boosting and direct injection

DEVELOPMENT Mixture Formation

6 MTZ worldwide 12/2004 Volume 65

To further reduce the Corporate Average Fuel Economy in order to meetthe ACEA target values agreed upon, more intense ef-forts are required in the areas of engine and drive traindevelopment by 2008 or 2012. FEV has thoroughly ana-

lyzed this kind of concept and analyzed the fundamentalsynergy effects resulting from the additional combination of

supercharging with direct injection in close detail.

1 Introduction and Status-Quo

After launching direct-injection gasolineengines with stratified charge on the mar-ket, gasoline direct injection has come tothe attention of the developers, in particu-lar in regard to its potential in combinationwith boosting [1, 2, 3]. It should be taken in-to account for which market or vehicle cat-egory these engines are being developed. Iffor instance the German vehicle market isanalyzed in regard to New European Dri-ving Cycle consumption and power/weightratio, Figure 1, the advantages in consump-tion become clear, in particular when usedfor niche or sports concepts. The low-end-torque behaviour that is actually disadvan-tageous looses its significance in sports en-gines or is compensated because automat-ed transmissions (e.g. Smart) are beingused.

However, higher volumes are only real-ized in (upper) middle class vehicles, whichare strongly dominated, at least in Europe,by manual transmission. A basic weaknessof these engines – the engine torque at lowengine speeds, the “low-end-torque” – be-comes apparent. Particularly, with tur-bocharged engines, the bmep that can bereached at 1000 rpm only lies slightly abovethat of naturally aspirated engines, Figure2. Reasons for this are the low mass flowrate, the low exhaust gas temperature, andthe relative low turbocharger efficiency inthis operating range. This disadvantage isalready effective in steady-state engine op-eration; in the case of transient operation,the effect of the necessary turbochargerspeed adaptation is added, and the disad-vantage compared to naturally aspiratedgasoline engines increases. On the otherhand, maximum torque can already be

By José Geiger,

Knut Habermann,

Oliver Lang,

Betina Vogt and

Michael Wittler

Aufladung und Direkteinspritzung

– Synergien für zukünftige Ottomotoren

You will find the figures mentioned in this article in the German issue of MTZ 12/2004 beginning on page 970.

Boosting andDirect InjectionSynergies for Future Gasoline Engines

7MTZ worldwide 12/2004 Volume 65

reached in turbocharged engines below2000 rpm, which sometimes requirestorque limitation in the low gears. The re-sult of the weak low-end-torque is that thetransmissions are usually designed low-geared, which again has a negative impacton the fuel consumption, e.g. in the NewEuropean Driving Cycle. On the other hand,the progressive steep torque rise is con-ceived as sporty or as difficult to drive de-pending on the driver. In principle, theboost pressure adjustment at low enginespeeds can be managed by the charger size,while there is a compromise with regard tothe specific output that can be reached as aresult of the turbine's throttling effect. En-gines with a high specific output of morethan 90 kW/l usually do not reach theirmaximum torque until the engine speedexceeds 2000 rpm.

The extrapolation of the specific torquesvalid for naturally aspirated engines that isshown in Figure 2 can be used to estimate atarget range (“Target Turbo DISI”) for theideal boosted engine. A bmep of more than14 bar should already be reached at 1000rpm. However, this target value depends onthe considered total displacement and thevehicle.

2 Basic Engine Design

Especially at boosted engines with a com-paratively low compression ratio, thecharge motion layout has a high impact onthe knocking behaviour. On the one hand,there is also a trade-off between the chargemotion level at low speeds and the maxi-mum possible air flow rate at rated power.However, in contrast to the naturally aspi-rated engine, this disadvantage can becompensated by adjusting the boost pres-sure level in the boosted engine after therated torque has been reached. All mea-sures that also aim at increasing the low-end-torque aggravate the knocking diffi-culties. Thus, for the design of boosted en-gines it is favourable to apply charge mo-tion intake port concepts. CAE methodolo-gy was used to optimize the intake portgeometry.

Figure 3 shows the layout of the flowcharacteristics of two different intake portversions, a filling and a tumble port. The re-sults of the engine tests show significantadvantages for the version with the highercharge motion [4], so that this is used insubsequent development.

In comparison to the port fuel injection(PFI), advantages due to both, increasedcylinder charge and reduced knocking sen-sitivity, can be observed with direct injec-tion. In order to better understand the dif-ferences between external and internal

mixture formation, borderline cases can beused that represent an idealized intakeprocess and vaporization, Figure 4. Exter-nal mixture formation, in other words theconventional PFI, is characterized by thecomplete fuel vaporization before the mix-ture induction process, during which theheat of vaporization is fully taken from theintake port wall. A “lower” borderline caseof the internal mixture formation describesthe full air intake, during which vaporiza-tion does not take place until the “inlet clos-es”, and which leads to an about 2 % highercylinder mass with λ = 1. The “upper” bor-derline case on the other hand describes thefuel vaporization during the inductionprocess, during which the heat of vaporiza-tion is fully taken from the intake air andwith a high cylinder mass of approximate-ly 8 %. Both, the port fuel injection and di-rect injection are real processes that are po-sitioned between the first and the last bor-derline case. Despite the increased cylindermass, the charge temperature decrease ofdirect injection usually causes a reducedknocking sensitivity. In the engine usedhere, it was possible to increase the com-pression ratio in comparison to the baseversion with port fuel injection by morethan one unit to ε = 10. The same air massflow was needed again in the event of rat-ed power. However, it was possible to re-duce fuel consumption in almost the entiremap.

Especially at the boosted engine withhigh injection rates per cylinder, direct in-jection also requires careful tuning of themixture formation in order to prevent anincrease in soot and HC emissions as wellas an increasing oil dilution as a result ofincreased wall wetting. CFD calculations [5]are used as a basis for the injection layout.Figure 5 serves as an example to showstudies on spray propagation and mixtureformation with swirl or multi-hole injec-tion nozzle. Multi-hole injection nozzles of-fer a high degree of freedom with regard tothe location of the individual holes andhave a spray pattern that is almost inde-pendent from the back pressure.

It can be seen, how the fuel mass is de-flected by the intake air down towards thepiston. When the piston is in bottom deadcentre, the swirl injector shows a clear wallcontact of the injection spray in the area ofthe exhaust valve. This also leads to a high-er stratification of mixture in the combus-tion chamber even during the compressionphase, so that at the end of compressionzones remain with relatively rich mixture.In contrast to that, it is possible to reducethe wall wetting with the multi-hole injec-tion nozzle. The rich mixture zones at theend of compression are reduced, and thus

the soot formation at high load is positive-ly affected.

In order to evaluate the wall contact acharacteristic number is used taking intoaccount the relative air/fuel ratio. Thecourse of the weighted wall contact overthe intake and compression stroke isshown in the right side of Figure 5 for bothinjector versions. The multi-hole injectionnozzle generally shows advantageous char-acteristics, since it permits the specific de-sign of the injection spray geometry.

Due to the cylinder wall contact and theoil adsorption, the quality of the mixtureformation with direct injection clearly af-fects the engine oil dilution with fuel. Oildilution is determined using a standardizedtest procedure, in which the engine is oper-ated cold conditioned, Figure 6. The drasti-cally reduced amount of fuel added speaksfor a clearly improved mixture formation.

Thus, this procedure yields an efficientmethodology, which includes the compre-hensive assessment of combustion systemand durability.

3 Gas Exchange

A GT-Power model of the engine was set upto optimize the FEV Turbo DISI. In particu-lar, the model takes into account the lowerengine speed range in order to permit low-end-torque studies and transients at lowspeeds with high accuracy. For example,the GT-Power extrapolation method for tur-bocharger maps was optimized and theburning characteristics of the engine as afunction of load and speed were imple-mented in the model [6].

When setting-up the model, special at-tention was also paid to a detailed model-ling of the engine periphery in order permitan exact modelling of gas dynamic phe-nomena, e.g. such as the exhaust blow-down reflection. This is required especiallyin those cases, when the engine is operatedwith large valve overlaps at low speed.

At full load and low engine speeds, tur-bo-charged gasoline engines typically fea-ture a positive scavenging pressure drop.This can be used in engines with inner mix-ture formation to scavenge the cylinderwith fresh air during gas exchange. In do-ing so, the volumetric efficiency raises andthe mass flow rate through compressor/turbine is increased. Thus, the operatingpoint of the compressor/turbine is shiftedto higher efficiencies. Figure 7 shows thefull load operating points for the PFI baseversion and the one of the Turbo DISI in thecompressor map. It is apparent that the fullload operating points at engine speeds of1000 and 1250 rpm clearly diverge from thesurge line and a higher pressure ratio can


8 MTZ worldwide 12/2004 Volume 65

be realized. In the mid speed range, scav-enging fresh air is used to reduce the knock-ing sensitivity, thus lowering the boostpressure demand. The latter also yieldsfavourable transient behaviour. In therange of high engine speeds, the efficiencyof the engine can be increased due to the re-duced knocking sensitivity (inner mixtureformation). If the power output is constant,the air mass flow is therefore reduced.Thus, the combination of gasoline direct in-jection and variable timing can significant-ly help to reduce the required compressormap width.

When tuning the combustion system,injection timing and air/fuel ratio are im-portant parameters in addition to valvetiming. Design of experiments (DoE) wasused here as a support tool, since the highdegree of freedom would lead to a largeamount of testing involved. For example,Figure 8 shows the solution space found toobtain a high torque and simultaneously ahigh efficiency for a speed of n = 1000 rpm.On the one hand, there is a limitation dueto a bmep of > 13 bar, and on the other handa specific fuel consumption of 400 g/kWhacts as a limit for the solution space. More-over, a relative air/fuel ratio of λ = 1.0 was arequested boundary condition for the ex-haust gas aftertreatment. It is apparentthat, in comparison to the large theoretical-ly possible adjustment range of the valvetiming, it was possible to notably reducethe solution space.

4 Engine-related Results

Figure 9 shows the steady-state full-loadresults. In comparison to the PFI base ver-sion, the gasoline direct injection and vari-able timing at a speed of n = 1000 rpm canbe used to increase the mean effective pres-sure by approximately 35 % to about 14 bar.Here, In principle, additional increase ispossible depending on the engine design.At a speed of n = 1250 rpm, a rated torque ofapproximately 260 Nm (bmep = 18 bar) canalready be achieved, whereas the value forthe PFI base version is “merely” 164 Nm,which corresponds to an improvement of58 %.

This significant increase of the low-end-torque can be achieved by utilizing the pos-itive scavenging pressure drop at lowspeeds. As a result, a volumetric efficiencyof up to 125 % is achieved (relative to the in-take manifold condition). The additionalscavenging air mass results in a leanerair/fuel ratio in the exhaust than in thecylinder, so that it increases from λ = 0.9 toλ = 1.0. Thus, it is possible to realize a mini-mum specific fuel consumption of below250 g/kWh at full load.


2 Basic Engine Design

Figure 4: Influence of mixture formation on volumetric efficiency

Figure 5: Influence of spray geometry on mixture formation (full load at 5000 rpm)

Figure 6: Oil dilutionassessment in FEV scatter-band

9MTZ worldwide 12/2004 Volume 65

Furthermore, it is possible to showfavourable consumption at partial loadswith the aid of the variable timing and thecompression ratio that is high for boostedengines. Figure 10 displays the Turbo DISIin the bsfc scatter band at n = 2000 rpm /bmep = 2 bar. It reveals the Turbo DISI be-ing clearly below the boundary line of tur-bocharged gasoline engines.

An inherent weakness of all downsizingconcepts with turbo-charging remains thesignificantly poorer dynamic behaviourcompared to that of naturally aspirated en-gines. Figure 11 shows the comparison ofengines with identical power for a tran-sient load step from bmep = 2 bar to fullload at 1250 rpm. The naturally aspiratedgasoline engine reaches its maximumtorque almost directly and steadily follow-ing the throttle angle in less than a half sec-ond, whereas the boosted versions are onlyable to keep up with this property in thefirst tenths of a second. The subsequentdrop in the torque gradient basically indi-cates that the particular engine has reachedthe naturally aspirated full load. A directadditional increase of the torque is onlypossible to a limited degree due to the mo-ment of inertia of the turbocharger that isto be accelerated. Anyway, a higher basictorque level offered by the turbo DISI en-gine at low speeds already has an approxi-mately 10 % advantage compared to thebase version. With regard to the transientbehaviour, the combination of boostingand variable timing with direct injection isclearly preferable compared to PFI. The Tur-bo DISI version with identical rated poweralready reaches the nominal torque of 1250rpm approximately two seconds earlierthan the PFI turbo version. About 1.5 sec-onds after the naturally aspirated engine itreaches its nominal torque and after anoth-er approximately 1.5 seconds the Turbo DISIproduces a torque 20 % higher comparedwith the naturally aspirated engine. Com-pared to the PFI base version, the torque is50 % higher at this time reference point.Overall, the torque build-up is much moreharmonious.

5 Conclusions and Outlook

The assessment of the fuel saving potentialfor the Turbo DISI engine is performed withregard to the estimated extra costs. For thispurpose, the fuel saving in the New Euro-pean Driving Cycle is shown in Figure 12. Avehicle with 1,500 kg test weight is used asa basis, which is equipped with a 2.0l natu-rally aspirated engine. The turbo versionsused for comparison have the same ratedoutput; thus, the swept volume decreasesto approximately 1.4 l. As a result of the

downsizing effect and the associated loadpoint shift, a consumption reduction of 13-18 % has been obtained with the tur-bocharged engine with manifold injection(“PFI turbo”). The associated extra cost com-pared to the naturally aspirated engine ofapproximately 200 Euro can be regarded asmoderate.

Moreover, using direct injection in com-bination with turbocharging leads to animproved low-end-torque behaviour, thusat least reducing the effect of the turbo lag.Besides, direct injection – as describedabove – can be used to generally increasethe compression ratio and gain advantagesin residual gas control. With these aspects,another approximately 6 % fuel consump-tion reduction can be expected in the NewEuropean Driving Cycle in combinationwith a possible extension of the final driveratio.

The outlook of additionally using a vari-able compression ratio indicates a total fu-el consumption reduction potential of ap-proximately 30 %.

The presented analysis results impres-sively demonstrate the synergy effectsthat can be achieved through the combina-tion of boosting and direct injection. Theyoffer a significant potential for the en-hancement of torque and efficiency ofgasoline engines; in addition, the respon-siveness of turbocharged engines will benotably improved. It can be expected thatthis concept will find broad acceptance inall vehicle and engine categories in thelong-term. Advantages compared to mod-ern diesel engines become apparent suchas the high low-end-torque or the wideroverall speed range.

Besides the detailed optimization of theoverall system the following issues can beaddressed for future development:■ Catalytic exhaust gas aftertreatment:The fluctuating admission of oxygen andrich exhaust in the catalytic converter atlow engine speeds may, under un-favourable circumstances, accelerate thecatalyst aging behaviour.■ Increase of the maximum mean effec-tive pressure: Improved drivability as wellas the favourable overall efficiency enablesa maximum torque level of approximately25 bars with single-stage boosting. Theproblem of the so-called pre-ignition ormega knocking has to be observed in thiscontext.■ New boosting technologies and high-pressure boosting: Variable turbine geome-try, 1050 °C turbine or two-stage boostingwill have to be measured in the future as tohow their potential can be applied to thisnew gasoline engine standard or whetherit supports further optimization.

There are no major technical obstaclespreventing this engine concept from beingused worldwide. Therefore, the boostedgasoline engine faces clearly rising marketshares – similar to what was the case withthe diesel engine in the past years.


References

[1] Alt, M.; Schaffner, P.; Krebs, W.; Quarg, J.:Benzindirekteinspritzung in Kombination mitAufladung. 10. Aachener KolloquiumFahrzeug- und Motorentechnik 2001

[2] Brinkmann, F.; Pingen, B.; Walder, K.: Ben-zindirekteinspritzung mit Turboaufladung – einBrennverfahren für Downsizingkonzepte. Hausder Technik, München, 2003

[3] Krebs, R.; Böhme, J.; Dornhöfer, R.; Wurms,R.; Friedmann, K.; Helbig, J.; Hatz, W.: Derneue Audi 2.0T FSI Motor – der erste direk-teinspritzende Turbo-Ottomotor bei Audi. 25.Int. Wiener Motorensymposium, April 2004

[4] Lang, O.; Geiger, J.; Habermann, K.; Sehr, A.;Vogt, B.: Optimierungspotenziale von Otto-Turbomotoren mit Direkteinspritzung. 12.Aachener Kolloquium Fahrzeug- und Mo-torentechnik 2003

[5] Geiger, J.; Breuer, M.; Adomeit, P.; Grünefeld,G.: Visualisierung und Modellierung derLadungsschichtung im DI Ottomotor – derSchlüssel zum besten Brennverfahren. Hausder Technik, Essen, 2001

[6] Lang, O.; Habermann, K.; Sehr, A.; Brackhaus,N.; Wohlberg, R.: Einsatz von Methoden undWerkzeugen zur Optimierung des transientenBetriebsverhaltens von aufgeladenen Verbren-nungsmotoren. HdT-Tagung “Ansaugsys-teme, Package und Kunststoffe”, Stuttgart,Juni 2004

boosting and direct injection

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