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International J of Engine Research2017, Vol. 18(10) 1055–1066� IMechE 2017Reprints and permissions:sagepub.co.uk/journalsPermissions.navDOI: 10.1177/1468087417699979journals.sagepub.com/home/jer

Thermal efficiency of a dual-modeturbulent jet ignition engine under leanand near-stoichiometric operation

Ravi Teja Vedula, Ruitao Song, Thomas Stuecken,Guoming G Zhu and Harold Schock

AbstractTurbulent jet ignition is a combustion technology that can offer higher thermal efficiency compared to the homogeneousspark ignition engines. A potential combustion-related challenge with turbulent jet ignition is the pre-chamber misfiringdue to improperly scavenged combustion residuals and maintaining the mixture composition there. Dual-mode turbulentjet ignition is a novel combustion technology developed to address the aforementioned issues. The dual-mode turbulentjet ignition is an engine combustion technology wherein an auxiliary air supply apart from an auxiliary fuel injection isprovided into the pre-chamber. This technology can offer enhanced stoichiometry control and combustion stability inthe pre-chamber and subsequently combustion control in the main chamber. In this work, engine testing of a single-cylinder dual-mode turbulent jet ignition engine having a compression ratio of 12.0 was completed with liquid gasolineand the indicated thermal efficiency was measured. High-speed pressure recordings were used to compare and analyzedifferent operating conditions. Coefficient of variation in the indicated mean effective pressure and the global air/fuelequivalence ratio values were used to characterize the engine operation. Lean operating conditions for a global air/fuelequivalence ratio of 1.85 showed an indicated efficiency of 46.8% 6 0.5% at 1500 r/min and 6.0 bar indicated mean effec-tive pressure. In addition, the combustion stability of this engine was tested with nitrogen dilution. The nitrogen diluentfraction was controlled by monitoring the intake oxygen fraction. The dual-mode turbulent jet ignition engine of com-pression ratio 12.0 delivered an indicated efficiency of 46.6% 6 0.5% under near-stoichiometric operation at 1500 r/minand 7.7 bar indicated mean effective pressure with a coefficient of variation in indicated mean effective pressure of lessthan 2% for all conditions tested.

KeywordsThermal efficiency, direct-injection spark ignition, dual-mode turbulent jet ignition, intake dilution, lean burn, stoichio-metric operation

Date received: 28 September 2016; accepted: 21 February 2017

Introduction

A major focus in light-duty vehicle development is toimprove the thermal efficiency of direct-injection sparkignition (DISI) engines and reduce exhaust emissionsusing various strategies such as high compression ratio,charge dilution, tumble enhancement, high ignitionenergy, and late intake valve closure timing (Miller orAtkinson cycle). Most passenger cars utilize less than10% of their maximum engine power during daily com-mute.1 This fact highlights the importance of improvingengine part-load efficiency. Here and throughout therest of this work, the term ‘‘part-load’’ is used to indi-cate the target range of 5–10 bar BMEP (brake meaneffective pressure). According to a recent benchmarking

study conducted by the United States EnvironmentalProtection Agency (US EPA), the current productionengines carry a part-load brake thermal efficiency rang-ing from 30% to 35%.2 An exception to this range arethe Mazda’s 2.0L engine that showed a peak brakeefficiency of 37% under part-load conditions and

Energy & Automotive Research Laboratory, Department of Mechanical

Engineering, Michigan State University, East Lansing, MI, USA

Corresponding author:

Ravi Teja Vedula, Energy & Automotive Research Laboratory,

Department of Mechanical Engineering, Michigan State University, East

Lansing, MI 48824, USA.

Email: vedulara@egr.msu.edu

Toyota’s 1.3L Atkinson cycle engine of compressionratio 13.5 with a brake efficiency of 38%. Reese3 madea ‘‘propulsion system efficiency’’ analysis based on2015 US EPA certification data to estimate the netimprovements required in the engine, transmission, anddriveline efficiencies to meet US 2025 greenhouse gasesregulations. According to this analysis, a further 30%reduction in fuel consumption is required for gasolineengines if all this improvement is to originate solelyfrom engine development.3,4 If a thermal efficiency of35% is considered as the current industry standard, thisanalysis would indicate a brake thermal efficiency of50% (=35/0.7) to meet the 2025 regulations.

Direct injection is more advantageous than the portfuel injection for engine part-load operating conditions.One of the reasons is with port fuel injection, the engineload is controlled by throttling the intake which incurspumping losses. However, direct-injection system offersunthrottled operation where the engine load can becontrolled by the amount of fuel injected, thereby mini-mizing pumping losses. Other advantages of directinjection include increased volumetric efficiency,reduced heat losses, and the ability to operate at highcompression ratios. Stratified charge direct-injectioncombustion has attracted a lot of attention in the lastfew decades to improve part-load efficiency of direct-injection engines and reduce exhaust emissions.5–9 Instratified charge DISI engines, the turbulent flame initi-ated upon ignition of a heterogeneously rich mixture atthe spark plug leads to diffusion-controlled combustionof the remaining bulk mixture. The ‘‘leanness’’ of thismixture increases with distance from the spark plug inthe combustion chamber. The stratification in DISIengines can be attained by one of the three conventionalsystems: wall-guided stratified charge, air-guided stratifiedcharge, or spray-guided stratified charge combustion sys-tems.8 Wall- and air-guided stratified combustion systemshave shown unfavorable emission, low specific power, andhigh fuel consumption due to the impingement of directly-injected fuel on the piston and cylinder wall.10–12 Spray-guided systems are capable of overcoming several unfa-vorable consequences seen with the other two systems.However, spray-guided systems are susceptible to randommisfires and partial burns.13

According to the vehicular emission regulations oflight-duty vehicles as reviewed by Johnson,14 a 75%reduction in fleet average non-methane organic gases(NMOG) and NOx will be phased in from 2015 to2025. By 2025, the target NMOG + NOx will be30mg/mile for both cars and light-duty trucks. Thesetargets should be achieved using a combination ofbetter-quality combustion and economically feasibleexhaust aftertreatment systems. Studies have shownthat NOx reduction and stable combustion can beattained simultaneously when exhaust gas recirculation(EGR) is implemented along with other combustion-enhancing strategies.15,16 Diana et al.17 noted that thebenefits of higher compression ratio and improvedengine efficiency can be attained by employing EGR

(this reduces NOx and mitigates knocking), along withenhanced in-cylinder flow, in agreement with the recentdevelopmental strategies of Toyota engines.18 The leanand EGR dilution limits might be extended usingadvanced ignition systems.19,20 Exhaust emissions aretypically controlled using exhaust aftertreatment sys-tems. Among the different aftertreatment systems avail-able, the three-way catalyst (TWC) has been mostwidely used, which is a low-cost alternative and uses awell-proven feedback system. The TWC is a multi-functional catalyst that can (1) oxidize CO and HC(hydrocarbon) to CO2 and (2) reduce NOx to N2 andH2O.21 A TWC has an optimal conversion efficiencyunder near-stoichiometric conditions.

Turbulent jet ignition (TJI) is a pre-chamber-initiatedtwo-chamber combustion system.22,23 The TJI systemmajorly involves three sections: a small pre-chamber, amulti-orifice nozzle, and a large main chamber. A spark-ignited reactive mixture in the pre-chamber flowsthrough the nozzle orifices and results in multiple, che-mically active, turbulent jets that emerge into the mainchamber. With this ignition strategy, TJI has beendemonstrated to ignite leaner mixtures and extend theflammability limits beyond those obtained using sparkignition combustion. Vaporized gasoline was used as thepre-chamber fuel by Attard and Blaxill,24 and the enginewas operated successfully at an air/fuel equivalence ratiogreater than 2.1 with minimal NOX emissions. Inanother work, TJI with gasoline in main chamber andgaseous propane in pre-chamber was demonstrated tomaintain stable combustion with 50% more diluentmass (air and residual gases) than spark ignition com-bustion.25 Apart from extending the lean limit, TJIoffers a faster burn rate than conventional spark ignitiondue to the rapidly propagating highly energetic reactingjet. Distributed, rapid burn rates of lean or dilute mix-tures reduce the chance of knocking as there is less timeavailable for the end gas to autoignite.

There are several existing barriers that need to beaddressed before the TJI combustion technology isviable for production-based applications. These bar-riers include the need to scavenge the pre-chamber andprecise control of pre-chamber stoichiometry. In a con-ventional TJI system with auxiliary injection, only fuelis injected into the pre-chamber. In these conventionalsystems, the air/fuel mixture from the main chamberback feeds into the pre-chamber thereby disadvanta-geously causing an unknown air-to-fuel ratio (AFR)within the pre-chamber. In the past two years, theauthors in collaboration with industry partners con-structed a dual-mode turbulent jet ignition (DM-TJI)system.26 A distinctive feature of this DM-TJI system isthe inclusion of a dedicated air supply to the pre-cham-ber. With this design, the AFR inside the pre-chamberand the main chamber can be controlled independently.This DM-TJI system not only permits ignition oflean mixtures but also permits the combustion rateto be controlled by management of pre-chamberstoichiometry.

1056 International J of Engine Research 18(10)

Developing an engine which can go from currentstate-of-the-art 35% to a brake efficiency of 40% atpart-load conditions for US light-duty vehicles has thepotential impact of saving nearly 1.2million barrels ofoil per day. In this work, the potential of the DM-TJIcombustion technology to offer high thermal efficiencyunder part-load operation is demonstrated:

� Under lean-burn operation consisting of a globalair/fuel equivalence ratio (l) of 1.85 without intakecharge dilution;

� Under near-stoichiometric operation with 30%intake charge dilution using nitrogen as the diluent.

The target performance of this naturally aspiratedDM-TJI engine having a compression ratio of 12.0 is40% brake efficiency at an engine load of 6–8bar indi-cated mean effective pressure (IMEP) and speeds rang-ing from 1200–2000 r/min.

Experimental setup

High compression ratio DM-TJI engine

The single-cylinder engine consists of a flat head thatwas pre-heated by flowing coolant through the flow pas-sages. The coolant used was 50:50 ethylene glycol–watermixture that was heated to 90 �C using an electricallypowered heating element. Using a PID (proportional-integral-differential) controller, the temperature wasmaintained at this set temperature during the entire test-ing time. Such an approach of pre-heating the head isusually employed for single-cylinder research engines toattain metal engine steady temperatures. Figure 1 shows

a drawing of the engine highlighting various parts of theDM-TJI system installed on the head. It is to be notedthat this engine configuration was obtained by modify-ing an earlier optical engine version. As a result of this,the DM-TJI engine has a metal piston with an attachedBowditch extension. A high-pressure fuel injector and ahigh-pressure air injector were inserted in the pre-cham-ber. The orientation plane of these pre-chamber injec-tors is out-of-paper and, hence, is not viewable in Figure1. Figure 2 renders a better view of the injector locationsin the pre-chamber. For the main chamber fueling, ahigh-pressure fuel injector was located in the intake port(see Figure 1). The fuel used was certification gradeliquid gasoline for both pre-chamber and main cham-ber. A measuring spark plug (Kistler 6117BFD17) thatincorporates a pressure sensor was employed for pre-chamber pressure measurements. A six-orifice nozzlewas used to connect the pre-chamber to the main cham-ber. Dimensions and other details of this engine aregiven in Table 1.

Figure 1. High compression ratio DM-TJI engine with metal piston.

Figure 2. Design details of the pre-chamber.

Vedula et al. 1057

Coefficient of variation (COV) in the IMEP deter-mined over 100 engine cycles was used as an index forcombustion instability. The stability limit was definedby a COV in IMEP less than 2.0%. A wideband oxy-gen sensor mounted in the exhaust manifold was usedto record the l values. The global l value was used as ametric to represent leanness of fuel–air mixture. Thesetwo variables, COV IMEP and global l, were chosento design the current set of operating conditions for thenaturally aspirated DM-TJI engine.

Intake charge dilution using nitrogen

EGR was simulated by adding nitrogen to the engineintake flow. No measures were taken to control thetemperature of nitrogen inflow. Intake dilution usingnitrogen allowed control of the oxygen content flowinginto the engine without influencing the combustionprocess as nitrogen is chemically inert to combustionunlike carbon dioxide.27 Hence, nitrogen was chosenfor intake dilution, and the combustion stability ofintake-diluted near-stoichiometric air/fuel mixture wastested under wide open throttle conditions. While usingpure gases for intake charge dilution, a gas havinghigher specific heat value can result in minimal NOxformation.28 The specific heat of nitrogen is higherthan CO2, and thus, intake dilution using nitrogen gasis expected to produce less NOx compared to that withCO2.

Nitrogen was supplied from a bank of nine com-pressed N2 cylinders, arranged next to the engine. Aquarter-turn ball valve was manually opened to allownitrogen to flow through a large plenum attached tothe intake manifold, upstream of the throttle plate, asshown in Figure 3. A hollow cylinder was attached totop of the plenum through which the nitrogen supplytube was inserted. The end of nitrogen tube that goesinto the plenum has five orifices (four tangential andone vertical). The air entered the engine at the plenumthrough the annular region of the hollow cylinder. Theorifice design as employed for the nitrogen tube wasintended for better air and nitrogen mixing. As shownin Figure 3, two oxygen sensors were employed while

working with N2 dilution. The nitrogen dilution wasquantified by the amount of intake oxygen measuredusing the intake oxygen sensor. The global l whichaccounts for the total pre-chamber and main chambermixture composition was measured from the exhaustoxygen sensor. A high exhaust O2 fraction is indicativeof a high temperature combustion, which further resultsin high NOx formation.29 A low exhaust O2 fractionindicates that all the oxygen was consumed for burningfuel during combustion. This could further indicateeither a stoichiometric or fuel-rich combustion. One ofthe primary goals of this study is to operate the DM-TJI engine under stoichiometric conditions with N2

dilution. The N2-simulated EGR dilution is aimed atreducing NOx formation during combustion, and stoi-chiometric operation is aimed at reaching efficient cata-lyst operation for a closely coupled TWC (oxidize COand HC). Therefore, for the current engine setup, oper-ating conditions that resulted in an exhaust O2 fractionof close to zero value were sought.

Table 1. Engine specifications.

Bore 95 mmStroke 100 mmConnecting rod length 190 mmCompression ratio 12.0:1Motored peak pressure 22.5 bar at TDCPre-chamber volume 2700 mm3 (;0.4% of displacement volume)Main chamber swept volume 0.709 LFuel injection High-pressure injectors for both chambersFuel EPA LEV-II liquid gasoline (both chambers)Engine speed 1500 r/minEngine load (naturally aspirated at wide open throttle) ;6.0 bar IMEP (without intake dilution)

;7.7 bar IMEP (with intake dilution)

TDC: top dead center; IMEP: indicated mean effective pressure.

Figure 3. Experimental setup of DM-TJI engine with nitrogendilution.

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Engine controls and data acquisition

As previously mentioned, a Kistler type 6117BFD17spark plug and pressure sensor combination was usedto record pre-chamber pressure data. A piezoelectricpressure transducer (Kistler type 6125C) inserted in themain chamber was used to record main chamber pres-sure data. The charge output from the transducers wasconverted to an amplified voltage using a DSPTechnologies model 1104CA charge amplifier. AMototron controller sent the control signals to an injec-tor drive box to control injection and spark timing, andinjection pulse widths as shown in Figure 4. The con-troller received the crank and cam signals from theengine for syncing with the compression top dead cen-ter (TDC) position. A Wineman System DCDynamometer was used to manage engine speed andload. The chamber pressure–related signals wererecorded during engine operation using an A&DTechnologies’ Phoenix-AM high-speed combustionanalysis system (CAS) capable of recording data at0.1 crank angle resolution. The data from the currenttesting were recorded at 1.0 crank angle resolution forsaving acquisition time and ease in data handling. TheCAS TDC position was synchronized with piston TDC

position using an encoder. After acquiring data withone test condition, the control parameters were chan-ged using the Mototune software that allows to sendsignals to the Mototron controller. Different operatingconditions tested in this work are listed in Table 2. Amulti-pulse fuel injection approach was employed formain chamber fueling in order to have longer mixingtimes and/or to reduce wall impingement.

Results and discussion

DM-TJI engine without intake dilution

All of the following tests without intake dilutioninvolved only one fuel injection pulse in the pre-cham-ber, with a pulse width of 0.90ms. The main chamberwas fueled using five split injections starting fromintake TDC with a pulse width of 0.75ms and a pulse-to-pulse duration of 4ms. With the pre-chamber injec-tion timing set at 80 CAD bTDC (crank angle degreebefore top dead center), different spark timings weretested at 1500 r/min and 6bar IMEP speed/load condi-tion and a global l value between 1.85 and 1.88. Bothpre-chamber peak pressure and main chamber heatrelease rate (HRR) increased with spark advance asshown in Figure 5. The HRR calculation was based onHeywood’s30 model, while the heat loss terms areneglected for a simple analysis. With a spark timing of25 CAD bTDC, the main chamber peak HRR occurredearlier than the second peak occurrence of pre-chamberpressure. The indicated efficiency defines the amount offuel energy being converted into mechanical energy thatdrives the piston during combustion. This fuel utiliza-tion efficiency was calculated using the followingequation

Indicated Efficiency=IMEP3Swept Volume

LHV3mfuel

� �

ð1Þ

where LHV is the measured lower heating value ofEPA-LEV II gasoline fuel=41.8MJ/kg and mfuel is

Table 2. Control parameters for pre-chamber and main chamber.

Pre-chamber air:Injection pressure 60 barInjection pulse width 1.2–1.6 msInjection timing 0.2 ms before pre-chamber fuel injection

Pre-chamber fuel:Injection pressure 103 barInjection pulse width 0.9–1.2 ms; single pulseInjection timing 160–40 CAD bTDC

Main chamber fuel:Injection pressure 103 barInjection pulse width Five pulses of 0.75–1.00 msInjection timing 360 CAD bTDCSpark timing 36–15 CAD bTDC

CAD bTDC: crank angle degree before top dead center.

Figure 4. Engine controls system.

Vedula et al. 1059

the total amount of fuel injected into the pre-chamberand main chamber.

There are several factors of uncertainty involved inthe process of measuring the indicated efficiency.Bench calibrations were performed for both fuel injec-tors (main and pre-chambers) at each fuel injectioncondition tested to determine the fuel mass injected.The injection command was sent to the pre- or mainchamber fuel injector and the injected fuel was collectedin a measuring tube placed under ambient conditions.As such, the actual amount of fuel injected into thepre-chamber might vary depending on the pre-chamberpressure at the time of injection during engine opera-tion. The main chamber injector calibrations, however,are not expected to deviate significantly from the actualvalues, given that these injection sequences occurred inthe intake port during engine operation at wide openthrottle. Overall, this calibration method is believed toprovide reasonable results as the amount of fuelinjected into the pre-chamber is only 3%–4% of thetotal fuel injected into both chambers per cycle. Afterrepeating the main chamber injector calibrations, it wasnoted that the deviation in the amount of mass mea-sured reflected in an efficiency difference of 60.5%.These factors of uncertainty are to be noted while dis-cussing the DM-TJI engine efficiencies in the rest ofthis work. Bowditch-style extended pistons can beprone to axial and radial deformations due to pistonacceleration and gas pressure forces during engineoperation. However, the resulting volume changes dueto piston deformation were noted to have insignificanteffect on IMEP values.31 Under the DM-TJI operatingconditions at 1500 r/min, the peak cylinder pressure isaround 45bar. Based on a structural analysis study pre-viously conducted for a similar engine configuration,these acceleration and pressure forces when combinedresulted in a piston axial deformation of less than0.5mm. This value falls in the lower range of the datareported by Aronsson et al.31 Thus, the indicated effi-ciency values are believed to be unaffected due to pis-ton deformation.

Hereafter, unless otherwise written explicitly, theefficiency values should be appended with 60.5% to

extract the actual values. For the same amount of fuelused, the spark timing of 25 CAD bTDC resulted inthe highest indicated efficiency of 46.0% compared tothe other two spark timings. This highlights that themaximum work was extracted with the advanced sparktiming for the given fuel and air injection conditions.

With spark timing set at 25 CAD bTDC, differentpre-chamber fuel timings were tested as follows: (1) 120CAD bTDC—early fuel injection (EFI); (2) 80 CADbTDC—middle fuel injection (MFI); (3) 40 CADbTDC—later fuel injection (LFI). The MFI pre-chamber fueling condition was employed in the earlierdiscussion related to Figure 5. Also recall that thiscombination of injection and spark timing resulted inan indicated efficiency of 46.0%. Ignition delay in thepre-chamber was defined as the duration between sparkinitiation and the time when the pre-chamber pressuredifferential first reached about 25% of its maximumpressure differential. This percentage was chosen basedon visual inspection of inflection point on the pressuretrace. With this definition, it was observed that the pre-chamber ignited faster with EFI condition followed byMFI and LFI conditions. As shown in Figure 6 (left),the first pre-chamber peak pressure was higher for LFIcondition indicating the occurrence of a high-speed jetin this injection condition.

Similarly, ignition delay in the main chamber wasdefined as the duration between spark initiation and thetime when the main chamber pressure differential firstreached about 25% of its maximum pressure differen-tial. It was observed that the main chamber ignitedalmost at the same time since the start of pre-chamberignition for all fuel injection conditions. The CA50angle of main chamber, however, was 2–3CADs earlierfor the LFI condition compared to the other two injec-tion conditions. The LFI condition showed higher com-bustion pressure (Figure 6, right), which was 3.5 barmore than that obtained with the EFI condition. Withthe global l ffi 1:85, the EFI injection condition resultedin 0.8% higher indicated efficiency compared to MFIand 0.6% higher than with LFI condition. The lowchamber pressure possibly resulted in low combustiontemperature with the EFI condition. Low temperaturecombustion would incur less heat losses, thereby, result-ing in high indicated work output with EFI condition.Therefore, with this pre-chamber injection testing, thenew indicated efficiency of the DM-TJI engine is 46.8%at 1500 r/min/6.0 bar IMEP. It should be noted that thisefficiency does not subtract the work required to pumphigh-pressure air into the pre-chamber. A preliminaryanalysis predicted less than 0.5% efficiency reductiondue to pumping work for pre-chamber air injection.Also, the current non-optimized results were obtainedfor a fixed spark timing.

DM-TJI engine with intake dilution

Benefit of injecting air in the pre-chamber. The advantageof having an air injection in the pre-chamber can be

Figure 5. Pre-chamber pressures and main chamber heatrelease rates for different spark timings.

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illustrated using the IMEP plots in Figure 7. These1500 r/min/6.0 bar IMEP conditions were created froma pre-chamber fuel injection timing of 80 CAD bTDC,fuel pulse width of 1.15ms, and spark timing of 25CAD bTDC. At these operating conditions, with nitro-gen dilution of 25.8% (intake O2% of 15.5%; global l

of 1.33), one misfire was detected without air injectionin the pre-chamber and no misfires occurred with pre-chamber air injection. When the charge dilution wasfurther increased to 27.3%, several misfires weredetected without pre-chamber air injection. Amongthese misfire cycles, majority of misfires occurred in themain chamber, while occasional misfires were detectedin the pre-chamber as well. The absence of misfires byhaving an air injection in the pre-chamber demon-strates the capability of the DM-TJI engine to operateat an ideal EGR-diluted near-stoichiometric mixtureburning condition. The EGR dilution helps in reducingNOx formation, and the stoichiometric operation pro-vides catalyst light-off temperatures in the exhaustmanifold to oxidize CO and unburnt HC emissionsand/or reduce NOx using a close-coupled TWC. In the

tested geometrical configurations, for the aforemen-tioned operating conditions with a global l;1:33, 25%seems to the dilution limit using nitrogen because theCOV in IMEP was greater than 2% (i.e. target stabilitylimit) as highlighted in Figure 7.

Spark sweep for an early pre-chamber injection timing. AnEFI in the pre-chamber resulted in higher indicatedefficiency without intake dilution as seen earlier whilediscussing Figure 6 results. This effect was tested herewith nitrogen dilution. Spark sweeps were conductedfor a different EFI timing of 160 CAD bTDC in thepre-chamber while maintaining a constant intake dilu-tion of approximately 30%. To achieve the goal ofnear-stoichiometric operation with nitrogen dilution,the amount of fuel injected into the main chamber wasfurther increased to a total fuel pulse width of 5.00ms.This higher fuel mass in the main chamber resulted in amaximum IMEP of 7.7 bar, compared to 6.0 bar asobtained for conditions without intake dilution. Also,the pre-chamber air injection pulse width was slightly

Figure 6. Pre-chamber and main chamber pressures for different pre-chamber fuel injection timings.

Figure 7. Effect of having pre-chamber air injection on the combustion of diluted mixture consisting an intake O2 of 15.5% (left)and 15.2% (right).

Vedula et al. 1061

increased to compensate for the predominantlynitrogen-diluted main chamber mixture entering thepre-chamber during the upward piston movement inthe compression stroke.

The corresponding results obtained for differentspark timings are shown in Figure 8. The COV inIMEP was less than 2% for a wide range of spark tim-ings demonstrating the high combustion stabilityobserved at these nitrogen-diluted operating condi-tions. The CA50 crank angle occurred earlier withspark advance. With a spark timing beyond 33 CADbTDC, however, the combustion instability increased.The exhaust O2 was observed to be around 1.0% forall the spark timings. The indicated efficiency reached amaximum of 46.6% at 27 CAD bTDC, then plateauedup to 33 CAD bTDC, and dropped at 36 CAD bTDCspark timing. Based on these results, the optimal oper-ating points that offered high combustion stability,near-stoichiometric combustion, and high thermal effi-ciency are circled in Figure 8.

Figure 9 shows the HRRs for selected spark timingswith the same pre-chamber fuel injection condition.The spark timing that delivered the highest indicatedefficiency resulted in an earlier and higher peak HRRin the main chamber. Spark timing beyond 27 CADbTDC, however, showed a decline in the peak HRR.This lower peak HRR and the slower combustionwould be the factors leading to lower efficiency and rel-atively high combustion instability at 36 CAD bTDCas seen earlier in Figure 8. In a different viewpoint, theprolonged duration for the HRR to reach 0–10 J/CAD

with a retarded spark timing could be related to themixture preparation times in the main chamber. For aretarded spark timing of 15 CAD bTDC, the HRRcurve showed a gradual rate of heat release than for theother spark timings. This could indicate that the mainchamber contained locally excess lean regions due toadditional mixing time available before spark. Thus,the reacting jet likely took longer to initiate combustionin the main chamber compared to the other spark tim-ings. These speculations on the possible inhomogeneityof fuel–air mixture in the main chamber are limitedbased on the results observed and can be investigatedin more detail using time-resolved experiments andcomputational models.

Thermal efficiency of the DM-TJI engine. The net meaneffective pressure (NMEP) of the engine can be attainedafter subtracting the pumping losses. As all the condi-tions discussed so far involved wide open throttle, theengine pumping losses were less than 3% of the IMEPvalues. It should be noted that no intake/exhaust cam-shaft and manifold dynamics were optimized to reducethe pumping work as this was out of scope of this study.The net thermal efficiency can be calculated by repla-cing IMEP with NMEP in equation 1. Figure 10 showsthe non-optimized net thermal efficiency maps for vari-ous engine speeds and loads obtained at throttled orwide open throttle conditions without intake chargedilution. The efficiency maps having throttled condi-tions are included here only to show various tests con-ducted with this engine, while the target operation ofthis DM-TJI engine is under wide open throttle condi-tions. As seen in Figure 10, a highest net thermal effi-ciency of 45.5%6 0.5% was obtained at 2000 r/min,6.5 bar NMEP, and a global l of 1.7.

Based on different operating conditions tested withnitrogen diluent, the most favorable conditions for theDM-TJI engine at 1500 r/min and 7.7 bar IMEP deliv-ered an indicated efficiency of 46.6%6 0.5% undernear-stoichiometric operation (exhaust O2ffi 0.9%) with

Figure 8. Graphical plots showing the DM-TJI engine testresults with approximately 30% intake dilution.

Figure 9. Main chamber heat release rates for different sparktimings with an early fuel injection in the pre-chamber.

1062 International J of Engine Research 18(10)

30% intake dilution. In terms of NMEP, this near-stoichiometric operation with 30% dilution resulted ina net thermal efficiency of 45.5%. Toulson et al.32

observed that the combustion stability can be main-tained at higher levels of dilution (at least 10% more)with EGR compared to nitrogen as used in this work.

Holmberg et al.33 performed a detailed study on fric-tion losses of a passenger car. According to their calcu-lations, friction losses of a 1.7-L four-cylinder engineaccount to 11.5% of the fuel energy used. Engine fric-tion losses are known to increase with engine speed andload. While the brake power was mentioned to be12kW in the work by Holmberg et al.,33 the enginespeed was not explicitly mentioned (vehicle speed=60km/h). For the single-cylinder DM-TJI engine, theindicated power at 1500 r/min is 6.9 kW. Few studieswere available that provided information related tofriction loss at an equivalent operating condition.Friction losses were noted to be around 10% at 1500 r/min and 4–8kW power.34 A heat balance analysis of aToyota 2ZR (1.8L) four-cylinder engine showed lessthan 4% mechanical losses at 2000 r/min and 7.0 barBMEP.18 In another study, Sellnau et al.35 determined

a friction mean effective pressure of 50–90 kPa for afour-cylinder gasoline compression ignition engine at1500 r/min and 8 bar IMEP. Based on these literaturefindings, the friction loss percentage is assumed to be11% of the indicated power under speed/load condi-tions equivalent to the DM-TJI engine operation. Onimposing this friction loss % for the DM-TJI engine,the brake thermal efficiency after subtracting theengine-related friction losses can possibly be 40% asillustrated in Figure 11. Alternatively, the engine fric-tion losses could be calculated more accurately usingempirical correlations developed by taking into accountthe engine speed, load, and in-cylinder pressure.36,37

However, the constants or the multiplicative factorsinvolved in these friction loss correlations are deter-mined experimentally for a specific engine design. Thistask was not attempted during the DM-TJI engineoperation. Hence, the nominal friction loss % for four-cylinder engines as identified from the literature surveywas assumed for the current discussion. An averagebrake thermal efficiency of 35% is the current industrystandard for light-duty vehicle engines under part-loadoperation.2,18,38 Thus, the DM-TJI engine with a possi-ble brake efficiency of 40% at near-stoichiometric oper-ation with intake dilution can possibly reduce the fuelconsumption by 12.5%.

It is to be emphasized that these brake efficiency cal-culations are based on moderate engine testing effortswhere some of the operating parameters such as sparkand valve timings were not completely optimized.Another important factor to consider is the amount ofNOx and particulates released from the DM-TJI engineat the operating conditions discussed in this work. Theamount of fuel injected was smaller in the lean-burnoperating mode compared to the nitrogen-diluted stoi-chiometric mode. An emission bench was installedrecently and the preliminary measurements showed asteady engine-out NOx of 140ppm under lean condi-tion and 260ppm under near-stoichiometric operatingcondition. An extensive study that involves the mea-surement of exhaust NOx, unburned HCs, and sootconcentrations of the DM-TJI engine at various oper-ating conditions is underway.

Figure 11. Estimated efficiency of the DM-TJI engine in converting fuel energy into useful work.

Figure 10. Net thermal efficiency maps of the DM-TJI enginewithout intake dilution.

Vedula et al. 1063

Conclusion

The DM-TJI engine with a compression ratio of 12.0was tested at wide open throttle using both lean air/fuelmixture and nitrogen-diluted near-stoichiometric mix-ture. The engine was demonstrated to have high com-bustion stability at both operating modes with a COVin IMEP less than 2%. An air injection in the pre-chamber favors the dilution limit extension of the TJIcombustion technology. During the time of ignition,the main chamber’s nitrogen-diluted mixture which hasless densely occupied oxygen is pushed into the pre-chamber. Injecting air into the pre-chamber wasbelieved to (1) compensate such lack of oxygen, thusaiding in combustion there; (2) scavenge some of thisextraneous mixture, thus creating desirable AFR in thepre-chamber. Other major observations made from thiswork are as follows:

� Without intake dilution (lean burn; global l ffi 1:85):For a pre-chamber fuel injection timing of 80 CADbTDC (middle injection condition), advancing thespark to 25 CAD bTDC resulted in highest peakHRR and indicated efficiency. When tested withdifferent injection timings with spark timing set to25 CAD bTDC, the pre-chamber ignited faster withearly injection condition compared to middle andlater injection conditions. Also, the early injectioncondition resulted in the highest indicated thermalefficiency of 46.8% at 1500 r/min/6.0 bar IMEP.

� With nitrogen dilution (near-stoichiometric):Nitrogen-diluted conditions showed several misfiresin the main chamber and/or the pre-chamber with-out air injection in the pre-chamber. For the sameintake dilution levels, no misfires were detected byincluding pre-chamber air injection. With 30%dilution and the pre-chamber fuel injection at 160CAD bTDC, the maximum spark advance resultedin lower efficiency due to the lower HRR. For agiven nitrogen dilution fraction, a retarded sparktiming is believed to delay the main chamber igni-tion and flame propagation. This was attributed tolonger mixing times before spark, thereby possiblyresulting in locally lean mixtures in the mainchamber.

� An earlier pre-chamber ignition time, a lower pre-chamber pressure before reaching its first peak, anda lower main chamber peak pressure resulted in highindicated efficiency. Thus, developing combustionphenomenon that leads to less heat losses in bothpre-chamber and main chamber could be an effec-tive strategy in improving engine thermal efficiency.

� With the DM-TJI engine demonstrated, near-stoichiometric mixture created using large amountsof nitrogen dilution (’30%) provided an indicatedefficiency greater than 46%. It should be noted thatthis efficiency includes work required to supplyhigh-pressure air to the pre-chamber. With this

operating condition, a conventional three-way cata-lytic converter can be employed to manage criteriaemissions, making this technology viable with thecurrent state-of-the-art aftertreatment technology.

� The net thermal efficiency obtained after subtract-ing the non-optimized pumping work and not sub-tracting friction losses was noted to be45.5%6 0.5% for both lean and near-stoichiometric operations. A preliminary estimateshowed that the DM-TJI engine could reduce thefuel consumption by 12.5% compared to the con-ventional spark ignition engines at the currentspeed/load conditions.

In the next steps, exhaust emission measurements areto be determined at various operating conditions (leanand stoichiometric). In addition, developmental effortsfor a DM-TJI engine with a higher compression ratioof 15:1 are underway.

Acknowledgements

The authors thank Jen & Jeff Higel, Brian Rowley, andJohn Pryzbyl for their services in building the DM-TJIcomponents and for engine maintenance. They alsothank the reviewers for taking their time in improvingthe quality of this presentation.

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research, authorship, and/or publi-cation of this article.

Funding

The author(s) disclosed receipt of the following finan-cial support for the research, authorship, and/or publi-cation of this article: This material is based upon worksupported by the United States Department of Energyand the National Science Foundation Partnership onAdvanced Combustion Engines under contract CBET-1258581.

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Appendix

Notation

l air/fuel equivalence ratio (=AFR/AFRStoichiometric); AFR is air-fuel ratio

mfuel total fuel mass injected

Vedula et al. 1065

bTDC Before top dead centerCAD Crank angle degreeCAS Combustion analysis systemCOV Coefficient of variationDISI Direct-injection spark ignitionDM-TJI Dual-mode turbulent jet ignitionEFI Early fuel injectionEGR Exhaust gas recirculationHC HydrocarbonHRR Heat release rateIMEP Indicated mean effective pressure

LFI Later fuel injectionLHV Lower heating valueMFI Middle fuel injectionNMEP Net mean effective pressureNOx Mixture of oxides of nitrogenNPI Normalized particulate indexrpm Revolutions per minuteSOI Start of injectionTDC Top dead centerTJI Turbulent jet ignition

1066 International J of Engine Research 18(10)