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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
4.2.2 Heat release rate . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26
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
4.2.7 Volumetric efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29
4.3 Rail pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30
4.3.1 Cylinder pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31
4.3.2 Heat release rate . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32
4.3.3 Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
4.3.4 Efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
4.3.5 Coefficient of Variation in IMEP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34
4.3.6 Volumetric efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 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
4.2.2 Heat release rate
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.
4.2.7 Volumetric efficiency
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
4.3.2 Heat release rate
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.
4.3.6 Volumetric efficiency
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.
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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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References
[1] Dr. Markus Rarbach and Yvonne Söltl. Cellu-losic ethanol from agricultural residues. MTZworldwide Edition: 2013-04, 2013.
[2] Dr.-Ing. Markus Schwaderlapp, Dr.-Ing. PhilippAdomeit, Dipl.-Ing. Andreas Kolbeck, andDipl.-Ing. Matthias Thewesl. Ethanol and itspotential for downsized engine concepts. MTZworldwide Edition: 2012-02, 2012.
[3] RFA: Renewable fuels association. World fuelethanol production, Downloaded: 2014-03-05. http://www.fermentas.com/techinfo /nu-cleicacids/maplambda.htm.
[4] U.S Energy Information Administra-tion. International energy outlook 2013.http://www.eia.gov/forecasts/ieo/.
[5] J.E. Anderson, D.M. DiCicco, J.M. Gin-der, U. Kramer, T.G. Leone, H.E. Raney-Pablo, and T.J. Wallington. High oc-tane number ethanol–gasoline blends:Quantifying the potential benefits in theunited states. Fuel, 97(0):585 – 594, 2012.dx.doi.org/10.1016/j.fuel.2012.03.017.
[6] Thomas Wallner and Scott A. Miers. Combus-tion behavior of gasoline and gasoline/ethanolblends in a modern direct-injection 4-cylinderengine. SAE Technical Paper 2008-01-0077, 2008.doi:10.4271/2008-01-0077.
[7] P. E. Kapus, A. Fuerhapter, H. Fuchs, andG. K. Fraidl. Ethanol direct injection on tur-bocharged si engines - potential and chal-lenges. SAE Technical Paper 2007-01-1408, 2007.doi:10.4271/2007-01-1408.
[8] Philip Price, Ben Twiney, Richard Stone, Ken-neth Kar, and Harold Walmsley. Particulate andhydrocarbon emissions from a spray guided di-rect injection spark ignition engine with oxy-
genate fuel blends. SAE Technical Paper 2007-01-0472, 2007. doi:10.4271/2007-01-0472.
[9] Heechang Oh, Choongsik Bae, and Kyoung-doug Min. Spray and combustion character-istics of ethanol blended gasoline in a sprayguided disi engine under lean stratified oper-ation. SAE Int. J. Engines, 3:213–222, 10 2010.doi:10.4271/2010-01-2152.
[10] Masaharu Chato, Suguru Fukuda, KiyotakaSato, Tatsuya Fujikawa, Run Chen, Zezheng Li,Jiangping Tian, and Keiya Nishida. Fuel sprayevaporation and mixture formation processesof ethanol/gasoline blend injected by hole-typenozzle for disi engine. SAE Int. J. Engines,5:1836–1846, 10 2012. doi:10.4271/2012-32-0018.
[11] Yongming Bao, Qing Nian Chan, SanghoonKook, and Evatt Hawkes. A comparative analy-sis on the spray penetration of ethanol, gasolineand iso-octane fuel in a spark-ignition direct-injection engine. SAE Technical Paper 2014-01-1413, 2014.
[12] Magnus Sjöberg and David Reuss. Nox-reduction by injection-timing retard in astratified-charge disi engine using gasoline ande85. SAE Int. J. Fuels Lubr., 5:1096–1113, 09 2012.doi:10.4271/2012-01-1643.
[13] Koichi Nakata, Shintaro Utsumi, AtsuharuOta, Katsunori Kawatake, Takashi Kawai, andTakashi Tsunooka. The effect of ethanol fuelon a spark ignition engine. SAE Technical Paper2006-01-3380, 2006. doi:10.4271/2006-01-3380.
[14] Mahmoud K. Yassine and Morgan La Pan. Im-pact of ethanol fuels on regulated tailpipe emis-sions. SAE Technical Paper 2012-01-0872, 2012.doi:10.4271/2012-01-0872.
40
[15] Francesco Catapano, Silvana Di Iorio, Paolo Se-menta, and Bianca Maria Vaglieco. Character-ization of ethanol-gasoline blends combustionprocesses and particle emissions in a gdi/pfismall engine, 2014.
[16] Håkan Sandquist, Maria Karlsson, and IngemarDenbratt. Influence of ethanol content in gaso-line on speciated emissions from a direct injec-tion stratified charge si engine. SAE TechnicalPaper 2001-01-1206, 2001. doi:10.4271/2001-01-1206.
[17] Robert A. Stein, James E. Anderson, and Tim-othy J. Wallington. An overview of the effectsof ethanol-gasoline blends on si engine perfor-mance, fuel efficiency, and emissions. SAE Int.J. Engines, 6:470–487, 04 2013. doi:10.4271/2013-01-1635.
[18] Juntao Wu, Ki Hoon Song, Thomas Litzinger,Seong-Young Lee, Robert Santoro, MiltonLinevsky, Meredith Colket, and David Liscin-sky. Reduction of pah and soot in pre-mixed ethylene–air flames by addition ofethanol. Combustion and Flame, 144(4):675 –687, 2006. doi:dx.doi.org/10.1016/j.combustflame.2005.08.036.
[19] Silvana Di Iorio, Maurizio Lazzaro, Paolo Se-menta, Bianca Maria Vaglieco, and FrancescoCatapano. Particle size distributions from adi high performance si engine fuelled withgasoline-ethanol blended fuels. SAE TechnicalPaper 2011-24-0211, 2011. doi:10.4271/2011-24-0211.
[20] Kenneth Kar, Wai Cheng, and Kaoru Ishii. Ef-fects of ethanol content on gasohol pfi enginewide-open-throttle operation. SAE Int. J. FuelsLubr., 2:895–901, 06 2009. doi:10.4271/2009-01-1907.
[21] Jialin Yang and Richard W. Anderson. Fuelinjection strategies to increase full-load torqueoutput of a direct-injection si engine. SAE Tech-nical Paper 980495, 02 1998. doi:10.4271/980495.
[22] Emmanuel Kasseris and John Heywood.Charge cooling effects on knock limits in si diengines using gasoline/ethanol blends: Part 2-effective octane numbers. SAE Int. J. Fuels Lubr.,5:844–854, 04 2012. doi:10.4271/2012-01-1284.
[23] P. A. Caton, L. J. Hamilton, and J. S. Cowart. Anexperimental and modeling investigation intothe comparative knock and performance char-acteristics of e85, gasohol [e10] and regular un-leaded gasoline [87 (r+m)/2]. SAE Technical Pa-per 2007-01-0473, 04 2007. doi:10.4271/2007-01-0473.
[24] Lukasz P. Wyszynski, C. Richard Stone, andGautam T. Kalghatgi. The volumetric efficiencyof direct and port injection gasoline engineswith different fuels. SAE Technical Paper 2002-01-0839, 2002. doi:10.4271/2002-01-0839.
[25] Geoffrey Cathcart and Christian Zavier. Fun-damental characteristics of an air-assisteddirect injection combustion system as ap-plied to 4-stroke automotive gasoline en-gines. SAE Technical Paper 2000-01-0256, 2000.doi:10.4271/2000-01-0256.
[26] Philip Price, Richard Stone, Tony Collier, andMarcus Davies. Particulate matter and hydro-carbon emissions measurements: Comparingfirst and second generation disi with pfi in sin-gle cylinder optical engines. SAE Technical Paper2006-01-1263, 2006. doi:10.4271/2006-01-1263.
[27] Masaaki Kawamoto, Tetsuya Honda, HideakiKatashiba, Mamoru Sumida, Norihisa Fuku-tomi, and Kazuhiko Kawajiri. A study of cen-ter and side injection in spray guided disi con-cept. SAE Technical Paper 2005-01-0106, 2005.doi:10.4271/2005-01-0106.
[28] James Smith, Gerald Szekely Jr, Arun Solomon,and Scott Parrish. A comparison of spray-guided stratified-charge combustion perfor-mance with outwardly-opening piezo andmulti-hole solenoid injectors. SAE Int. J. En-gines, 4:1481–1497, 04 2011. doi:10.4271/2011-01-1217.
41
[29] Mikael Skogsberg, Petter Dahlander, and Inge-mar Denbratt. Spray shape and atomizationquality of an outward-opening piezo gasolinedi injector. SAE Technical Paper 2007-01-1409,2007. doi:10.4271/2007-01-1409.
[30] Ch. Schwarz, E. Schünemann, B. Durst,J. Fischer, and A. Witt. Potentials ofthe spray-guided bmw di combustion sys-tem. SAE Technical Paper 2006-01-1265, 2006.doi:10.4271/2006-01-1265.
[31] M. Skogsberg, P. Dahlander, R. Lindgren, andI. Denbratt. Effects of injector parameters onmixture formation for multi-hole nozzles in aspray-guided gasoline di engine. SAE TechnicalPaper 2005-01-0097, 2005. doi:10.4271/2005-01-0097.
[32] Petter Dahlander and Ronny Lindgren. Multi-hole injectors for disi engines: Nozzle holeconfiguration influence on spray formation.SAE Int. J. Engines, 1:115–128, 04 2008.doi:10.4271/2008-01-0136.
[33] N. Mitroglou, J. M. Nouri, Y. Yan, M. Gavaises,and C. Arcoumanis. Spray structure gener-ated by multi-hole injectors for gasoline direct-injection engines. SAE Technical Paper 2007-01-1417, 2007. doi:10.4271/2007-01-1417.
[34] Luigi Allocca, Francesco Catapano, AlessandroMontanaro, Paolo Sementa, and Bianca MariaVaglieco. Study of e10 and e85 effect on airfuel mixing and combustion process in opticalmulticylinder gdi engine and in a spray imag-ing chamber. SAE Technical Paper 2013-01-0249,2013. doi:10.4271/2013-01-0249.
[35] Vincent Knop and Eddie Essayem. Comparisonof pfi and di operation in a downsized gasolineengine. SAE Int. J. Engines, 6:941–952, 04 2013.doi:10.4271/2013-01-1103.
[36] Satoshi Taniguchi, Kaori Yoshida, and Yuki-hiro Tsukasaki. Feasibility study of ethanolapplications to a direct injection gasoline en-gine. SAE Technical Paper 2007-01-2037, 07 2007.doi:10.4271/2007-01-2037.
[37] E. Achleitner, H. Bäcker, and A. Funaioli. Directinjection systems for otto engines. SAE TechnicalPaper 2007-01-1416, 2007. doi:10.4271/2007-01-1416.
[38] Yunlong Bai, Zhi Wang, and Jianxin Wang.Knocking suppression using stratified stoichio-metric mixture in a disi engine. SAE TechnicalPaper 2010-01-0597, 2010. doi:10.4271/2010-01-0597.
[39] Magnus Sjöberg, Wei Zeng, and David Reuss.Role of engine speed and in-cylinder flow fieldfor stratified and well-mixed disi engine com-bustion using e70. SAE Int. J. Engines, 7, 04
2014.
[40] Susanne Philipp, Ruediger Hoyer, FrankAdam, Stephan Eckhoff, Rolf Wunsch, ChristofSchoen, and Guido Vent. Exhaust gas aftertreat-ment for lean gasoline direct injection engines -potential for future applications. SAE TechnicalPaper 2013-01-1299, 2013. doi:10.4271/2013-01-1299.
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