suppression [7, 8]. A drawback of the higherheat of evaporation is its negative effect on coldstart where excessive enrichment is needed in order

10

Several studies investigating ethanol and DI exists. This section will present a review of theprevious research applicable to the object of this thesis.

to start the engine. This is a waste of fuel and ithas a direct impact on HC emissions. One possiblesolution is to combine DI and PFI systems. TheDI is used with E85 during warm operation and asecondary PFI system using gasoline is activatedduring severe cold start conditions [7].

Because of the different properties of ethanolsit will behave differently from gasoline when in-jected into the engine. Visualization studies (usingLaser absorption scattering, planar laser-inducedfluorescence and MIE scattering) of ethanol sprayswith around 10 MPa injection pressure have shownthat it produces a more homogeneous (locally) andless liquid clouds due to its higher vapor pressure aswell as faster diffusion rate and evaporation [9, 10].A consequence of this is also a reduced spraypenetration, desirable for late injections [10, 11].

2.1.2 Emissions from ethanol

combustion

A straight switch from petroleum fuels to ethanolwill have several advantages in terms of emissions.As discussed in Sec. 2.1.1 the higher heat of vapor-isation leads to a cooler charge and subsequently alower peak temperature. NOx is known to form athigh temperatures and will therefore be reduced.Several studies confirm this [6, 12, 13, 14, 15].The ethanol has a much higher oxygen contentthan gasoline. This leads to lower CO emissionssince there is more oxygen available to create CO2

[16, 14, 15]. The overall CO2 output is still lowerwith ethanol than petroleum since it has a higherH/C ratio leading to a higher H2O/CO2 ratio fora complete combustion. [17]. Hydro carbons areshown to decrease with ethanol due to the shortercarbon chains of ethanol which evaporates easier.[13, 15]

An example of such a straight switch discussedabove is performed by Nakata et.al [13] where astandard 1500cc 4 cylinder engined is fueled withboth ethanol and gasoline. The measurements aretaken at part load (BMEP 0.2 MPa) with WOT (Wide

Open Throttle). The ignition timing is changedto ensure MBT (Maximum Brake Torque) for allblends. The result is shown in Figure 2.1.

Figure 2.1: Emissions from ethanol andgasoline [13]

It is clear that NOx, THC and CO2 decreases as theethanol content increases. The faster combustion ofethanol is also evident as the ignition timing hasto be advanced closer to TDC (Top Dead Center)in order to obtain MBT as the ethanol contentincreases.

PM and soot is shown to decrease with ethanol andthis is again because of the higher oxygen content.Ethanol contains both less soot-prone hydrocarbonsand hinders them to form soot [18, 15]. Di Iorioet.el. [19] compares the mass concentration ofparticles on a DISI engine. The concentration ismeasured with an opacitimeter and the particle sizewith a differential mobility spectrometer (DMS500).The results are shown in Figure 2.2.

The test compares gasoline (E0) to pure ethanol(E100) for homogeneous, stratified stoichiometricand stratified lean combustion. PM is shown to de-crease with the use of ethanol and while it increaseswith stratified combustion, it still remains low. Theslight increase is probably due to the locally richermixture.

11

Figure 2.2: Particulate emissions fromethanol and gasoline [19]

2.2 PFI

PFI has traditionally been used in some form bySI engines since its inception. First in the form ofcarburetors and later as a electrically controlledinjector placed inside the intake manifold. The fuelis injected into the intake maniforl and mixes withthe air before it enters through the intake port andinto the combustion chamber.

In a PFI engine the fuel is injected and pre-pared outside of the combustion chamber, andtherefore it impinges the inside of the port and theintake valve. In this case, the injected fuel meets thewalls of the intake channel and the required heat forthe fuel vaporization is mainly taken from the hotwalls instead of the air, leading to a hotter chargethan compared to DI where most of the heat istaken from the air [20, 21, 22]. However, the chargecooling effect of a gas-ethanol mixture with PortFuel Injection is not yet fully clarified. The paperby Kar et. al. [20] investigates this. In theory if theevaporation process had been performed adiabat-ically, the temperature of the mixture would havebeen decreased by 60 C degrees. Thereby the airflowwould have been increased by 20%. However inthe case investigated for the PFI arrangement, theairflow measured was not more than 1%. Thus thedirect conclusion is that for PFI engines workingat λ = 1 and WOT, the major amount of heatneeded for the fuel’s vaporization is taken by thesurrounding walls. Thus the effect of charge coolingis not utilized but in a very small extent [20].

The volumetric efficiency is therefore expected toincrease when using ethanol. The heat is taken fromthe walls and the charge can therefore not utilizeethanol’s higher heat of vaporization. Moreover,the increase in density is counteracted by the extramass of ethanol needed for a stoichiometric mixture[23, 13, 24].

An emissions comparison between DI and PFIis discussed in Section 2.3.6.

2.3 Direct Injection

More and more modern engines are moving awayfrom the traditional PFI and towards DI. This sec-tion will discuss the differences as well as the bene-fits and drawbacks of the DI system in general andparticularly using ethanol.

2.3.1 Spray, Air and Wall guided

injection

As discussed (See Section 2.3.4 & 2.3.5) there aremany benefits to be gained with DIs ability to in-ject arbitrarily during either the intake or compres-sion stroke to form an either homogeneous or strat-ified mixture. The stratified injection requires injec-tion equipment configurations that enables a mix-ture with a lambda gradient across the combustionchamber. Such configurations can be either spray, airor wall guided injection systems as shown in Figure2.3.

Figure 2.3: DI systems [25]

The wall guided systems use a side mounted injec-tor that injects onto the piston. The pistons shapewill then direct the fuel towards the spark plug.This will not alone create a satisfactory mixtureand needs a specifically designed air motion. One

12

drawback is the increased PM emissions due todiffusion controlled combustion of the fuel on thepiston surface as well as lower fuel pressure andpoorer mixing [26].

The air guided system will, as described inSection 2.3.4, use the in-cylinder air motion tocreate a locally rich mixture around the sparkplug. This system is highly dependent on the airmotion and might require partial throttling leadingto an increase in pumping losses. Since there isno interaction between the walls and the spray theair guided system produces less HC than the wallguided.

Finally, there is the spray guided system whichhas a central or side mounted injector aimed at thespark plug. A good mixture can thus be createdunhindered around the spark plug without the aidof either the walls or in-cylinder flows. This leadsto a better control of the charge and therefore thehighest potential when it comes to emissions andfuel consumption. The demands on the equipmentin this system is higher than for the other since thecharge is solely created by the injector as well aspotential cyclic temperature fatigue of the sparkplug since it is cooled by the spray. Studies haveshown that there is a slight advantage in combustionstability when using a centrally mounted injector[27].

2.3.2 Injectors and fuel jets

Different types of injectors exists for use in a DIengine. The most predominant are the multi-holeor the outwardly opening injectors. The multi-holeis generally solenoid driven and injects fuel via aseries of openings creating multiple jets. The out-wardly opening injector on the other hand is usuallyof piezo driven and uses a circular opening whichcreates an even curtain of fuel. A MIE scattering im-age of the sprays from the two types is shown inFigure 2.4.Experimental investigations of the sprays usinga constant pressure chamber have shown thatthe outwardly-opening injector results in a better

Figure 2.4: Mie scattering image of the twoinjector types, piezo (left) and multi-hole

(right) [28]

atomization than the multi-hole. This is due tothat the very thin curtain or cone interacts withthe air in such a way that it creates vortexes whichtransforms the kinetic energy into rotational energy.The multi hole injects several high-energy jets thatare more difficult to retard and mix with the air[29]. An other study adds more advantage with theoutwardly opening injector and that is the fasterresponse, improved mixing and quantity controlwhich leads to the possibility of multiple injectionsper cycle [30]. Although the outwardly-openinginjector produces a better spray, steady state enginetests with the two injector types have been con-ducted with the conclusion that they are similar inperformance [28].

An important aspect of the (multi-hole) injec-tors in a DI system is the umbrella angle, i.e. theangle between two opposing jets. This angle needsto be adjusted so that impingement is avoided.Several studies investigate the effect of the um-brella angle both numerically (CFD) and optically[31, 32, 33]. The study Skogsberg et.al. [31] showsthat an increased angle helps reduce the penetrationdue to a shift in axial (downwards) and radial(sideways) velocity. It will thus reduce the riskfor impingements during late injections when thepiston is close to TDC as well as improving theevaporation since the longer distance between jetsutilizes the air better. The A/F mixing will howeverbe reduced as shown by Dahlander [32] since the

13

individual jets/plumes never collide and interact toform a coherent A/F cloud. There is also evidencethat an increased L/D ratio of the nozzle holes willreduce the diameter of the fuel drops which willimprove the vaporization [31, 33].

2.3.2.1 Injection pressure

The effects of injection pressure on penetrationand emissions is studied by Allocca et. al. [34].A multi hole injector injects into a vessel withair at atmospheric pressure and a temperature ofabout 25

◦C. The penetration is then measuredoptically. E10, E85 and gasoline is tested at injectionpressures of both 50 and 100 bar. A higher injectionpressure leads to a longer penetration. For E85 thepenetration increases from around 50 to 60 mm forE85. The increase is greater for gasoline where thepenetration goes from 55 to 70 mm. This is likelydue to the faster vaporization of ethanol becomesmore apparent at higher pressures.

The study continues with engine tests wherethe flame front velocity as a function of injectionpressure is investigated. It is seen that the fastestflame fronts with E85 is obtained with the lowerinjection pressure (50 bar) while the opposite istrue for E10 where 100 bars produces a faster flamefront. The unexpected result regarding E85 is likelydue to the longer penetration impinging the piston.

2.3.3 Charge cooling, Volumetric

efficiency & Knock tenden-

cies

One advantage of the DISI engines in general, com-pared to PFI engines, is the higher utilization of thecharge cooling effect. The heat for fuel vaporizationis mainly taken by the charge instead of the walls.Consequently the charge is cooled further and theknock limit is increased. This charge cooling effectcan be utilized even more when using ethanol sincethe heat of vaporization for ethanol is much higherthan for gasoline. This cooler charge leads to aneven further increased knock limit as well as a spark

timing closer to MBT. In addition, due to the higherdensity of the cooler charge, more air can enter thecombustion chamber and subsequently increase thevolumetric efficiency [35, 24].

A comparison of volumetric efficiency and theignition timing for MBT between gasoline and E100

at full load is shown in Figure 2.5.

Figure 2.5: Full load comparison of ethanoland gasoline [36]

The volumetric efficiency is shown to decrease withethanol for PFI (See Section 2.2) while the oppositeis true for DI. The increase is around 5%. No knockis produced with ethanol and MBT spark timing canbe achieved at all engine speeds while an advancedspark timing is needed for gasoline during lowengine speeds [36].

The DI system is more resistant to knock dueto the previously mentioned charge cooling effects.There is however a trade-off between volumetricefficiency and knock resistance. An early injection(good for volumetric efficiency) will lead to the fuelspending more time and move more freely insidethe chamber resulting in higher charge temperaturedue to the heat transfer from the walls ratherfrom the charge to the fuel. This will increase thetendency for knock. A later injection will on theother hand better utilize the cooling effect of thefuel vaporisation and give a cooler final charge and

14

thus suppress knock [21].

2.3.4 Charge mixture and motion

How well the air and fuel mixes during the intakeand compression is largely dependent on injectiontiming. As the air is induced it creates tumblesinside the cylinder. When the compression strokebegins the spaces for these tumbles will reduce mak-ing the structured motions transform into turbulentmotions. The turbulent motions are desired sincethey increase the flame speed and improves thecombustion efficiency. A major advantage of the PFIis its capability to produce a homogeneous mixture.When fuel is injected in the intake it both has moretime to properly mix and that mixing occurs beforeentering the cylinder, thereby reducing the effectsof air motions inside the cylinder. PFI will on theother hand never be able to use the injection as animpulse to increase air motions.

A homogeneous mixture will always be moredifficult with DI. The injection will effect thein-cylinder motions, creating a trade-off betweenturbulence and mixture homogeneity. This is exten-sively studied by Knop et. al. [35]. The influence ofinjection timing on turbulence is shown in Figure2.6.

Figure 2.6: Turbulence intensity relative PFIfor different injection timings [35]

A late injection will inject into a fully developedtumble and enhance it further, leading to more tur-bulence. This is however not beneficial to mixingsince the fuel will be trapped in the tumble creatinga layer between rich and lean mixtures. The oppo-

site is true for early injections where there are nostructured motions and the fuel can move aroundmore freely and thereby offer a more uniform mix-ing. Since the tumble is not enhanced by the injec-tion the resulting turbulence will be lower.

2.3.5 Injection strategies and tim-

ing

The combustion is highly dependent on the injec-tion since the mixture quality is determined by theinjection strategy and timing.

An early injection will improve mixing but itmight however create an overall lean mixtureinstead of the desired slightly rich around the sparkplug and lean around the walls. A lean mixturewill produce lower soot but higher NOx emissions.If the mixture becomes too lean it might not beignitable. The opposite is true for a late injectionwhere NOx is likely to decrease while soot willincrease and where the lean mixture is difficult toignite the overly rich mixture produced by a too lateinjection might pre-ignite [37, 38]. HC is expectedto increase with late injection due to the poorer andricher mixture [38].

The DI engine has the ability to run globallylean, meaning that it is richer around the sparkplug but overall there is an excess of oxygen insidethe chamber. This results in the advantageouscombustion of a rich mixture but with the lowerfuel consumption of a lean. This can also reducethe pumping losses since the need for throttling isreduced. The mixture might be too rich locally andheterogeneous at higher engine loads and speeds.It is then better to run the engine with a morehomogeneous stoichiometric mixture by injectingearly [38].

2.3.5.1 Split injections

As mentioned in Section 2.3.3, knock is the mainlimitation when trying to improve the volumetricefficiency and thermal efficiency (by increased CR).

15

As explained in Section 2.3.3 there is a trade-offbetween volumetric efficiency and knock resistancesince an early injection will increase the volumetricefficiency while a late injection increase the knocklimit. Maximizing both is therefore not possible.This also means that when the injections is forced tobe retarded due to knock, the power output of theengine will decrease. A way to utilize the benefitsof both injection timings is to use split injectionswhere one portion of the fuel is injected duringthe intake stroke and the remainder of the fuel isinjected during the compression stroke. Studieshave shown an increase in IMEP of 2-3% comparedto a single injection strategy, mainly due to theincreased volumetric efficiency [21].

A stratified mixture is favorable for mediumloads. In that case only a small amount of fuel is in-jected in the intake stroke for improved volumetricefficiency. The majority of the fuel is injected lateduring compression in order to achieve stratifiedconditions. The opposite is true for full load wherea homogeneous mixture is advantageous. Here themajority of the fuel is injected early, giving it timeto become homogeneous. The remainder is injectedlate to suppress knock [38].

2.3.5.2 Effects of injection timing on ef-

ficiency, IMEP and heat release

Each engine configuration has a combustion win-dow where no misfires occurs. By sweeping the endof injection (EOI) inside this window it is possibleto study how the injection timing affects the com-bustion efficiency. Such study is done by Oh et. al.[9]. By advancing the EOI, less fuel will result inmore lean mixtures which will result in failed strat-ification under low charge temperatures. Retardingthe EOI on the other hand leads to the fuel beinginjected at very high charge temperatures whichmay lead in a sort of stagnation and produce locallyrich mixture, hard to ignite. This will result in theincrease of the HC and CO emissions. Moreover, toomuch retardation (combustion phasing) increasesthe exhaust gas temperature. This is because the

late ignition moves the burn to the exhaust valveopening and thus higher exhaust gas temperature.An optimum EOI where maximum combustionefficiency is found where these are balanced [9].Ethanol blends have a more retarded combustionwindow than gasoline. The explanation lies in thealready mentioned higher vaporization rate and thelocally enhanced homogeneity (See Section 2.1.1).

The retardation of EOI results in an increase ofIMEP since the effective work by retarding the com-bustion phasing is increased. IMEP is subsequentlyexpected to increase with increased ethanol contentsince the EOI can be retarded [9, 23]. However,IMEP decreases with ethanol content for the sameEOI. The advanced combustion phase when burn-ing ethanol blends causes a sharper in cylinderpressure. This results in increased negative workand therefore a decreased IMEP [9].

Figure 2.7: Heat release rates for a)stratified and b) homogeneous DI operation

at different engine speeds [39]

The heat release rates for both stratified lean andhomogeneous stoichiometric combustion at 1000,1400 and 2000 rpm is shown in Figure 2.7. SOI is

16

advanced in order to sustain CA50 at 8◦C for the

stratified.

The heat release rate is stretch and subsequentlylowered with increased engine speed for stratifiedcombustion but is unaffected by engine speed whenthe heat release is plotted in time instead of crankangle. For the homogeneous it remains constantover crank angle but increases with increasingengine speed when plotted in time due to theincreased flame speed associated with increasedengine speed.

The previously discussed charge cooling (seeSection 2.3.3) by direct injection is clearly visiblein Figure 2.7a. where the heat release rate turnsnegative at SOI .

2.3.6 Emissions

The emission output of a DI equipped enginedepends on many factors. These factors are contin-uously discussed throughout this literature review.For a detailed look at what and how these factorsand parameters affect emissions go to the individualsections.

The only real general differences in terms ofemissions regarding DI compared to PFI is theincreased CO emission for DI. This is due to themore heterogeneous mixture discussed in Section2.3.4. The heterogeneous mixture leads to locallyrich combustion where the lack of oxygen willstop CO from becoming CO2 [35, 38]. The seconddifference is the higher HC for PFI since the fuel/airmixture will enter crevices and come in contact withoil films along the liners.

As mentioned in Section 2.3.5 the stratifiedcharge engine has the benefits of a reduced fuelconsumption without sacrificing performance byrunning lean. This will however cause problemswhen it comes to exhaust after-treatment since thethree-way-catalyst cannot be applied due to non-stoichiometric conditions [40]. Stratified operationis shown to increase PM engine out emissions,

especially when running lean. This is mainly due toimpingements and worse quality mixing [19].

17

Experimental Methodology

3.1 Hardware

The evaluation of the two injection systems is doneby performing engine tests at the internal combus-tion engine laboratory at KTH. A single cylinder en-gine based on the Scania D12 heavy duty engine ismodified to run both injection systems. The enginespecifications is shown in Table 3.1.

Compression ratio 13.1Bore [mm] 127Stroke [mm] 154Connecting rod length [mm] 255Displacement volume [dm3] 1.95Exhaust valves open [aTDCf] 145Exhaust valves close [aTDCf] 355Intake valves open [aTDCf] 346Intake valves close [aTDCf] 566

Table 3.1: Engine specifications

A compressor supplies the engine with air at adesired pressure. A throttle fitted in the intakemakes the engine capable of both over-boost andthrottled operations. The emissions were measuredusing a Horiba EXSA-1500 exhaust gas analyzer. Itwas calibrated before each test day to assure correctmeasurements.

The engine uses a slightly modified diesel cylinderheader. The ignition system consists of a VW115Rignition module and a Bosch sparkplug which isplaced in a widened pressure sensor channel ataround 3/4 radius of the bore. A custom pistonwith minimal squish was used in order to avoidconflict between the piston and the spark plug at

TDC.

The PFI system used a BOSCH 909 in-line fuelpump supplying a rail equipped with two injectors,one for each intake port. The rail pressure wasregulated at 3.5 bars by an adjustable pressureregulator fitted after the rail.

Scanias XPI system was used as the DI systemwhere an electrically driven high pressure fuelpump supplies a common rail which in turn sup-plies the XPI injector. This system is originallybuilt for diesel engines and slight modifications tothe injector was required in order to run E85. Theinjector is detailed in Table 3.2.

No of holes 6Flow [PPH] 235Umbrella angle [deg] 50

Table 3.2: Injector specifications

Ethanol E85 of Swedish standard SS 15 54 80 wasused for all tests.

3.2 Test plan

The test plan presented in this Section is designed toanswer the latter two questions detailed in Section1.2.

The injection systems were evaluated with the4 operating points shown in Figure 3.1. All pointswere run with MBT ignition timing (if not knocklimited), λ = 1 and the rail pressure at 1000 bar.

18

The thesis methodology is presented in this section. How the hardware was used and how it wasbuilt along with a detailed description of the performed tests is discussed.

Optimum SOIs were found by running the sweepsdiscussed below.

1,000 1,300

8.25

16.5

Engine speed

BM

EP

Figure 3.1: The 4 points chosen forevaluation

The optimum injection angles with respect to emis-sions were found by performing a Start Of Injection(SOI) sweep. The load during these sweeps waskept constant at 11 bar bmep (full throttle) withλ = 1 and MBT ignition. The PFI system was sweptaround the whole cycle, 720

◦CA, with 60◦CA

intervals. A second sweep with shorter intervals of10

◦CA were then performed in the window whereemissions had their minimum. The chosen SOI wasthen used for all further evaluation. The DI injectionsystem was swept for the compression and intakestroke in order to find two SOIs, one stratifiedand one homogeneous. The stratified sweep wasperformed from 20 to 90

◦CA bTDC fire, with 10

◦CA intervals. The interval was then increased to 30

◦CA from 90 to 330 bTDC fire ◦CA in order to findthe optimum homogeneous SOI angle.

In order to investigate how rail pressure effectsDI operation the 4 load points were repeated at 1600

bar.

All pressure traces and its offsprings are aver-ages of 100 cycles. All other measurements areaverages of a 1 minute measurement.

3.3 Calculation and evalua-

tion

3.3.1 Heat release rate

The heat release rate is calculated from the pressuretraces with Equation 3.1.

dQdθ

=γ

γ − 1p

dVdθ

+1

γ − 1V

dpdθ

(3.1)

where

γ =γair + γexh

2=

12·( cpair

cvair+

cpexhcvexh

)(3.2)

Crevice effects and heat transfer to the walls are ne-glected. The pressure trace is filtered by a first or-der low-pass filter. Pressure and volume derivativesare calculated by a numerical difference calculationshown in Eq. 3.3.

dxdθ

=x2 − x1

θ2 − θ1(3.3)

3.3.2 Mass fraction burned

A Wiebe function is fitted to each heat release tracein order to study the mass fraction burned of thedifferent injection systems. The Wiebe function isshown in Equation 3.4.

xb =

1 − exp(−α

(θ−θsθe−θs

β+1))

1 − exp(−α)(3.4)

where

θ = Crankangle

θs = Start of combustion

θe = End of combustion

and α and β are adjustable constant used to fit thecurve. These parameters are found by fitting thederivative of Equation 3.4

dxbdθ

=α(

θ−θsθe−θs

)β

θe − θs

1 + β

1 − exp(−α)exp

(−α

(θ − θs

θe − θs

)β+1)

(3.5)

19

to the heat release calculated from the measuredpressure curves. Figure 3.2 shows an example ofa fitted Wiebe function and a measured heat releasetrace.

−20 0 20 40−50

0

50

100

150

200

Crank angle

Hea

trel

ease

rat

e [J

/CA

]

dQ dxb/dθ

Figure 3.2: Example of fitted Wiebe functionto measured heat release

Table 3.3 shows all the values of α and β for the loadpoints discussed in Section 3.2. S, H and P standsfor Stratified, Homogenous and Port injection.

α β

8.251000 rpm

S 3.5 5H 4.5 2.5P 3.75 3

8.251300 rpm

S 2.9 5.25H 4.8 2.25P 4.75 2.5

16.51000 rpm

S ∼ ∼H ∼ ∼P 3.5 3.2

16.51300 rpm

S 2.9 5.25H 4.5 2.55P 4.5 2.75

Table 3.3: α and β fitted for all load points

3.3.3 Efficiency and fuel consump-

tion

The fuel flow could not be measured during the ex-periments and is instead calculated by

B =mair

AFR · λ(3.6)

where AFR = 9.675 (AFR = 15%AFRethanol +

85%AFRgasoline) and λ = 1 for all operation points.mair is measured with a rotary piston flow meter.

The efficiency can then be calculated by

η =2πn60 · T

B · LHV(3.7)

where T is the torque and LHV = 29.2 MJ/KG forE85.

3.3.4 Volumetric efficiency

The volumetric efficiency is calculated by dividingthe flow of air into the cylinder and the amount ofair displaced by the piston according to Equation3.8.

ηvol =2mairρVdN

(3.8)

Where ρ is the density of inducted air calculated bythe ideal gas law

ρ =mV

=Pin

RTin(3.9)

R = 287.058 is the specific gas constant for dry airand Vd is the displacement volume listed in Table3.1.

20

Results

4.1 Start of injection sweeps

4.1.1 PFI

In Figure 4.1 the emissions of HC, CO and NOxversus SOI are shown.

In Figure 4.1a it is obvious that the lowestHC emissions appear during the overlap of thevalve openings. Namely between 300 and 420

degrees aTDCnf . The HC are lowest at 346 wherethe intake valve opens. When injecting at the sametime the intake valve opens, the air and fuel aremixing and inserted inside the chamber reducingthe amount of time the fuel comes in contact withthe walls. As the SOI chosen are moving closer tothe time where the intake valve closes (514), the HCare increasing. This since fuel is still injected afterthe valve has closed and it stays longer period oftime inside the intake port, until the next time itthe intake valve opens. The measurement at SOI0◦CA and SOI 720

◦CA should be equall since itis the same point. However these two points donot coincide, probably because the time the enginewas allowed to run in order to stabilize was notenough. Thus there seems to be an incorrect readingregarding the SOI at 0

◦CA.

0 120 240 360 480 600 720

800

1000

1200

1400

SOI [aTDCf]

EXH INT

HC

[PP

M]

(a) HC

0 120 240 360 480 600 7201500

2000

2500

3000

SOI [aTDCf]

EXH INT

CO

[PP

M]

(b) CO

0 120 240 360 480 600 7203300

3400

3500

3600

3700

3800

3900

4000

SOI [aTDCf]

EXH INT

NO

x [P

PM

]

(c) NOx

Figure 4.1: Emission measurements fromthe PFI SOI sweep at 11 bars BMEP, λ = 1

and MBT. Tinj ≈ 18 CA

21

The results from the tests described in Chapter 3 is presented in this section.

Figure 4.1b shows the CO emissions from the PFISOI sweep. The measurement on the exhaust sideof aTDCnf is unstable and it is difficult to draw anyclear conclusions. It does appear however that aminimum exists somewhere around valve overlap.

When it comes to the NOx emissions (See Fig-ure 4.1c) the dominating mechanism is quitestraightforward and closely related to the in-chamber temperature. From the 0 to 360 degrees thetemperature is higher compared to the SOI between360 and 600. The explanation for this is that for thefirst range of SOIs, the intake valve is closed and thefuel therefore left for a long period of time insidethe intake channel. Thus the vaporization heat istaken from the walls instead of the charge leadingto a hotter mixture. In the SOI cases where the inletvalve is open, fresh and cool air is being introducedinto the chamber. Consequently, the amount of heattaken from the air for the fuel to vaporize increases.For SOIs over 600 degrees, the inlet valve is alsoclosed and thus the temperature rises again to thesame level as 0 to 360 showed.

Figure 4.2 shows SOI at smaller steps aroundthe area where the emissions were at their mini-mum. Naturally the same principles apply here.From these figures the optimum SOI was found tobe 340

◦CA which gave the best result in terms ofemissions.

320 330 340 350 360 370 380 390 400 410 420 430800

850

900

950

1000

SOI [aTDCf]

HC

[PP

M]

(a) HC

320 330 340 350 360 370 380 390 400 410 420 4302200

2400

2600

2800

3000

3200

SOI [aTDCf]

CO

[PP

M]

(b) CO

320 330 340 350 360 370 380 390 400 410 420 4303550

3600

3650

3700

3750

3800

SOI [aTDCf]

NO

x [P

PM

]

(c) NOx

Figure 4.2: Emission measurements fromthe narrower PFI SOI sweep at 11 barsBMEP, λ = 1 and MBT. Tinj ≈ 18 CA

22

4.1.2 DI

Emission measurments and ignition angle vs. SOIsweep is shown for DI in Figure 4.3. The sweepstarts at 20

◦ and ends at 330◦ bTDC fire. The

mixture is initially stratified and gradually becomeshomogeneous as the SOI moves from the compres-sion to the intake stroke.

Figure 4.3a shows the HC emissions. Both endsof the sweep appear to have low HC emissions.SOIs in the beginning of the intake stroke (330

bTDC) leads to homogeneous mixture. The HCemissions are around 1000 ppm. In this case thedominant source of HC emissions are crevicesand wall wetting effects. As the SOI moves intothe compression stroke between 150 and 20 themixture becomes more stratified (between 150 to20) and consequently increasingly locally rich andthe HC emissions is therefore seen to increase. HCmeasurements are out of range for SOIs during thecompression stroke, more specificlly between 70

and 155 bTDC ◦CA.

The same trend does not appear for SOIs dur-ing the compression (180-360 bTDC). This is partlydue to the piston’s location where if it is close toBDC the injected fuel ends up on the liner. Thischanges as the piston approches TDC. However,while a part of HC emissions can be explainedby the piston’s location it cannot explain theasymmetry between the two strokes (intake andcompression). One speculation for partly findinganswer to why this asymmetry exists could be thepiston velocity direction. When the piston travelsupwards, the mixture is pushed to the sides andends up at the liner. It only starts to decline forSOIs less than 60

◦CA degrees. Namely after thetop speed of the piston which in general is around75

◦CA degrees.

0 60 120 180 240 300 3600

1000

2000

3000

4000

SOI [bTDCf]

HC

[PP

M]

(a) HC

0 60 120 180 240 300 3602000

4000

6000

8000

10000

12000

SOI [bTDCf]

CO

[PP

M]

(b) CO

0 60 120 180 240 300 3602500

3000

3500

4000

4500

5000

5500

SOI [bTDCf]

NO

x [P

PM

]

(c) NOx

0 60 120 180 240 300 360−10

−5

0

5

10

15

SOI [bTDCf]

Mixinglimited

Knock limited

Igni

tion

angl

e

(d) Ignition angle

Figure 4.3: Emission measurements from DISOI sweep at 11 bars BMEP and λ = 1.

Tinj ≈ 2.5 CA

23

On the other side, when the piston travels down-wards, the fuel is drawn downwards. This helpsreduce the amount of fuel ending up on the walls.The lowest emissions during the intake strokeappear to be when SOI is close to TDC. At thatpoint the piston is closest to TDC and it willstart to travel downwards with increasing speed.Therefore, forces affecting the fuel are greater andare applied for a longer time compared to when thepiston has already travelled closer to BDC. Furtheranalysis and tests would needed to support such ahypothesis.

The crevice and wall wetting effect is negligi-ble during stratified combustion. The emissions ofHC is due to flame quenching of the lean peripheralof the fuel cloud. HC is as low as 250 PPM forhighly stratified combustion at 20 to 60

◦CA.

The CO emissions are shown in Figure 4.3b.CO emissions at the beginning of the sweep, wherethe mixture is stratified and rich, appears to belower than at the other end where it is homoge-neous. This indicates that the early SOIs causes"pools" on the piston which leads to diffusionsflames. It is unclear whether or not the spike at 60

◦CA bTDC is a measurement error.

It is well known that NOx production is highlydependent on temperature and λ-value. Duringthese sweeps the temperature is fairly constantinside the cylinder. A different explanation is thusneeded. The mixture is locally rich during strat-ified operation and there is therefore less oxygenavailable for the NOx to be created. The mixture isstoichiometric during homogeneous operation andNOx production increases slightly as it is knownthat NOx peaks at a slightly lean A/F. The sparktiming has to be advanced as the mixture becomesmore homogeneous in order to avoid knock. Thislowers the temperature and explains the peak at120

◦ bTDC, just before spark retard is needed, andthe decreasing NOx for earlier SOIs.

4.2 DI vs PFI

The two injection systems are compared with the op-timum SOIs found in Section 4.1. Figure 4.4 showsthe pressure traces and Figure 4.5 shows the heat re-lease traces of the load points detailed in Table 4.1.

Inja

ngle

[CAbT

DCf]

Injd

ur1[m

s]

Injd

ur2[m

s]

Ignangle[CAbT

DCf]

Pin

[bar]

8.25 BMEP1000 rpm

S 40 2.01 ∼ 0* -0.21H 330 1.9 ∼ 12 -0.2P 340 14.2 14.1 5 -0.2

8.25 BMEP1300 rpm

S 40 2.1 ∼ 3* -0.225H 330 1.81 ∼ 14* -0.27P 340 14.6 14.5 8 -0.23

16.5 BMEP1000 rpm

S ∼ ∼ ∼ ∼ ∼H ∼ ∼ ∼ ∼ ∼P 340 27.5 27.5 0 0.4

16.5 BMEP1300 rpm

S 40 2.71 ∼ 0* 0.3H 330 3.6 ∼ 8 0.27P 340 27.8 27.5 6 0.33

Table 4.1: Operational settings for the 4 loadpoints. *MBT

The DI system proved highly resistant to knockand MBT ignition was possible for most load pointswhen running at stratified conditions. It is seen fromfigure 4.3d, that for the homogeneous strategy, theprocess is knock limited and therefore the ignitionangle remains more or less constant. As the SOIare close to TDC at stratified conditions, the MBTis achieved. However, it is very interesting to ob-serve that there is a very steep drop as the SOI anglechosen is less than 60 bTDCf and on. In that case, inthe effort to explain this phenomenon a new conceptis introduced as hypothesis. This hypothesis claimsthat, at that point, the limited parameter is neitherMBT nor knock, but instead the time available forthe fuel and air to be mixed in a level that the mix-ture is combustible. The combustion itself can be

24

realized, even at that extremely late SOI, due to theintense combustion of the E85 (as it will be shownlater on). MBT was, as expected, not possible for allload points at homogeneous conditions or any of theload points for the PFI arrangement.

4.2.1 Cylinder pressure

Figure 4.4 compares the pressure traces. It is clearthat stratified DI results in a steep pressure rise asa result of its fast combustion (see Section 3.3.2). Itpeaks sooner than both homogeneous DI and PFIdespite being ignited later. Homogeneous DI is seento have a faster combustion and proved less prone toknock than PFI, leading to an ignition timing closerto MBT. Maximum cylinder pressure reach over 60

bar at medium load for stratified DI and is around10 and 20 bar lower for homogeneous DI and PFIrespectively. At high load and speed the stratifiedDI peak at 120 bar while both homogeneous DI andPFI peak at 80 bar.

It thus seems that the advantages of stratifiedDI compared to PFI are greater at high load andspeed. The opposite seems to be true for homoge-neous DI where the advantages compared to PFIare found at low load. Figure 4.4d shows that thereis little difference between homogeneous DI and PFIat high load and speed. The knock resistance forhomogeneous DI is reduced and the ignition had tobe retarded to the same level as PFI’s which resultsin a similar pressure trace.

Table 4.2 shows the maximum positive gradi-ent for the pressure traces in BARS/CA.

As seen in the pressure traces there is littleseparating homogeneous DI and PFI. The differenceis at most 1 BARS/CA and again it is stratifiedDI that stands out with a value as high as 14.2BARS/CA at high load and speed. This high valuemight cause problems with both noise and materialstress.

−45 0 45 900

20

40

60

80

Crank angle

Cyl

inde

r pr

essu

re [B

ar]

STRAT HOM PFI

(a) BMEP = 8.25 @ 1000 rpm

−45 0 45 900

20

40

60

80

Crank angle

Cyl

inde

r pr

essu

re [B

ar]

(b) BMEP = 8.25 @ 1300 rpm

−45 0 45 900

20

40

60

80

100

120

Crank angle

Cyl

inde

r pr

essu

re [b

ar]

(c) BMEP = 16.5 @ 1000 rpm

−45 0 45 900

20

40

60

80

100

120

Crank angle

Cyl

inde

r pr

essu

re [B

ar]

(d) BMEP = 16.5 @ 1300 rpm

Figure 4.4: Pressure traces for stratified DI,homogeneous DI and PFI for 4 operating

points

25

BARS/CA

8.25 BMEP1000 rpm

S 5.9H 2.0P 1.6

8.25 BMEP1300 rpm

S 7.1H 2.3P 1.7

16.5 BMEP1000 rpm

S ∼H ∼P 3.1

16.5 BMEP1300 rpm

S 14.2H 3.3P 4.3

Table 4.2: Pressure rise in BARS/CA for theinjection concepts during the 4 load points


Figure 4.5 shows the heat release traces calculatedwith Eq. 3.1 in Section 3.3.1.

The fast combustion of stratified DI is also evi-dent in the heat release rate traces. The heat releaserate increases faster and peaks at a greater valuethan the other concepts. A negative heat releaserate can be seen at the time of injection (-40

◦CA)since the fuel extracts heat from the air in order tovaporize.

The heat release rate from homogeneous DIand PFI is similar. The slightly more advancedignition timing, due to knock insensitivity, of thehomogeneous DI is evident at low load in Figure4.5a and 4.5b where the heat release rate trace isequally advanced.

As discussed in Section 4.2.1 there was littleseparating the homogeneous DI and PFI at highload and speed (Figure 4.5d) since both whereknock limited.

−45 0 45 90−100

0

100

200

300

400

Crank angle

Hea

trel

ease

rat

e [J

/CA

]

STRAT HOM PFI

(a) BMEP = 8.25 @ 1000 rpm

−45 0 45 90−100

0

100

200

300

400

Crank angle

Hea

trel

ease

rat

e [J

/CA

](b) BMEP = 8.25 @ 1300 rpm

−45 0 45 90−200

0

200

400

600

800

Crank angle

Hea

trel

ease

rat

e [J

/CA

]

(c) BMEP = 16.5 @ 1000 rpm

−45 0 45 90−200

0

200

400

600

800

Crank angle

Hea

trel

ease

rat

e [J

/CA

]

(d) BMEP = 16.5 @ 1300 rpm

Figure 4.5: Heatrelease rate traces forstratified DI, homogeneous DI and PFI for 4

operating points

26

4.2.3 Mass fraction burned

The mass fraction burned, xb (See Section 3.3.2), forthe 4 load points and different injection systems isshown in Figure 4.6. The Wiebe function proved dif-ficult to fit to the sharp heat release trace of strati-fied DI, especially at the ends of the trace. This leadsto both that some mass fraction burned traces con-verge above 100% and that some values on DUR5

and DUR10 are before or at ignition. This is obvi-ously impossible. Table 4.3 shows 5, 10, 50 and 90%heat released in relation to crank angle.

DUR5 DUR10 DUR50 DUR90 IGN

8.251000

S -0.6 0.9 6 9.9 0H 9.3 12.1 22.7 32.9 -12P 5.3 8.4 19.3 29 -5

8.251300

S 2 3.5 8.1 11.5 -3H 9.1 11.9 22.7 33.4 -14P 9.1 11.8 22.3 32.5 -8

16.51000

S ∼ ∼ ∼ ∼ ∼H ∼ ∼ ∼ ∼ ∼P 3.9 6.7 16.6 25.2 0

16.51300

S 0 1.5 6.1 9.5 0H 8.9 11.6 21.7 31.4 -8P 8.4 11.2 21.7 31.4 -6

Table 4.3: CA duration from IGN until 5, 10,50 and 90% burned. Ignition (IGN) in ◦CA

bTDCf.

Stratified DI leads to the fastest combustion fol-lowed by homogeneous DI and lastly PFI for all loadpoints. The stratified DI combustion is significantlyfaster than the other concepts and reached 90%burned around 20 CA faster. The explanation forthis is, as discussed in Section 2.3.4, the increasedturbulence gained from the impulse of the injection.

Homogeneous DI and PFI follow each otherclosely.

−20 0 20 400

0.5

1

1.5

Crank angle

Mas

s fr

actio

n bu

rned

[%]

STRAT HOM PFI

(a) BMEP = 8.25 @ 1000 rpm

−20 0 20 400

0.5

1

1.5

Crank angle

Mas

s fr

actio

n bu

rned

[%]

(b) BMEP = 8.25 @ 1300 rpm

−20 0 20 400

0.5

1

1.5

Crank angle

Mas

s fr

actio

n bu

rned

[%]

(c) BMEP = 16.5 @ 1000 rpm

−20 0 20 400

0.5

1

1.5

Crank angle

Mas

s fr

actio

n bu

rned

[%]

(d) BMEP = 16.5 @ 1300 rpm

Figure 4.6: Mass fraction burned for all loadpoints

27

Combustion speed remains fairly constant with DIand increased load but increases, especially earlyflame development, with increased engine speed.This is due to the increased turbulence and chargemotion velocity at higher engine speeds.

4.2.4 Emissions

Table 4.4 shows HC, CO and NOx emissions for thethree different injection systems.

HC CO NOx

8.25 BMEP1000 rpm

S 643 2864 3660H 1066 6281 3412P 1122 3108 3669

8.25 BMEP1300 rpm

S 377 2701 3930H 1118 6245 3977P 1007 2965 4119

16.5 BMEP1000 rpm

S ∼ ∼ ∼H ∼ ∼ ∼P 1611 2085 4025

16.51300 rpm

S 183 4651 3100H 520 5894 3684P 2284 2341 4832

Table 4.4: Emissions comparison of theinjection concepts at 4 load points. All values

in PPM

PFI consistently has the highest HC output ofthe three concepts. Mainly due to wetting of theintake port and channel during injection as well aspremixed charge entering the crevices. HC increasesat higher load for PFI simply because there is moreair and fuel available. The stratification helps theDI system by completing the combustion before itreaches the walls and quenches, especially in thecase of the stratified strategy. There is not muchdifference in HC for homogeneous DI and increasedspeed. For stratified DI however, the HC emissionsare halved due to the improved turbulence andin-cylinder motions helping to homogenize thecharge as well as improved oxidation due to theslightly faster combustion. As load increases theHC decreases, again by about half, for both DIstrategies due to improved after-oxidization due to

higher exhaust temperatures.

CO emissions are around 3000 PPM for PFI atlow load and drops to around 2300 PPM for highload. Stratified DI produces surprisingly less CO,at low load, than both PFI and homogeneousDI. The stratified combustion is rich and shouldproduce high CO due to the local lack of oxygenand consequently less oxidization of CO and CO2.The low CO emissions are probably explaned byhigh turbulence of the end of combustion created bylate SOI. The retarded combustion of homogeneousDI ends much later in the expansion stroke whichleaves it with less time, a lower temperature andless turbulence to oxidize into CO2. A second ex-planation might be that the homogeneous injection,which occurs with the piston closer at TDC andlower cylinder pressure, hits the piston and creates"puddles" which causes diffusion flames whichincrease the CO output of the homogeneous DIto levels higher than the stratified. PFI producesless than homogeneous DI even though these twoconcepts are similar. It might be an indicator thatthe homogeneous DI can’t produce the same levelof homogeneity as the PFI.

The NOx level is in between 3500 and around4000 PPM for all injection concepts with the ex-ception of PFI at high load and speed where it iseven higher. During high load and speed there isa decline in NOx emissions for the stratified DIand the combustion seems to be is rich (confirmedby the increase in CO). NOx emissions are higherwith PFI than DI since the charge is warmer due tovaporization heat taken from intake ports. This ismost evident at high load and speed where peakcylinder pressure is higher. The result is NOxemissions up to 4800 PPM

28

4.2.5 Efficiency

Table 4.5 shows the results from the efficiency calcu-lation presented in Section 3.3.3

eta

8.25 BMEP1000 rpm

S 0.405H 0.408P 0.400

8.25 BMEP1300 rpm

S 0.396H 0.397P 0.391

16.5 BMEP1000 rpm

S ∼H ∼P 0.356

16.5 BMEP1300 rpm

S 0.365H 0.367P 0.364

Table 4.5: Efficiency, η, of the injectionconcepts at the four load points

As noted, the stratified DI offers significantly lowerHC emissions, lower CoV IMEP and a very fast com-bustion. While everything indicate higher total ef-ficiency when running stratified DI, this is not ap-parent in the actual results. Table 4.5 shows the re-sulting efficiencies for all load points and concepts.According to the results there is no significant dif-ference among them. It is therefore logical to as-sume that a possible error in the air measurementreadings has occurred. Consequently, the fuel mea-surement was also wrongly calculated and the finaltotal efficiency does not reflect the obvious benefitsof stratified DI.

4.2.6 Coefficient of Variation in

IMEP

Table 4.6 shows the Coefficient of variation (COV) ofIMEP for the injection concepts during the four loadpoints.

CoV in IMEP [%]

8.25 BMEP1000 rpm

S 1.04H 1.88P 2.12

8.25 BMEP1300 rpm

S 1.05H 1.42P 2.72

16.5 BMEP1000 rpm

S ∼H ∼P 1.09

16.5 BMEP1300 rpm

S 0.61H 0.98P 1.45

Table 4.6: CoV in IMEP for the three injectionsystems at the four load points

CoV IMEP is consistently lowest for stratified DI.The difference is greatest at low load and speed anddecreases as load and speed increases.

The stratified DI is more stable due to the increasedturbulence which leads to a faster combustionwhich in turn is more predictable.

The sidemounted spark plug means that the flamehas to propagate from one side of the chamber tothe other. This increases the unpredictability of itand therefore also the CoV. This is more significantwith the "pre-mixed" concepts homogeneous DI andPFI since the mixture is more spread throughoutthe chamber. It is also more likely that the injectionevent, i.e. flow and pressure, is more predictablewith the DIs high pressure XPI system than withthe PFI system. These two combined should explainthe higher CoV for the "pre-mixed" concepts.


The volumetric efficiencies of the different injectionconcepts are shown in Table 4.7

29

ηvol

8.25 BMEP1000 rpm

S 0.752H 0.764P 0.741

8.25 BMEP1300 rpm

S 0.813H 0.823P 0.806

16.5 BMEP1000 rpm

S ∼H ∼P 0.965

16.5 BMEP1300 rpm

S 0.991H 1.017P 1.000

Table 4.7: Volumetric efficiency, ηvol, for thethree injection concepts at four load points

As expected (See Section 2.3.3) homogeneous DIhas an advantage in terms of volumetric efficiency.The fuel is injected while the intake valve is openand as it extracts heat from the air in order tovaporize, the density increases and more air canbe induced into the cylinder. The same principleis true for PFI but the volumetric efficiency isbetween 2 and 3% lower since the heat is takenfrom the walls of the intake system instead of the air.

The fuel injection does not affect the air induc-tion process with stratified DI since the injectionoccurs when the intake valve is closed. It doeshowever have a higher volumetric efficiency (1%)compared to PFI at low load which would indicatethat PFI actively deteriorates volumetric efficiency.This might be due to the lower molecular weight ofethanol which increases its volume, it is thereforeconceivable that PFI would have a more similarvolumetric efficiency, compared to stratified DI,when running gasoline.

4.3 Rail pressure

The effect of different injection pressure was inves-tigated. Except the standard injection pressure of1000 bars used for the comparison, the load pointswere repeated at 1600 bars. Table 4.8 shows the op-erational settings for the rail pressure investigations.

Inja

ngle

[CAbT

DCf]

Injd

ur[m

s]

Ignangle[CAbT

DCf]

Pin

[bar]

8.25 BMEP1000 rpm

1600 S 40 1.51 4* -0.21H 330 1.33 14* -0.2

1000 S 40 2.01 0* -0.21H 330 1.9 12 -0.2

8.25 BMEP1300 rpm

1600 S 40 1.57 4* -0.23H 330 1.27 14* -0.27

1000 S 40 2.1 3* -0.26H 330 1.81 14* -0.27

16.5 BMEP1300 rpm

1600 S 40 1.81 0* 0.3H 330 2.73 8 0.27

1000 S 40 2.71 0* 0.3H 330 3.6 8 0.27

Table 4.8: Operational setting for the 4 loadpoints and 2 different railpressures. *MBT.

Injection duration in ms.

Rail pressure seems to have an effect on knock ashomogeneous DI could be run at MBT for low loadand low speed at 1600 bars of injection pressure butnot at 1000 bars.

It is important to note that, when running stratified,the ignition angle at low load and speed are 4 and 0

◦CA for 1600 and 1000 bar respectively. Both theseangles are MBT but unfortunately this causes somedifficulty when directly comparing them as in thepressure and heat release traces. It also seems thatthe air flow measurement is implausibly low forthe point with 1600 bars. This leads to a too highefficiency and a too low volumetric efficiency.

30

4.3.1 Cylinder pressure

Figure 4.7 shows how the cylinder pressure changesfor the different load points at the two differentinjection pressures.

Figure 4.7a clearly shows the different ignitionangles (discussed in Section 4.3) with S16s higherand more advanced peak. At low load and highspeed, where the ignition angles are almost equal,there is little separating S16 and S10. At high loadand speed on the other hand (See Figure 4.7c) bothS16 and S10 are ignited at the same time. Here,S16 is seen to have a faster combustion and peakslightly higher and sooner than S10. This is due tothe fact that a higher injection pressure at the samehole diameter causes a higher degree of atomizationof the fuel which consequently leads to a fastervaporization.

What can be observed directly is that for thelow load and low speed the H16 is a bit highercompared to H10 while the H10 is slightly higherat higher engine speed (see Figure 4.7b). The latteris also true when both the engine speed and loadincreases (see Figure 4.7c).

−45 0 45 900

20

40

60

80

Crank angle

Cyl

inde

r pr

essu

re [B

ar]

S16S10H16H10

(a) BMEP = 8.25 @ 1000 rpm

−45 0 45 900

20

40

60

80

Crank angleC

ylin

der

pres

sure

[Bar

]

(b) BMEP = 8.25 @ 1300 rpm

−45 0 45 900

50

100

150

Crank angle

Cyl

inde

r pr

essu

re [B

ar]

(c) BMEP = 16.5 @ 1300 rpm

Figure 4.7: Cylinder pressure comparison ofstratified and homogeneous DI for 1600 and

1000 bars for 3 load point

31


Since the heat release rate is closely related to the incylinder pressure similar trend appear in the heatrelease diagram (see Figure 4.8). The resulting heatrelease rate curves are credible and they seem tofollow the theory.

In Figure 4.8a the heat release for the stratifiedconditions starts earlier for 1600 bar compared to1000 bar of injection pressure. This is due to theearlier ignition angle discussed in Section 4.3.

At low speed and low load the slightly bettercharge cooling effect utilization can be seen for thehigher injection pressure as a larger negative heatrelease.

As mentioned before, higher injection pressureresults in higher level of atomization of the injectedfuel which in turn results to a faster vaporizationand therefore the advanced heat release seen in Fig-ure 4.8c. On the other hand, for the homogeneousstrategy, the differences are negligible. Since thepressure traces are almost identical the heat releaserate follows also the same trend. The early injectionof the fuel (beginning of intake stroke) even athigher injection pressure leaves enough time for theincrease in turbulence due to a higher rail pressureto be diminished.

−45 0 45 90−100

0

100

200

300

400

Crank angle

Hea

trel

ease

rat

e [J

/CA

]

S16S10H16H10

(a) BMEP = 8.25 @ 1000 rpm

−45 0 45 90−100

0

100

200

300

400

Crank angleH

eatr

elea

se r

ate

[J/C

A]

(b) BMEP = 8.25 @ 1300 rpm

−45 0 45 90−200

0

200

400

600

800

Crank angle

Hea

trel

ease

rat

e [J

/CA

]

(c) BMEP = 16.5 @ 1300 rpm

Figure 4.8: Heatrelease rate comparison ofstratified and homogeneous DI at 1600 and

1000 bars for 3 load point

32

4.3.3 Emissions

Table 4.3.3 compares the HC, CO and NOx emis-sions for an increasing rail pressure.

HC CO NOx

8.25 BMEP1000 rpm

1600 Bars S 789 4001 3632H 1036 5254 3537

1000 Bars S 643 2864 3660H 1066 6281 3412

8.25 BMEP1300 rpm

1600 Bars S 496 4362 3744H 1111 5053 4125

1000 Bars S 377 2701 3930H 1118 6245 3977

16.5 BMEP1300 rpm

1600 Bars S 259 4561 3004H 609 5759 3634

1000 Bars S 183 4651 3100H 520 5894 3684

Table 4.9: HC, CO and NOx emissions forhomogeneous and stratified DI at two rail

pressures and three load points. All values arein PPM.

It is clear that for an increase of the injectionpressure from 1000 bar to 1600 bar the HC areincreased for stratified conditions while for thehomogeneous strategy they do not change. Forthe homogeneous strategy case the explanation israther straightforward and that is the injected fuelhas enough time to be mixed with the air and endup in the same mixing state when combustion startsirrespectively of injection pressure. For stratifiedconditions this is not the case. The increasedinjection pressure changes the way the fuel is beingdistributed inside the combustion chamber. As thishappens close to TDC there is not enough timefor the mixture to end up in the same conditionsirrespective of the injection pressure as in the casefor the homogeneous case.

For stratified DI there is a large increase in COemissions at low load, but not at high load, asinjection pressure increases. This is most likelycaused by an increase in stratification and thuslocally richer combustion. The exact reason behindthis is unclear. It might be due to the increase

in penetration resulting in different flow patterninside the chamber causing a locally richer regionin the center. For homogeneous conditions, the COemissions for low load are lowered with increasedinjection pressure. Possibly due to the improvedatomization causing a more well mixed chargeand therefore a more complete combustion. COremains constant at full load for both stratified andhomogeneous DI with increased injection pressure.

The NOx emissions remains constant with in-creased injection pressure for both homogeneousand stratified DI. The improved fuel vaporization(seen in Figure 4.7a in Section 4.3.1) should lead toa cooler charge and lower NOx but for some reasonthere is no evidence of that.

4.3.4 Efficiency

Table 4.10 shows the calculated efficiencies. The val-ues with 1000 bars rail pressure are repeated fromTable 4.5 in Section 4.2.5 for easier comparison.

η

8.25 BMEP1000 rpm

1600 Bars S 0.426H 0.406

1000 Bars S 0.405H 0.408

8.25 BMEP1300 rpm

1600 Bars S 0.409H 0.393

1000 Bars S 0.396H 0.397

16.5 BMEP1300 rpm

1600 Bars S 0.369H 0.368

1000 Bars S 0.365H 0.367

Table 4.10: Efficiency, η, for homogeneous andstratified DI at two rail pressures and three

load points

Different values of efficiencies had been expectedas already discussed in Section 4.2.5. However, inthis case, conclusions can be drawn since it is asystematic error that applies equally. The efficiencyof 0.426 for low load and speed at 1600 bar is notto be trusted due to the, possibly, incorrect air flow

33

reading discussed in the beginning of Section 4.3.

Rail pressure does not have an effect on homo-geneous DIs efficiency and any differences arewithin the measurement error. The fuel is injectedearly and any potential benefits of a higher railpressure, i.e. increased turbulence, is dissipatedduring the intake and compression strokes. A slightincrease in efficiency can be seen for stratified DIat low load and high speed where it increases with3.3% as it changes from 0.396 to 0.409. Sections 4.3.2and 4.3.3 indicates a faster and richer combustionwhich would also translate to a higher efficiency.

4.3.5 Coefficient of Variation in

IMEP

Table 4.11 shows the CoV IMEP for the three oper-ating points and two rail pressures. As with Table4.10 above the values for 1000 bars is repeated fromTable 4.6 in Section 4.2.6.

CoV in IMEP [%]

8.25 BMEP1000 rpm

1600 S 0.97H 1.61

1000 S 1.04H 1.88

8.25 BMEP1300 rpm

1600 S 1.03H 1.51

1000 S 1.05H 1.42

16.5 BMEP1300 rpm

1600 S 0.53H 0.95

1000 S 0.61H 0.98

Table 4.11: CoV in IMEP measured forhomogeneous and stratified DI at two rail

pressures and three load points

Increased rail pressure has a positive effect, regard-ing the COV IMEP, for both homogeneous and strat-ified DI at all load points. Homogeneous DI im-proves more at low load and speed, while stratifiedimproves more at high load and high speed.


Table 4.12 compares how the rail pressure affects thevolumetric efficiency.

ηvol

8.25 BMEP1000 rpm

1600 S 0.707H 0.763

1000 S 0.752H 0.764

8.25 BMEP1300 rpm

1600 S 0.794H 0.824

1000 S 0.813H 0.823

16.5 BMEP1300 rpm

1600 S 0.988H 1.021

1000 S 0.991H 1.017

Table 4.12: Volumetric efficiency, ηvol, forhomogeneous and stratified DI at two injection

pressures and three load points

As with the efficiency in Section 4.3.4 the volumetricefficiency of 0.707 for low load and speed at 1600

bar is incorrect. This is due to the fact that theintake valve is closed at that time. Thus, any changeof injection pressure could not possibly affect theamount of air inserted into the chamber. Thiscan only strengthen the speculation about havingincorrect air flow measurement.

No gain in volumetric efficiency can be seenwith homogeneous DI and increased rail pressureeven tough the atomization of the fuel improves.

34

Summary & Conclusions

All concepts; stratified DI, homogeneous DI and PFI,were able to work with E85 on the Scania D12 en-gine. Both delivered reliable and continuous opera-tion with the exception of DI, at high load and lowspeed, where it is likely that the ignition coil (orig-inating from a passenger car) was too weak to sup-port the combustion.

5.1 DI vs PFI

The different injection systems was compared with4 load points. The systems were evaluated in termsof consumption, emissions, combustion stabilityand load potential (knock resistance) (See Sections5.1.1 to 5.1.4).

A sweep was performed in order to find theoptimum SOIs for both PFI and DI. The sweep forthe whole 720

◦CA SOI range for PFI showed thatthere is an optimum window where all emissionshad a global minimum. Within this window therewas not much variation and 340

◦CA was chosenas optimum SOI for the PFI operation. The 360

◦CA sweep with DI showed that both HC andNOx were low at both ends of the sweep while COremained roughly constant. 40

◦CA where thuschosen as SOI for stratified operation and 330

◦CAas homogeneous operation.

5.1.1 Load potential

Stratified DI proved resilient to knock and MBT igni-tion was possible at most load points. It is believedthat the fast combustion combined with the locally

rich mixture is responsible for this. MBT could notbe run either with homogeneous DI or PFI, and igni-tion had to be retarded. Homogeneous DI was lesssensitive and could run with ignition closer to MBT,compared to PFI.

5.1.2 Combustion speed

From the pressure and heat release rate traces itis clear that the homogeneous DI and PFI producesimilar combustions. Stratified DI stands out withmuch sharper pressure rise due to faster combus-tion. The combustion was further investigated bystudying the mass fraction burned traces where itwas found that stratified DI reaches DUR90 around20 CA faster than the other concepts. PFI were foundto produce a faster flame than homogeneous DI.

5.1.3 emissions

Stratified DI has, in general, the lowest HC and COemissions compared to both homogeneous DI andPFI. There is not much separating these concepts interms of NOx.

5.1.4 Efficiency and combustion

stability

No conclusions can be drawn from the presented ef-ficiencies due to the lack of a reliable fuel flow mea-surement. Values for CoV IMEP were found to bewithin the acceptable range. Stratified DI showedthe best COV IMEP, as low as 0.6% and the PFI the

35

A Summary along with the main conclusions of the results presented in Section 4 is presented inthis section

worst, with a CoV IMEP of almost 3%. Homoge-neous DI’s CoV IMEP came in between, howevercloser to stratified DI. Homogeneous DI had, as ex-pected, the highest volumetric efficiency, followedby stratified DI in second place and PFI last.

5.2 DI

Time restricted further studies of the DI system andonly rail pressure was investigated.

5.2.1 Rail pressure

The rail pressure study suggests that an increase ininjection pressure leads to slightly faster combustiondue higher turbulence. It increases both HC and COemissions which indicates that a higher rail pressureleads to increased stratification due to increasedpenetration and thus a locally richer mixture whenrunning stratified. Efficiency increases with around3% and CoV IMEP decreases with higher railpressure.

Rail pressure had little or no effect on homo-geneous DI.

36

Discussions and future work

6.1 Discussion

The aim of this thesis was to investigate the poten-tial benefits and drawbacks of direct injection ona heavy duty ethanol internal combustion engineswith an experimental investigation using PFI as areference.

The results proved and confirmed many of thebelieved potentials of a such system. E.g. the muchfaster combustion as well as lower emissions andmore stable combustion. Most of the results arederived from cylinder pressure measurements anddeemed reliable. Same goes for the emissions whichwere measured with a high quality equipmentcalibrated on a daily basis. Unfortunately there wasno way to measure the fuel flow and values on effi-ciency had to be calculated from air flow and air fuelratio. These values are improbable and it is morelikely that the air flow measurements were incorrect.

One of the planned load points, namely BMEP16.5 at 1000 rpm proved impossible with DI. Thispoint had already been run with PFI and due tolack of time it was not possible to revert the engineback to PFI and choose a different load point knownto work with DI. This point is thus, sadly, left blankfor DI in the comparisons between the injectionsystems.

6.1.1 Restrictions of Hardware

and test bed

Some less than optimal solutions regarding theengines hardware were used due to restrictions inboth time and finance. The cylinder head used anexisting channel for the spark plug. This meantan off-center position of the spark plug which isless than ideal. Furthermore, the ignition systemcame from a passenger car due to availability andfinance. It is believed that a more powerful systemis required with this high in-cylinder numberdensity at the time of spark. Another drawbackof the off-center spark location is that the pistoncrown shape had to be designed primarily to avoidcolision with the spark plug, rather than producingthe desired air flow.

The PFI system used an existing inlet pipe withinjector mounting far upstream from the intakeport. This inlet pipe is originally built for gasinjection where the distance is not a problem. Forliquid fuels however this is far from optimal.

The engine test bed unveiled some weaknessesas testing went on. One large obstacle in these testswhere the highly unstable inlet pressure controlwhen running throttled, which made reliable andrepeatable measurements difficult. The originallyplanned load and speed load points therefore hadto focus more at full trottle in order to avoid this.The measurement time was lengthened to ensure amore correct average value.

37

This section discusses the results presented in Section 4 as well as their limitations along withrecomendations on subjects and hardware for future work

6.2 Future work

There is still a lot to investigate with this technology.This thesis originally set out to investigate a lot morethan is presented in this report but time ran out dueto problems with the equipment.

6.2.1 Test plan

Further studies are needed to find the exact reasonsfor some of the phenomena seen in the tests. Firstand foremost, a full comparison of both injectionsystems with new load/speed point, that bothsystems can handle. An optical investigation of theinjection and combustion would help to answermany of the question that still remain.

A proposal for future work and understandingwould be to run the same tests again, but that timeto calculate the efficiency for each of this points.This would presumably show that at very lateignitions where MBT is not achieved, the efficiencyis dropped dramatically. That would also provewhether the SOI chosen was the best choice or not.

The literature review identified many interest-ing questions to be studied. Unfortunately thetime available was not enough to investigate allof those questions. The most interesting were;using different umbrella angles to examine theeffects of in-cylinder motion, especially at stratifiedconditions. Split injections to combine the positivesfrom both homogeneous and stratified.

6.2.2 Hardware and test bed

The installation and control of the PFI system waseasy. The same can not be said for the XPI DIsystem where it seems that the long injector onduration overloaded the cells power supply andforced the control system (actuating the ignition)to momentarily shut off and subsequently shutoff the ignition. Several attempts to find out theexact cause, such as individual power sources, weredone but failed. One suggestion for future studiesis therefore to invest in an after market engine

management system, preferably one designed forOtto engines.

The laboratory did not support running theengine with fuel from a barrel inside the fuel depot.Instead the tests had to be run with a cannisterfrom inside the cell. This causes many problems. Asolution where a barrel could be connected to thefuel lines leading to the cell would increase safety,remove the need of refilling the cannister and leadito longer uninterrupted tests and more accuratefuel flow reading.

The ability to run throttled is necessary whenrunning an Otto engine. The control of thecompressor supplying the engine is out of thedepartemet’s control, and when problem discoveredit was difficult to solve it. If further Otto enginetests are to be performed this does however needto be fixed together with the parties responsible forthe compressor.

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Nomenclature

A/F Air-Fuel ratio

BDC Bottom Dead Center

CO Carbon Monoxide

CO2 Carbon Dioxide

CoV Coefficient of variation

CR Compression ratio

DI Direct Injection

EOI End of injection

H10 Homogenous DI at 1000 bar

H16 Homogenous DI at 1600 bar

HC Hydro carbons

HD Heavy duty

ICE Internal Combustion Engines

IMEP Indicated Mean Effective Pressure

L/D Length-Diameter ratio

LHV Lower heating value

MBT Maximum Brake Torque

NOx Mono Nitrogen Oxides

PFI Port Fuel Injection

PM Particulate matter

S10 Stratified DI at 1000 bar

S16 Stratified DI at 1600 bar

SI Spark ignited

SOI Start of injection

TDC Top Dead Center

TDC/BDCf Top/Bottom Dead Center fire

TDC/BDCnf Top/Bottom Dead Center non-fire

WOT Wide Open Throttle

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