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Combustion Chambers for Natural Gas SI Engines Part 2: Combustion and Emissions

Olsson, Krister; Johansson, Bengt

Published in:SAE Transactions, Journal of Engines

1995

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Citation for published version (APA):Olsson, K., & Johansson, B. (1995). Combustion Chambers for Natural Gas SI Engines Part 2: Combustion andEmissions. SAE Transactions, Journal of Engines, 104(SAE Technical Paper 950517).

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950517

Combustion Chambers for Natural Gas SIEngines Part 2: Combustion and Emissions

Krister Olsson and Bengt JohanssonLund Institute of Technology

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950517

Combustion Chambers for Natural Gas SI EnginesPart 2: Combustion and Emissions

Krister Olsson and Bengt JohanssonLund Institute of Technology

ABSTRACT

The objective of this paper is to investigate howthe combustion chamber design will influencecombustion parameters and emissions in a naturalgas SI engine.

Ten different geometries were tried on a convertedVolvo TD102 engine. For the different combustionchambers emissions and the pressure in the cylinderhave been measured. The pressure in the cylinderwas then used in a one-zone heat-release model toget different combustion parameters. The engine wasoperated unthrottled at 1200 rpm with different valuesof air/fuel ratio and EGR. The air/fuel ratio was variedfrom stoichiometric to lean limit. EGR values from 0 to30% at stoichiometric air/fuel ratio were used.

The results show a remarkably large difference inthe rate of combustion between the chambers. Thecycle-to-cycle variations are fairly independent ofcombustion chamber design as long as there is somesquish area and the air and the natural gas are wellmixed.

Geometries that give the fastest combustion givethe highest NOx values at λ=1.2, but at λ>1.5, whichis normally designated lean-burn, the differences aresmaller. The lowest NOx values for lean burn wereobtained with the geometries that gives fastcombustion.

The HC emissions display some correlationbetween high combustion rate and low levels of HCemissions, but combustion chambers with dead zonesand large total combustion chamber areas give higherHC contents than the combustion rate alone wouldindicate.

Indicated efficiency is reduced for combustionchambers with a large total combustion chambersurface area and thus large heat losses. High levels ofturbulence also tend to reduce the efficiency for thesame reason.

INTRODUCTION

Most natural gas commercial vehicles are usingconverted relatively large diesel engines. Thecombustion chamber in these engines is mostcommonly located in the piston crown and a flatcylinder head is used. The inlet port of these enginesoften generates a highly swirling gas motion toenhance the diesel combustion process. In theconversion to spark ignition operation, the originalinlet port is most often used. The original combustionchamber is, however, not directly suitable for SIoperation as the compression ratio often is too highand the flow structure is optimised for spraycombustion rather than the flame propagation of a SIengine. But the question is how the piston crownmodification should be performed to get the minimumamounts of emissions and at the same time a highthermal efficiency.

To get an indication of the importance of thecombustion chamber geometry ten differentgeometries were manufactured for a Volvo TD102 1.6litre single cylinder engine. For these geometries, thein-cylinder flow, combustion and the emissioncharacteristics were measured. In the previous paperthe in cylinder flow measurements were presentedtogether with combustion parameters when theengine operated at λ=1.5. In that paper was foundthat the rate of combustion changed significantly forthe chambers.

In the present paper the emissions as well as thecombustion parameters will be presented with theengine operating with air/fuel-ratios fromstoichiometric to the lean limit as well as with differentamounts of EGR. The combustion events weremeasured by using the cylinder pressure and a simpleheat-release analysis. The indicated mean effectivepressure (IMEP) and the duration of 0-10% and10-90% heat-released were registered among otherparameters.

2

Figure 1: Geometry of the combustion chambers used.

3

The different combustion duration for thecombustion chambers would be expected to producedifferent levels of NOx emissions as the maximumpressure and hence temperature would increase fora fast combustion chamber. The different surfacearea for the combustion chambers would heexpected to influence the levels of HC emissions.

COMBUSTION CHAMBERS

The combustion chambers used to study theeffect of chamber geometry on emissions andcombustion rate were designed in a cut and tryfashion. To make it easy to change combustionchamber geometry a two-piece piston wasdeveloped for a 1-cylinder experimental engine. Thelower part of the piston consists of a standard pistonthat has been cut above the upper piston-ring grooveand then threaded internally. Upper parts withdifferent combustion chamber geometries have thenbeen made. The advantage of this piston is that oneonly needs to remove the cylinder head whenchanging the geometry of the combustion chamber.The disadvantage is that it is not possible withabsolute certainty to compare results with thoseobtained using a single-piece piston, as efficiency isaffected by the different heat transfer characteristicsof the pistons.

The nominal compression ratio for mostchambers was set to 12:1. This ratio correspondswell to the ratio used in present natural gas heavyduty engines [1],[2],[3]. Three geometries were,however, designed with a higher ratio to study theeffect of compression ratio. The main geometric datafor the selected combustion chambers are presentedin figure 1 and table 1. More detailed descriptions ofthe geometries can be found in Part 1.

Table 1: Geometry of the used combustion chambers

EXPERIMENTAL APPARATUS

The engine- The measurements were made in asingle cylinder engine based on a six-cylinder VolvoTD 102 diesel engine. Its main geometric propertiesare shown in table 2.

Table 2: Geometric properties of the engine.

The exhaust gas analysis system- For theexhaust gas measurements a ’P7450 AutomotiveExhaust Gas Analysis System’ from Cusson wasused. The exhaust gas was continuously sampledfrom the exhaust gas stream via a heated sampleline to a heated distribution module. The heateddistribution module contains sample pump and filters.Total hydrocarbons were analysed in a heated flameionisation detector that is fed via a heated line fromthe distribution module. Oxides of nitrogen wereanalysed by a chemiluminescent analyser with aheated capillary module that was also fed via aheated line from the distribution module. Carbondioxide and carbon monoxide were measured bynon-dispersive infrared analysers while oxygen wasmeasured by a paramagnetic analyser. Beforeentering these analysers the sample was cooled toremove excess water vapour, and filtered. See figure2.

Figure 2: The exhaust gas analysis system. [4]

Each analyser was calibrated with appropriatecalibration gas from bottles before and during everymeasurement. More details about the analysers canbe found in [4]. In order to be able to evaluate thepercentage

Comb.Chamber

Squish/Bore

Area/Bore

Bowldepth

Bowldiam.

mm mm

Flat 0 1.3572 "10.7" -

Cylinder 0.66 1.5709 32.7 D=70

Square 0.74 1.7297 42.5 54×54

Cross 0.42 1.7830 22,37 see fig 1

Nebula 0.25 - 28.5 -

Hemi 0.30 1.2631 28.5 R=60

Turbine 0.55 1.4069 25 D=80

Square 16 0.74 1.4740 29.1 54×54

Hemi 16 0.55 1.2257 29 R=45

Cone 0.55 1.2886 42 D=80,d=25

Displaced volume 1600 cm3

Bore 120.65 mm

Stroke 140 mm

Connection rod 260 mm

Exhaust valve open 39 CAD BBDC

Exhaust valve close 4 CAD ATDC

Inlet valve open 2 CAD BTDC

Inlet valve close 42 CAD ABDC

4

of exhaust gas recycled (EGR), the oxygen fractionin the inlet mixture was measured. This O2 analyserwas also a paramagnetic analyser.

The Pressure measurement system- Thepressure in the cylinder was measured with a AVLQC42 piezo-electric transducer connected to aKistler 5001 charge amplifier. The charge amplifiervoltage output was connected to a 486/66 PC with aData Translation DT2823 100 kHz 16-bit A/D-card. Amore detailed description can be found in [5].

The Flow measurement system- The air flowwas measured with a Bronkhorst mass flow meterF-106A-HC. The natural gas flow was measured witha Bronkhorst mass flow meter F-106B-HD.

The control system- The ignition timing wascontrolled with a PC-controlled system. Triggeringsignals to the pressure-system were also included inthis system. Input signals to the control system werea sync-pulse (1 pulse per 2 revs), a TDC pulse (1pulse per rev) and a crank angle-pulse (5 pulses percrank angle degree, CAD).

OPERATING CONDITIONS

The engine was run on natural gas that was fed tothe engine through two pulse width-modulatedsolenoid valves.

The valves were controlled by an IntelligentControl IC5460 engine management system. To geta homogenous charge the mixing length from thesolenoid valves to the engine was 3 m with a 16 litremixing tank in the middle. The influence of differentamounts of fuel and air into the engine from cycle tocycle is therefore reduced.

During all emissions and cylinder pressuremeasurements, the engine was run at 1200 rpm andno throttling was applied.

DATA REDUCTION

One-zone heat-release model- To extractinformation on the flame development, acycle-resolved heat-release calculation wasperformed. In the computations Wochnis heattransfer model [6] was applied and the ratio ofspecific heats was assumed to have a lineardependence on temperature. Further detailsconcerning the heat-release calculation have beendescribed elsewhere [5].

Relative air/fuel ratio- To evaluate emissionlevels and compare them under differentcircumstances, the following formulas were used tocalculate the relative air/fuel ratio, λ [4].

λ is defined as follows:

where A/F is air fuel ratio by mass

Global combustion reaction is assumed to be:

λ is calculated as follows:

where

and

The * denote dry exhaust gases and K is theequilibrium constant in the water gas reaction. Forthe calculations in this paper K = 3.5 and m/n = 3.76.

Exhaust Gas Recycled (EGR)- The EGR(%)used in this paper is defined:

λ

AF----

actual

AF----

stoichiometric

-------------------------------------------=

CnHmOr λ no2O2 3.773 N2⋅+( ) ⇒⋅ ⋅+

np (⋅ x̃ CHCaHb x̃coCO x̃co2CO2 x̃o2

O2+++

x̃N2N2 x̃NONO x̃NO2

NO2 x̃H2OH2O x̃H2H2)+ ++ + +

where no2n m

4----- r

2---–+=

λnp x̃H2O x̃NO 2 x̃NO2

⋅+ +( )⋅

2 no2⋅

--------------------------------------------------------------------------=

np 1 x̃H2O–( ) x̃CO2

* 2 x̃O2*⋅+

⋅ ⋅

2 nO2⋅--------------------------------------------------------------------------------------+

x̃H2O2

m x̃CO* x̃CO2

*+ ⋅

2 n 1x̃CO*

K x̃CO2*⋅

---------------------- m2 n⋅----------- x̃CO

* x̃CO2*+

⋅+ +⋅ ⋅

----------------------------------------------------------------------------------------------------------------=

npn

x̃CH 1 x̃H2O–( ) x̃CO* x̃CO2

*+ ⋅+

-------------------------------------------------------------------------------------=

EGR %( ) 100m· EGR

m· i------------------

⋅=

5

where mEGR is the mass of exhaust gas recycledand mi is the total intake mixture. The relative air/fuelratio was maintained λ=1.0 during all the EGRmeasurements by the λ control option in theIntelligent Control engine management system.

The following formulas were used to calculateEGR(%):

EGR* = EGRflow / (Airflow+ Fuelflow)

E = EGRflow has the O2 fraction OE

A + F = Airflow + Fuelflow has the O2 fraction OA+F

A + F + E = Airflow + Fuelflow + EGRflowhas the O2 fraction OA+F+E

An oxygen balance gives:

OA+F · (A + F) + OE · E = OA+F+E · (A + F+ E)

The equation can be rearranged to

(A + F)·(OA+F - OA+F+E) = E · (OA+F+E - OE)

and thus

EGR*(%) = 100·(OA+F - OA+F+E) / (OA+F+E - OE)

EGR(%) and EGR* (%) are related by

EGR(%) = 100 · EGR*(%) / (100 + EGR*(%))

Lean Burn or EGR?- To get lowest possibleemissions there are two possibilities. The first is touse as lean an air/fuel mixture as possible and then ifnecessary add an oxidising catalyst. The second isto run the engine at λ=1.0 and use as much EGR aspossible and if necessary add a three-way catalyst.Both alternatives were tested and the results arepresented below, first lean burn and then EGR. Allmeasured emissions are from raw undiluted exhaustgas sampled close to the exhaust port.

LEAN BURN

Combustion- The different combustionchambers have as indicated before a large spread inthe rate of heat-release. To obtain optimumperformance the ignition angle θi has to be adjustedfor the different combustion chambers and fordifferent λ. Figure 3 below shows the ignition angle θiwhen the engine was run with MBT and with differentλ. As can be seen the Square combustion chamberhas the latest ignition timing for MBT. This indicates

a very fast combustion rate. The Flat combustionchamber with no squish, has the earliest ignitiontiming.

Figure 3: The crank angle position for ignition timingwhen different combustion chambers are used. Theengine operated at MBT and 1200 rpm.

Figures 4 and 5 show the crank angle for 0 to 10%and 10 to 90% of the total heat released when theengine was run with MBT and with different λ. As canbe seen it is the Square combustion chamber thathas the fastest 10 to 90% combustion.

Figure 4: The flame-development angle ∆θd for 0 to10% of the total heat released when differentcombustion chambers are used. The engineoperated at MBT and 1200 rpm.

6

The Flat combustion chamber, with no squish, isclearly the slowest with almost twice as long durationfor the combustion. There are less differences for the0 to 10% combustion but the combustion chambersfollow more or less the same order. The exception isthe Hemi that has the smallest flame developmentangle but the largest rapid-burn angle. Thecombustion 15 CAD after ignition is still mostlylaminar and for the same λ the laminar flame speedis about the same. The laminar flame speed is alsoinfluenced by pressure and temperature and asdifferent MBT values give somewhat differentpressure and temperature histories can this be thereason for these small differences detected.

Figure 5: The rapid-burn angle ∆θb for 10 to 90% ofthe total heat released when different combustionchambers are used. The engine operated at MBTand 1200 rpm.

Figure 5 also shows the lean limit for the differentcombustion chambers as the right end point of thecurves. The lean limit in this paper is the highestλ-value or EGR(%) that can be used for acombustion chamber with 0% misfire during 300continuous cycles. The reason for this rather harshdefinition is that the λ-value is calculated fromemission values and with misfires the values from theanalysers will not be stable enough for a correct λcalculation. The Square and the Cross have thehighest lean limit and the Flat and the Hemi thelowest.

Cycle-to-cycle variations- The cycle-to-cyclevariations for these combustion chambers areobtained as the standard deviation of indicated meaneffective pressure for 300 continuous cycles dividedby the mean value (COVimep). Figure 6 shows theCOVimep for the different combustion chamberswhen the engine was run with MBT and with different

λ. The figure shows that with one exception COVimepis almost the same for the different combustionchambers and with just a small increase forincreasing λ-values up to the lean limit. The Flatcombustion chamber with no squish is the only onewith a COVimep over 1% without any misfires andprobably as a result of partial burn near the lean limit.

Figure 6: The COVimep for different combustionchambers. The engine operated at MBT and 1200rpm.

Figure 7: The indicated efficiency for differentcombustion chambers. The engine operated at MBTand 1200 rpm.

Efficiency- The indicated efficiency, , isinfluenced of heat losses and flow losses, bothdependent on the generated turbulence. Figure 7shows the indicated

ηi

ηi

7

efficiency when the engine was run at MBT andwith different λ The figure indicates that turbulencegenerating design features such as large squishareas decrease the efficiency.

Emissions- Figure 8 shows the HC emissions forthe different combustion chambers. The Crossshows the highest levels of HC for all λ-values. Asthe Cross has the largest piston surface area this canresult in a rather large contribution of HC due to wallquenching but there are also a lot of "dead" cornerswith slow flow rate and a high possibility of partialburn. The Nebula shows the lowest levels of HC forall λ-values. Unfortunately, the crevice volume for theNebula is only 56% compared to the others withsquish area because the first piston ring is locatedhigher on that one-piece piston. This is probably theexplanation for the Nebula’s low HC emissions [7].The crevice volume for the Flat is 68% compared tothe others with squish area because the piston crownis lower on that piston to get the correct compressionratio. The small crevice volume and the smallestpiston area compensate the lack of turbulenceexcept when the λ-values are greater than 1.6. Thenthe increase of HC as a function of λ is greater thanaverage due to a higher amount of partial burn. TheSquare with large squish areas and the fastestcombustion maintains good HC values even for thehighest λ-values.

Figure 8: The HC emissions for different combustionchambers. The engine operated at MBT and 1200rpm.

Figure 9 shows the NOx emissions for thedifferent combustion chambers. The maximumvalues at λ=1.2 show that the Square with the fastestcombustion has the highest value and the Flat hasthe lowest. The other combustion chambers show noeasily explained pattern.

The NOx maximum value is very sensitive toignition timing and a small offset from MBT givesnoticeable results. The maximum value is of courseof less interest than the NOx value close to lean limit.Figure 10 shows the NOx emissions for the differentcombustion chambers when λ>1.4. The Square andthe Cross have the lowest NOx values because theyhave the highest lean limit. The Flat and Nebula alsoshow a low minimum level of NOx.

Figure 9: The NOx emissions for different combustionchambers. The engine operated at MBT and 1200rpm.

Figure 10: The NOx emissions for differentcombustion chambers. The engine operated at MBTand 1200 rpm.

As the sum of HC and NOx is used as a measurein legislation of emission levels this is shown in figure11

ηi

8

plotted as a function of λ. This figure indicates thatthe Nebula and the Square are the best combustionchambers for lean burn.

Figure 11: The HC+NOx emissions for differentcombustion chambers. The engine operated at MBTand 1200 rpm.

EXHAUST GAS RECYCLED (EGR)

Figure 12: The crank angle position for ignition timingwhen different combustion chambers are used. Theengine operated at MBT and 1200 rpm.

Combustion- Figure 12 above shows the ignitionangle θi when the engine was run with MBT and withdifferent EGR. The Square combustion chamber hasthe

latest ignition timing for MBT, followed by the Cross.The Flat has the earliest ignition timing.

Figures 13 and 14 show the crank angle for 0 to 10%and 10 to 90% of the total heat released when theengine was run with MBT and with different EGR.

Figure 13: The flame-development angle ∆θd for 0 to10% of the total heat released when differentcombustion chambers are used. The engineoperated at MBT and 1200 rpm.

Figure 14: The rapid-burn angle ∆θb for 10 to 90% ofthe total heat released when different combustionchambers are used. The engine operated at MBTand 1200 rpm.

As can be seen it is the Square combustion chamberthat has the fastest 10 to 90% combustion. The Flatchamber,

9

with no squish, and the Hemi are clearly the slowestwith almost twice as long duration for thecombustion.

There are less differences for the 0 to 10%combustion but the combustion chambers followmore or less the same order. These figures look asexpected very much like figures 4 and 5 for leanburn.

Figure 14 also shows the EGR limit for thedifferent combustion chambers as the right end pointof the curves. The Square has the highest EGR limitand the Cylinder the lowest. This low value for theCylinder is somewhat surprising but it was ratherdifficult to obtain stable EGR values over 25% due tothe EGR control system and it is possible that thelast EGR value without misfire for the Cylinder issomewhat higher than the one in the figure.

Cycle-to-cycle variations- Figure 15 shows theCOVimep for the different combustion chamberswhen the engine was run with MBT and with differentEGR. The figure shows that with one exceptionCOVimep is almost the same for the differentcombustion chambers and is almost constant up tothe EGR limit. The Flat combustion chamber with nosquish is the only one with a COVimep over 1%without any misfires and this only as a result of partialburn near the EGR limit.

Figure 15: The COVimep for different combustionchambers. The engine operated at MBT and 1200rpm.

Efficiency- Figure 16 shows the indicatedefficiency when the engine was run with MBT andwith different EGR. The figure indicates thatturbulence generating design features such as largesquish areas decrease the efficiency. Compared withlean burn the values for indicated efficiency are lowerfor EGR and only show a small increase forincreasing EGR values.

Figure 16: The indicated efficiency for differentcombustion chambers. The engine operated at MBTand 1200 rpm.

Emissions- Figure 17 shows the HC emissionsfor the different combustion chambers. The Crossshows the highest levels of HC for all EGR-values asit did for lean burn. The reasons are believed to bethe same as for lean burn. The small crevice volumeand the smallest piston area for the Flat compensatethe lack of turbulence and give this combustionchamber the lowest level of HC emissions. It isinteresting to note the much smaller increase in HCwhen EGR is applied compared to lean burn.

Figure 17: The HC emissions for differentcombustion chambers. The engine operated at MBTand 1200 rpm.

ηi

ηi

10

Figure 18 shows the NOx emissions for thedifferent combustion chambers. The Cylinder, theSquare and the Cross have lowest NOx values. Apossible explanation is that these combustionchambers have the largest residual gas volumes dueto the perpendicular walls of the recess in the piston.The same tendencies could be found for mediumλ-values in Figure 9 for lean burn.

Figure 18: The NOx emissions for differentcombustion chambers. The engine operated at MBTand 1200 rpm.

Figure 19: The HC+NOx emissions for differentcombustion chambers. The engine operated at MBTand 1200 rpm.

As the sum of HC and NOx is used as a measurein legislation of emission levels this is plotted as afunction of EGR in figure 19. This figure indicatesthat the Flat is the best combustion chamber for

EGR. This is mainly due to the low HC emissionsfrom the small crevice volume.

CORRELATIONS

Emissions and compression ratio- NOxemissions and compression ratio showed nocorrelation for the tested geometries. However thiswas not expected [8].

Increasing compression ratio increases the HCemissions. Figure 20 shows the HC emissions as afunction of compression ratio for different combustionchambers for λ=1.8.

Figure 20: HC emissions as a function ofcompression ratio for different combustionchambers. The engine operated at MBT and 1200rpm.

Several factors could contribute: increasedimportance of crevice volumes at high compressionratio, lower gas temperatures during the latter part ofthe expansion stroke, thus producing less HCoxidation in the cylinder; lower exhausttemperatures, and hence less oxidation in theexhaust system [8].

Figure 21: HC emissions as a function of relativepiston area for different combustion chambers. Theengine operated at MBT and 1200 rpm.

11

Piston area and HC emissions- A correlationcan be found between piston area and HC emissionsfor the different combustion chambers but only atλ=1.3 when the HC emissions are low. Then theeffect of wall quenching is dominating and no bulkquenching due to partial burn is likely to appear.Figure 21 above shows the HC emissions as afunction of relative piston area (piston area/borearea) for different combustion chambers for λ=1.3.

Piston area and NOx emissions- A correlationcan be found between piston area and NOxemissions for the different combustion chambers.Figure 22 shows the NOx emissions as a function ofrelative piston area for different combustionchambers for λ=1.5.

Figure 22: NOx emissions as a function of relativepiston area for different combustion chambers. Theengine operated at MBT and 1200 rpm.

DISCUSSION

Is it now possible to choose the best strategy,lean burn or EGR, and the best combustion chamberfrom the presented results? The choice depends onif emissions or efficiency or a mixture of both are theobjectives to meet [10]. In figure 23 below HC+NOxis plotted as a function of indicated efficiency for thedifferent combustion chambers and for λ>1.4. Theinformation from the figure is that the low emissionvalues for the Square are combined with a rather lowefficiency and it is best to choose between the Flatand the Nebula. All emissions that have beenmeasured are raw emissions. If a hypotheticaloxidation catalyst with a 70% reduction of HCemissions is added the possible result is shown infigure 24. With this hypothetical catalyst the lowestemissions are obtained with the Square and theNebula. The efficiency of the Nebula is howeversuperior and is thus the recommended choice.

Figure 23: The HC+NOx emissions as a function ofindicated efficiency for different combustionchambers and # values. The engine operated atMBT and 1200 rpm.

Figure 24: The 0.3*HC+NOx emissions as a functionof indicated efficiency for different combustionchambers and # values. The engine operated atMBT and 1200 rpm.

The same figures for different EGR values arefound below. In this case the hypothetical catalyst isa three-way catalyst with a 75% reduction of HCemissions and a 90% reduction of NOx emissions.

12

Figure 25: The HC+NOx emissions as a function ofindicated efficiency for different combustionchambers and EGR values. The engine operated atMBT and 1200 rpm.

Figure 26: The 0.25*HC+0.1*NOx emissions as afunction of indicated efficiency for differentcombustion chambers and EGR values. The engineoperated at MBT and 1200 rpm.

The increase from 70% to 75% reduction of theHC emissions is justified by the increased exhausttemperature of EGR operation. The information fromthe figures shows that it is best to choose the Flatcombustion chamber. This is somewhat surprising.Aside from the small crevice volume for the Flatcombustion chamber, a possible explanation for thisresult is that the definitions of lean limit and EGR limitnever give the turbulence generating

designs a chance to prove that they can extend leanlimit and EGR limit compared to the Flat. But in thefuture HC emissions can not be neglected and a leanlimit defined at 5% COVimep [1],[2],[9] is likely toproduce too much HC emissions.

EGR combined with a three-way catalyst shows thelowest emission values at the price of a lowerefficiency than lean burn.

CONCLUSIONS

1. Regardless of combustion chamber design λ=1.0plus EGR gives somewhat better total (HC+NOx)emissions, but with a lower efficiency, than lean burn.

2. Combined with a three-way catalyst λ=1.0 plusEGR shows a potential for really low total emissionvalues that are not possible to achieve with lean burnand an oxidising catalyst.

3. The Square combustion chamber which has thefastest combustion has the lowest NOx emissions atlean burn.

4. If no misfire is accepted the Flat combustionchamber and the Nebula combustion chamber showthe lowest total emissions together with the bestefficiencies.

REFERENCES

[1] T. Naganuma, M. Iko, T. Sakonji, F. Shoji:"Basic Research on Combustion Chambers forLean Burn Gas Engines"1992 Int. Gas Research Conference.

[2] T. Sakurai, M. Iko, K. Okamoto, F. Shoji:"Basic Research on Combustion Chambers forLean Burn Gas Engines"SAE932710

[3] W.R. Dietrich, W. Grundmann, G. Langeloth:"Pollutant Reduction on Stationary S.I. Enginesfrom Motoren-Werke Mannheim for Operationon Natural Gas Applying the Lean-BurnPrinciple" (In German) MTZ, MotortechnicheZeitschrift 47(1986) 3 pp83-87

[4] J Nilsson:"A Study on a Low Emission Combustor - LeanPremix Prevaporize Concept" Licenciate ThesisISRN LUTMDN/TMVK--7014-SE Dept. of Heat& Power Engineering, Lund Inst. of Techn.1993.

[5] B. Johansson:"Correlation Between Velocity ParametersMeasured with Cycle-Resolved 2-D LDV andEarly Combustion in a Spark Ignition Engine"Licenciate Thesis ISRN LUTMDN/TMVK --7012-SE Dept. of Heat & Power Engineering,Lund Inst. of Techn. 1993

[6] G. Woschni:"A Universally Applicable Equation forInstantaneous Heat Transfer Coefficient in theInternal Combustion Engine"SAE670931

13

[7] H. Carstensen:"Systematische Untersuchung derKonstruktions- und Betriebsparameter einesZweiventilmagermotors auf Kraftstoffverbrauch,Schadstoffemission und Maximalleistung"Dissertation TUW 1991 Vortschrittberichte VDIReihe 12 Nr 163

[8] J. B. Heywood:"Internal Combustion Engine Fundamentals"ISBN 0-07-100499-8 p844.

[9] M.G. Kingston Jones, M.D. Heaton:"Nebula Combustion System for Lean BurnSpark Ignited Gas Engines"SAE890211

[10] Y. Nakajima, K Sugihara and Y. Takagi:"Lean Mixture or EGR - Which is better for FuelEconomy and NOx Reduction?"I Mech E Conference Publication C94/79 1979.