Energy Conversion and Management 112 (2016) 369–381

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Energy Conversion and Management

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Energy efficiency impact of EGR on organizing clean combustionin diesel engines

http://dx.doi.org/10.1016/j.enconman.2016.01.0420196-8904/� 2016 Elsevier Ltd. All rights reserved.

⇑ Corresponding author at: Department of Mechanical, Automotive & MaterialsEngineering, University of Windsor, 401 Sunset Avenue, Windsor, Ontario N9B 3P4,Canada. Tel.: +1 (519)253 3000x2636; fax: +1 (519)973 7007.

E-mail address: mzheng@uwindsor.ca (M. Zheng).

Prasad S. Divekar a, Xiang Chen b, Jimi Tjong a, Ming Zheng a,⇑aDepartment of Mechanical, Automotive & Materials Engineering, University of Windsor, Windsor, Ontario, CanadabDepartment of Electrical and Computer Engineering, University of Windsor, Windsor, Ontario, Canada

a r t i c l e i n f o a b s t r a c t

Article history:Received 30 November 2015Accepted 16 January 2016Available online 28 January 2016

Keywords:Diesel engineEGRDual-fuelEthanolUltralow NOx and smokeThermal efficiency

Exhaust gas recirculation (EGR) is a commonly recognized primary technique for reducing NOx emissionsin IC engines. However, depending on the extent of its use, the application of EGR in diesel engines isassociated with an increase in smoke emissions and a reduction in thermal efficiency. In this work,empirical investigations and parametric analyses are carried out to assess the impact of EGR in attainingultra-low NOx emissions while minimizing the smoke and efficiency penalties. Two fuelling strategies arestudied, namely diesel-only injection and dual-fuel injection. In the dual-fuel strategy, a high volatilityliquid fuel is injected into the intake ports, and a diesel fuel is injected directly into the cylinder. Theresults suggest that the reduction in NOx can be directly correlated with the intake dilution caused byEGR and the correlation is largely independent of the fuelling strategy, the intake boost, and the engineload level. Simultaneously ultra-low NOx and smoke emissions can be achieved at high intake boost andintake dilution levels in the diesel-only combustion strategy and at high ethanol fractions in the dual-fuelstrategy. The efficiency penalty associated with EGR is attributed to two primary factors; the combustionoff-phasing and the reduction in combustion efficiency. The combustion off-phasing can be minimized bythe closed loop control of the diesel injection timing in both the fuelling strategies, whereas the combus-tion efficiency can be improved by limiting the intake dilution to moderate levels. The theoretical andempirical analyses are summarized and the control of intake dilution and in-cylinder excess ratio isdemonstrated for the mitigation of NOx and smoke emissions with minimum efficiency impact.

� 2016 Elsevier Ltd. All rights reserved.

1. Introduction

The exhaust gas recirculation (EGR) is achieved by redirecting afraction of the exhaust gases into the intake. The resulting dilutionof the intake charge causes a significant reduction in the engine-out NOx emissions [1–3]. Over the years, the EGR rates appliedin diesel engines have increased consistently with the more strin-gent NOx emission targets. Likewise, a number of EGR configura-tions have been investigated [4,5]. Until recently, the mostcommon EGR configuration has been the high pressure EGR sys-tem. The high pressure EGR path connects the upstream of thevariable geometry turbine (VGT) to downstream of the compressor,and the VGT vanes maintain a positive backpressure to drive theexhaust gases through the EGR path. In order to accommodatehigher EGR flow rates without compromising the turbocharger

efficiency, the recent trend has been to implement a dual loopEGR configuration [6,7], where a low pressure EGR path is addedto the existing high pressure EGR system. In the low pressureEGR path, the treated exhaust gas from downstream of the tur-bocharger is recirculated into the fresh air, prior to entering thecompressor.

Although high EGR rates are common in current diesel enginesfor NOx reduction, the smoke penalty associated with EGR remainsa challenge (the NOx–smoke trade-off) [2,3,8]. The high smokeemissions can be partly mitigated by advancements in the com-mon rail fuel injection system. Research suggests that higher injec-tion pressures and smaller nozzle diameters improve the fuelatomization process and result in lower smoke emissions.Increased intake boost pressure has also shown advantages in low-ering the smoke emissions. However, the advancements in the fuelinjection and intake boosting systems cannot completely eliminatethe NOx and smoke trade-off [9,10]. At high engine load levels, inparticular, the implementation of EGR for NOx control is associatedwith an accelerated increase in the smoke, which limits the maxi-mum allowable EGR rates at the high load conditions.

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Nomenclature

AcronymsCA50 crank angle of 50% heat releaseCO carbon monoxideCI compression ignitionDPF diesel particulate filterEGR exhaust gas recirculationFSN filter smoke numberHCCI homogenous charge compression ignitionHTC high temperature combustionIDL intake dilution levelIMEP indicated mean effective pressureLTC low temperature combustionMAF mass air flowNOx oxides of nitrogenSCR selective catalytic reductionTDC top dead centerTHC total hydrocarbons

SymbolsC1HbOc equivalent fuel molecular formulaEGRCO2 CO2 EGR ratio

EGRmass mass EGR ratioEGRO2 O2 EGR ratiomcyl mass of cylinder chargemegr mass of recirculated exhaust gasesnCO2 in-cylinder moles of CO2nf in-cylinder moles of equivalent fuelnH2O in-cylinder moles of H2OnN2 in-cylinder moles of N2nO2 in-cylinder moles of O2½O2�amb ambient O2 concentration½O2�int cylinder O2 concentration½O2�exh exhaust O2 concentrationR molar EGR ratiox total mole number of combustion productsy total mole number of in-cylinder gas before combustionz total mole number of fresh air inducted per cyclekair fresh air excess ratiokcyl in-cylinder excess ratiow port-injection fuel energy ratio

370 P.S. Divekar et al. / Energy Conversion and Management 112 (2016) 369–381

Several researchers have studied the fundamental effects of EGRthat influence the combustion process and its impacts on the NOxand smoke emissions. Ladommatos et al. [1] and Zheng et al. [2]conducted experiments on diesel engines using mixtures of purebottled gases to study the EGR composition effects. Similar studieshave been conducted by Li et al. [11] on a natural gas spark ignitionengine. EGR leads to a reduction in the oxygen concentration of thecylinder charge by replacing a fraction of the air with CO2 and H2O.The dilution of the intake charge causes the largest reduction inNOx. However, the reduced in-cylinder oxygen is usually associ-ated with the increase in smoke emissions. Additionally, the higherheat capacities and the endothermic dissociation of CO2 and H2Ocontribute to the NOx emission reduction, albeit to a lesser extentcompared to the dilution effect. The increased intake charge tem-perature from the mixing of the hot exhaust gases with the freshair can cause a ‘thermal throttling’ effect that reduces trapped gasmass, negatively impacts the thermal efficiency and increases thesmoke emissions.

The NOx and smoke trade-off associated with the application ofEGR has required the use of exhaust after-treatment for meetingthe current (and future) emission standards [12]. Most on-roaddiesel engines are equipped with a diesel particulate filter (DPF)for smoke filtration and a selective catalytic reduction system(SCR) for the NOx reduction. The DPF requires a periodic injectionof fuel for regeneration of the filter and the SCR system employsurea injection to convert the NOx into N2 and O2. The consumptionof the supplemental fuel and the urea solution can be reduced ifthe trade-off between the engine-out NOx and smoke emissionscan be minimized. This will require a precise control of the EGRamount under all operating conditions including transient opera-tion [6]. The implementation of online EGR models [13], analyticalapproaches [14], and high speed measurements [15] have beeninvestigated to enable the precise EGR control.

Ultra-low levels of engine-out NOx and smoke emissions can beattained through the implementation of low temperature combus-tion (LTC) strategies, wherein the fuel–air mixing is enhanced byusing early or late injection timings or by excessive dilution. Byemploying LTC strategies, the dependence on aftertreatment

systems can be greatly reduced. Although LTC has been studiedextensively at steady-state conditions and moderate engine loads[16–21], the transient and high load operation in the LTC regimecontinues to be a challenge [22]. The air-path control exhibits amajor hurdle in the LTC implementation. Precise control of boostand EGR is necessary to deliver the simultaneously high intakeboost levels and EGR rates necessary for the diesel LTC operation[23].

The interest in renewable alternate fuels has driven theresearch towards the combustion of bio-fuels [24,25] in dieselengines. Studies conducted with ethanol [26], butanol [27], andvarious biodiesels [28] have shown that EGR is necessary attainthe low NOx emissions. The oxygen molecule contained in thesefuels lessens the negative effect of EGR on the smoke emissions.Although the effects of EGR have been extensively studied usingthe alternate fuels, the impact of the fuel type on the EGR effective-ness has not been highlighted. Under a fixed equivalence ratio, theexhaust gas concentrations largely depend on the fuel molecularcomposition. Hence, it is necessary to assess the impact of differentfuels on EGR.

The interactions among the EGR rates, fuelling amounts, fueltypes and the resultant emissions are highly complex. For a fixedrate of EGR, e.g. 30%, the impact on the NOx and smoke productionis altered by the engine operating conditions, as shown in Fig. 1.The intake oxygen concentrations and NOx–smoke emissions forthree test cases, representative of the diesel-only and the dual-fuel combustion strategies are summarized. In case I and case II,the diesel-only strategy is employed at low and high engine loads,respectively. The 30% EGR is less effective in reducing the intakeoxygen concentration, and results in high NOx emissions in caseI. However, in case II, the same EGR rate causes a much largerreduction in the intake oxygen concentration, which in turn lowersthe NOx emissions and results in high smoke emissions. In case III,the same engine load level as that of the case II is achieved in thedual-fuel strategy by port injection of ethanol and direct injectionof diesel. Although the same EGR rate is applied and a similarintake oxygen concentration is achieved, the NOx and smoke emis-sions are considerably lower in the dual-fuel case. The port injected

3.7

0.09

18.2

15.6

Dual-fuel operationsupresses smoke

while EGR effectiveness in NOx reduction

increases

EGR effectiveness in NOx reduction is improved at

high load, but smoke penalty is significant

30% EGR is insufficient for effective NOx

reductionCase IIMEP: 5.2 barDiesel only

Case IIIMEP: 15.6 bar

Diesel only

Case IIIIMEP: 15.9 bar

Dual fuelDiesel+Ethanol

0.009

1.4

15.2

0.007

0.34

Smoke [g/kWh]

NOx[g/kWh]

[O2]int[%]

EGR: 30% for all test cases

Fig. 1. Illustration of the EGR impact on the intake oxygen concentration and on theexhaust smoke and NOx emissions.

1 Compressed air supply

2 Intake surge tank3 Exhaust surge

tank4 Exhaust back-

pressure valve

5 EGR valve6 EGR cooler7 Diesel injector8 Port fuel injector9 Intake gas sampling

10 Exhaust gas sampling11 Cylinder pressure

transducer12 Optical encoder

1

6 5

2 3

8

1179 4

10

12

Portfuel

Fig. 2. Schematic of the experimental test setup.

Table 1Test engine specifications.

Model 4 Cylinder, DI Ford DuraTorq ‘‘Puma”

Displacement 1998 cm3

Bore � Stroke 86 mm � 86 mmCompression ratio 18.2:1Maximum cylinder pressure �18 MPa (180 bar)Injection systems Common-rail (up to 160 MPa)

P.S. Divekar et al. / Energy Conversion and Management 112 (2016) 369–381 371

ethanol supresses the smoke, while the NOx is also lower than thediesel-only case.

In this work, an analytical method is presented for the study ofEGR based on a molar balance across the air-path and a simple in-cylinder combustion chemistry calculation. Expressions arederived for the air–fuel ratio considering the fresh air and the in-cylinder gases. The analysis is then extended to incorporate theeffects of dual-fuel combustion. Thereafter, test results are pre-sented to develop an understanding of the EGR effects on theNOx and smoke emissions at different engine loads and differentfuelling strategies. The impact of EGR application on thermal effi-ciency is studied for both the diesel-only and dual-fuel modes.Based on the EGR analysis and the experimental study, fuel strat-egy independent EGR control considerations are laid out withlow NOx and smoke as the primary targets and thermal efficiencyas a constraint.

2. Experimental method

The test data presented in this work is collected from a singlecylinder, common-rail diesel engine. The engine is coupled to aneddy current dynamometer used for load dissipation and enginespeed control. The test platform is equipped with independentlycontrolled intake boost, EGR, exhaust back-pressure and fuelling(port injection and direct injection) systems. Intake boost isachieved using an oil-free dry air compressor to generate a highpressure combustion air supply and an electro-pneumatic valveto regulate the boost pressure. An air flow meter measures theair volume flow rate which is converted into a mass flow rate usingthe measured air temperature and pressure. An intake surge tank ismounted between the flow meter and the intake manifold to iso-late the cyclic pulsations generated by the valve events that wouldotherwise introduce errors in the flow rate measurement. Exhaust

back-pressure is adjusted using a pneumatically controlled valve.An EGR valve is used in combination with the exhaust back-pressure valve for EGR flow regulation. An exhaust surge tank,mounted in the exhaust path, ensures stable EGR flow and mini-mizes the impact of exhaust pressure wave action. A schematicof the test setup is shown in Fig. 2 and the major specificationsof the test engine are summarized in Table 1.

A high pressure diesel common rail fuel injection system is usedfor direct injection of diesel. The intake manifold is fitted with asecondary port fuelling system for the port injection of the highvolatility fuel. Two pre-calibrated gasoline port fuel injectors sup-ply the fuel into the intake runners. During the diesel-only tests,the common rail system is used to deliver the diesel fuel directlyinto the combustion chamber. For the dual-fuel tests, three highvolatility fuels, ethanol, butanol, and gasoline, are tested for port-injection and diesel is used as the direct injection fuel. The mainproperties of the tested fuels are listed in Table 2. The fuel flowrates are measured using volumetric fuel flow meters. Thesteady-state flow measurements are averaged over 60 s, whichare then converted to an average fuel mass flow rate by applyinga fixed fuel density conversion using the fuel densities summarizedin Table 2.

The cylinder pressure indicating system consists of a glow plugmounted pressure transducer (AVL GU13P) and a crank shaftmounted optical encoder with a 0.1� CA resolution. The intakeand exhaust gas compositions are measured using a dual-bankgas analyzer system, one for the exhaust emission measurements(NOx, HC, CO, CO2, O2 and smoke) and the other for the intakegas concentration measurements (CO2 and O2). The details forthe gas analyzer system are presented in Table 3. During everysteady state engine test point, cylinder pressure is recorded for200 consecutive engine cycles and all other measurements arelogged at 2 Hz and averaged over a period of 10 s.

It is noted that the results presented in this paper are subject tosmall variations in the measurement accuracies of the equipment

Table 2Test fuel specifications [29–31].

Fuel Diesel Ethanol Butanol Gasoline

Density (15 �C, kg/m3) 846 788 810 720Viscosity (30 �C, cSt) 3.5 1.52 3.5 0.64Cetane number (–) 46.5 8–11 17–25 10–17Octane number (–) �25 110–115 87 91Lower heating value (MJ/kg) 43.5 26.9 33.1 42.4Oxygen content (% mass) 0 34.78 21.6 NegligibleBoiling Temp (1 bar, �C) 246–388 78.3 117.5 60–200Equivalent molecular formula (–) C1H1.78 C1H3O0.5 C1H2.5O0.25 C1H1.87

Table 3Emission analyzers for emission and gas concentration measurements.

Analyzer type Measured emissions Model

Paramagnetic O2 (%) CAI 602PHeated flame ionization THC (ppm) CAI 300M HFIDNon-dispersive infrared CO (ppm). CO2 (%) CAI 200/300 NDIRChemiluminescence NOx (ppm) CAI 600 HCLDSmoke meter Smoke (FSN, mg/m3) AVL Model 415S

372 P.S. Divekar et al. / Energy Conversion and Management 112 (2016) 369–381

used. Although the individual equipment is calibrated regularlyand the linearity and measurement drift for the device is typicallywithin 1%, the derived measurements presented here may exhibitan uncertainty of up to 5%. For further details of the test setup andthe measurements, the reader is referred to Asad et al. [32].

3. EGR analysis

In this work, the concept of EGR is revisited from a fundamentalunderstanding stand point. The five primary components of thecylinder charge are considered, namely, N2, O2, CO2, H2O, and anequivalent fuel. Concentrations of the primary components in theintake and exhaust vary depending on the intake boost, EGR rate,fuel type, and fuel quantity. Although the by-products of combus-tion, such as CO, HC, NOx and smoke, are crucial from the emissioncontrol perspective, the concentrations of these combustion prod-ucts are usually at negligible levels for EGR ratio calculations.Therefore, they are not included in this analysis.

3.1. Analytical approach

Using the five components of the cylinder charge, the combus-tion reaction is written in the following form for an equivalent fuel,C1HbOc.

nf ðC1HbOcÞ þ nO2O2 þ nN2N2 þ nH2OH2O

þ nCO2CO2��!ðnCO2 þ nf ÞCO2 þ nH2O þ b2nf� �

H2Oþ nN2N2

þ nO2 þc2nf � nf � b4nf

� �O2 ð1Þ

The intake charge comprises a mixture of N2, O2, CO2, and H2O.The fuel is added to the mixture, which after combustion yields theproducts that consist of the same gaseous components but in vary-ing concentrations. The notation ‘ni’ denotes the mole number ofthe respective specie ‘i’ before combustion. The total mole numberof the intake charge is represented by ‘y’, while that of the productsis denoted by ‘x’. The total number of moles of products in Eq. (1)can be expressed in terms of the total number of moles of the reac-tants, as shown in Eq. (2).

x ¼ yþ c2þ b4

� �nf ð2Þ

In this work, a conceptual EGR ratio is used to express the EGRas a volumetric fraction of the total cylinder charge. Assuming thefresh air and the circulated gas mix under isothermal conditions,the volumetric EGR fraction is equivalent to the molar EGR frac-tion. Thus, the molar EGR ratio (R) can be written as,

R ¼ y� zy

ð3Þ

where ‘z’ is the mole number of fresh air inducted into the cylinder.

3.2. Quantification of EGR amount

Several methods for the evaluation of the EGR amount havebeen implemented in academia and industry such as the one notedin Eq. (3). The ability to quantify the EGR amount largely governsthe development of further understanding of the EGR impact.The frequently used EGR definitions are, therefore, summarizedin this sub-section and their equivalence is shown.

The mass based EGR is the most commonly applied definition intheoretical [33] and control studies [34,35]. The mass EGR ratio(EGRmass) is the ratio of the recirculated gas mass to the total cylin-der charge mass. The fuel mass may or may not be included in thecylinder charge mass depending the fuel delivery strategy. In thisstudy, the fuel mass is not considered as a part of the gaseouscylinder contents. Nonetheless, the consideration of fuel mass asa part of the cylinder flow introduces a minor discrepancy in themass EGR ratio calculations. The mass EGR ratio is expressed as,

EGRmass ¼ megrmcyl ð4Þ

where ‘megr’ is the mass of recirculated exhaust gases and ‘mcyl’ isthe cylinder charge mass. Although the mass based definition isuseful for theoretical analysis, calculation of EGRmass from practicalmeasurements is extremely challenging due to the lack of accurateand robust EGR flow measurements. However, the mass based EGRratio can be derived from the estimation of EGR flow (orifice flowacross the EGR valve) and cylinder flow (assuming a volumetric effi-ciency), and this approach is commonly used in control models [36].The difference between the molar EGR ratio in Eq. (3) and the massEGR ratio is negligible under the following assumptions;1. The molecular weight of the intake and exhaust gases is equal2. Molar concentration of fuel is very low (

P.S. Divekar et al. / Energy Conversion and Management 112 (2016) 369–381 373

The O2 based EGR ratio is defined as the ratio of the O2 concen-tration reduction in the intake charge to the O2 concentrationreduction in the exhaust gases, relative to the ambient O2concentration.

EGRO2 ¼½O2�amb � ½O2�int½O2�amb � ½O2�exh

ð6Þ

In Eq. (6), ½O2�amb, ½O2�int, and ½O2�exh are the ambient, intake, andexhaust O2 concentrations, respectively. If wet CO2 and O2 concen-trations are considered for the EGR ratio calculations, the EGRCO2and EGRO2 definitions are equivalent to the molar EGR ratio (R).

3.3. Intake dilution level

Even though the EGR ratio quantifies the displacement of thefresh air with the recycled gases, the consideration of the effective-ness of EGR towards emission reduction is of utmost importance.Since EGR is primarily used to reduce the NOx emissions in dieselengines, the effectiveness of EGR can be gauged by the extent ofNOx reduction attained by the application of EGR. Severalresearchers have studied the individual effects of the componentsof EGR-diluted intake charge on the NOx emissions using bottledlaboratory gases [1–3,8]. The results suggest that the partialreplacement of oxygen from ambient air yields the most significantNOx reduction. Thus, an intake dilution level (IDL) is defined hereto represent the EGR effectiveness using the O2 concentrations ofthe ambient air and the engine intake as shown in Eq. (7). Essen-tially, the IDL is the concentration ratio between the O2 replace-ment and the ambient O2.

IDL ¼ ½O2�amb � ½O2�int½O2�ambð7Þ

From the EGR analysis method summarized in Section 3.1, theIDL can be correlated to the molar EGR ratio, ‘R’, by Eq. (8).

IDL ¼ R 1� ½O2�exh½O2�amb

� �ð8Þ

Since the ambient O2 concentration is nearly a constant in mostcases, the IDL is a function of the molar EGR ratio, ‘R’, and theexhaust oxygen concentration. The correlation between the EGRratios and the IDL is presented in Fig. 3 for two engine operatingconditions. The upper graph shows the calculated molar EGR ratio,‘R’, the calculated EGRmass and the measured EGRCO2 plotted againstthe IDL for two engine load levels in the diesel combustion mode.

0

20

40

60

80

0

102030

4050

0 20 40 60Intake dilution level [%]

EGR

ratio

[%]

MAF EGRCO2 EGRR

Speed: 1500 rpmpint: 2 bar a

12 bar IMEP

6 bar IMEP

Speed: 1500 rpmpint: 2.5 bar aIMEP: 15 bar

Diesel+Ethanol

Diesel

EGR

ratio

[%]

Fig. 3. Relation between the intake dilution level and the EGR ratio at differentengine load and fuel combinations.

The EGRmass is calculated from the steady-state measurement ofthe fresh air mass flow rate and an estimation of the cylinder flowassuming a constant volumetric efficiency. As the fuel consump-tion increases at a higher engine load, more oxygen is consumedduring combustion, hence reducing the oxygen in the exhaustand in the EGR stream. As a result, the same EGR ratio causes agreater IDL in the higher load test case.

The lower part of Fig. 3 shows two test cases under the sameengine load and intake boost conditions but with different fuellingstrategies. A fuel energy ratio of 1:4 (diesel:ethanol) is used in thedual-fuel strategy. Comparing the two cases, a very similar EGR-IDL correlation is observed regardless of the different fuellingstrategies. The dual-fuel combustion requires a slightly lowerEGR ratio for the same IDL due to the higher fuel quantity requiredfor producing the same indicated load level.

3.4. Air–fuel ratio considerations

The air excess ratio defined by the actual air–fuel ratio relativeto the stoichiometric air–fuel ratio is commonly used to evaluatethe strength of the cylinder charge. However, this concept needsto be revisited when the intake air is diluted with the recirculatedexhaust gases. In order to address the effect of EGR on the air–fuelratio, the authors have defined two excess ratio terms in the previ-ous work [37]. The excess ratio based on the flow of ambient air iscalled the fresh air excess ratio (kair) while the excess ratio based onthe cylinder charge is called the in-cylinder excess ratio (kcyl). Thein-cylinder excess ratio takes into account the recirculation ofthe oxygen through the EGR path and hence is greater than thefresh air excess ratio for lean combustion. Using the EGR analysisadopted in this work, the fresh air excess ratios can be expressedas follows,

kair ¼ 1þ Cf ½O2�exh½O2 �amb�½O2 �exh½O2 �amb

ð9Þ

Similarly, the in-cylinder excess ratio is,

kcyl ¼ 1þ Cf ½O2�exh½O2 �int�½O2 �exh½O2 �int

ð10Þ

where ‘Cf’ is a constant for a given fuel depending on the equivalentfuel formula as shown in Eq. (11).

Cf ¼c2 þ b4� �1þ b4 � c2� � ð11Þ

Fig. 4. The correlation between the fresh air excess ratio and the in-cylinder excessratio at varying EGR rates and engine load levels.

Fig. 5. NOx reduction versus intake dilution ratio for single injection dieselcombustion.

374 P.S. Divekar et al. / Energy Conversion and Management 112 (2016) 369–381

As illustrated in Fig. 4, the fresh air excess ratio and the in-cylinder excess ratio are calculated from empirical test data of fourEGR sweeps at different engine load levels. The excess ratios arecalculated using the measured O2 concentrations in the intakeand the exhaust gases (following Eqs. (9) and (10)). The color ofthe solid dots represents the engine load level. The in-cylinderexcess ratio is higher than the fresh air excess ratio for all testpoints due to the recirculation of the O2 in the exhaust throughthe EGR path. At the same engine load, the difference betweenthe in-cylinder excess ratio and the fresh air excess ratio peaks atintermediate EGR levels. Under low EGR conditions, the cylindercharge primarily consists of the fresh air, and therefore the differ-ence between the two excess ratios is insignificant. At high EGRlevels, the combustion approaches stoichiometric conditions andthe O2 amount in the EGR stream depletes. As a result, the excessratio diminishes despite the high EGR flow rate. An extreme caseis the stoichiometric burning condition caused by high EGR usewhere no O2 is present in the EGR stream and the fresh air excessratio is equal to the in-cylinder excess ratio.

The IDL and the excess ratio expressions are useful to correlateengine emissions and efficiency with the effectiveness of EGR fordifferent fuelling strategies, as discussed in later subsections. Theexpressions for the IDL and the excess ratios presented in Eqs.(8)–(10) can be easily applied in dual-fuel combustion by adaptingthe fuel formulas and fuel energy ratio as presented in Eq. (A.1) inAppendix A. However, it must be highlighted that the equivalentmolecular formula for diesel and gasoline fuels is solely an approx-imate representation and does not account for the actual composi-tion of the fuels. As a result, the molar analysis is not an accuraterepresentation of the combustion of these multi-component fuelseven though it aids in the understanding of the EGR impacts. Adetailed discussion of the modification of the molar analysis tech-nique to accommodate the dual-fuel combustion is shown inAppendix A.

4. EGR versus NOx and smoke emissions

Since the composition of the recirculated gases is dependent onthe amount of fuel, air, and EGR, it is necessary to analyze theimpact of EGR on the emissions under a wide range of engine oper-ating conditions. The data presented in the following sections isselected from a large pool of engine tests over a wide range ofengine operating conditions and the trends are representative ofthe entire dataset.

4.1. Single injection diesel combustion

The conventional diesel high temperature combustion (HTC) ischaracterized by a short ignition delay with (typically) an overlapbetween the fuel injection and the combustion events, which pro-vides effective controllability over the combustion process buttends to produce high NOx emissions. EGR is therefore applied todilute the cylinder charge and achieve a significant NOx reduction,however usually with an associated smoke penalty. A simultane-ous lowering of the NOx and smoke emissions can be achieved ifan adequate fuel–air mixing time is allowed by increasing the igni-tion delay to enter into the partially premixed LTC regimes [16,17].A long ignition delay can be attained by using combustion phasingretard and high amounts of EGR. In order to maintain adequatecombustion stability and to use the thermal efficiency benefitsfrom a large expansion ratio, the phasing retard is normally limitedto 15–25 �CA after TDC, the exact value depending on the engineload and other operating conditions [21,38,39]. Therefore, theenabling of simultaneously low NOx and smoke emissions relieson heavy EGR application.

The intake dilution level through EGR is a dependent on theengine load and the intake boost at a given EGR ratio. The extentof NOx reduction caused by EGR is also dependent on the engineoperation. The authors have previously shown that the effective-ness of EGR in reducing NOx can be decoupled from other engineoperating conditions if the intake dilution is used as a measureof the EGR instead of the EGR ratio itself [37]. Thus, the IDL isthe preferred parameter from the NOx control perspective. InFig. 5, the reduction in the NOx emissions is plotted against theIDL for test data collected over a wide range of engine operatingconditions. The upper plot presents the measured NOx againstthe dilution level, whereas in the lower plot the same data set ispresented as the NOx reduction ratio normalized to the NOx levelsat zero EGR condition for each individual case. The representationhighlights two important trends;

a. EGR is very effective in NOx emission reduction up to an IDLof around 25–30%. Thereafter, the effectiveness reduces sub-stantially highlighting the challenge of achieving ultra-lowNOx emissions.

b. Although different levels of EGR are necessary to achieve thedesired IDL, all the test data sets tend to overlay highlightingthe sensitivity of NOx emissions to the IDL. It must be notedhowever, that the actual value of NOx emissions at 0% EGRvaries with engine operating conditions, and the EGR effectis only relative to the base NOx emission value.

While the NOx emissions depict a straightforward trend withthe IDL, the smoke is affected by a number of engine operatingvariables. From the air-path point of view, a major factor thatimpacts the diesel engine smoke emissions is the cylinder excessratio which is a function of the intake boost, EGR ratio and the fuel-ling quantity (as shown in Eq.(10)). Since the fuelling quantity isprimarily determined by the user torque demand, the control overthe in-cylinder excess ratio can be exercised by a combination ofintake boost and EGR ratio regulation.

Fig. 6. Effect of intake boost on smoke emissions and dilution ratio. Fig. 7. Effect of fuelling quantity on smoke emissions and dilution ratio.

1 For interpretation of color in Figs. 4, 5, 6, 9 and 15, the reader is referred to theweb version of this article.

P.S. Divekar et al. / Energy Conversion and Management 112 (2016) 369–381 375

The smoke emissions from two EGR sweep tests are presentedagainst the in-cylinder excess ratio in the upper plot of Fig. 6.The fuelling amount and the fuel injection pressure are held con-stant while the intake boost pressure is increased from 1.75 barto 2 bar abs. As shown by the test results, the lower boost caseexhibits lower in-cylinder excess ratios throughout the EGR sweep.When the intake pressure is 1.75 bar abs, the in-cylinder excessratio is lower throughout the EGR sweep test. At kcyl = 1.4, thesmoke emissions approach as high as 5 FSN which prohibits fur-ther increase in the EGR amount. However, when the intake boostpressure is increased to 2 bar abs, the EGR sweep curve shifts to aleaner in-cylinder excess ratio, and the EGR can be furtherincreased without producing excessive smoke, while avoiding theregions of very high smoke. Following the initial increase in theEGR amount, the reduced oxygen availability and the loweredcombustion gas temperatures negatively impact the oxidation rateof the smoke formed in the diesel flame, resulting in increasedexhaust smoke emissions. However, when EGR is very high, thediluted cylinder charge supresses the ignition of diesel such thatthe partially premixed LTC mode is enabled and the formation ofsmoke is avoided [21].

In the lower plot of Fig. 6, the test data is overlaid on contours ofthe intake boost and in-cylinder excess ratio in order to betterunderstand the impact of the intake boost level and the EGR rateon the achievability of LTC. The contours are generated using theEGR analysis discussed in Section 3.1. The shaded contours repre-sent the IDL against the intake boost and the cylinder excess ratio.Contour lines of fixed EGR ratio are also included in the plot, and

are represented by thick blue1 lines. The reduction of NOx requiresa sufficient IDL (�30%). However, the desired intake dilution canreduce the in-cylinder excess ratio towards near stoichiometriclevels when low boost is applied, e.g. in the 1.75 bar abs test case.The resultant lack of the excess in-cylinder oxygen causes the smokeemissions to increase rapidly. When the intake boost pressure is ele-vated, a similar intake dilution is attained at a higher in-cylinderexcess ratio resulting in an overall reduction in the smoke. Moreover,the higher in-cylinder excess ratio at the same IDL allows furtherEGR increase which ultimately enables the LTC operation.

The impact of increasing the engine load (fuelling amount) onthe exhaust smoke emissions is presented in Fig. 7. The smokeemissions are plotted against the cylinder excess ratio at threeengine load levels in the upper part of Fig. 7 for fixed intake boostand fuel injection pressures. At the low load condition (i.e. the6 bar IMEP test case), the cylinder charge is over lean and as theEGR is increased, the in-cylinder excess ratio decreases causingthe smoke to increase. The peak smoke, however, is relativelylow as a result of the sufficiently large amount of in-cylinder oxy-gen. With further EGR application, LTC is enabled during which thein-cylinder excess ratio remains sufficiently lean (kcyl � 2). Whenthe fuelling amount is increased to achieve 10 bar IMEP, the EGRsweep curve shifts towards a richer air–fuel ratio operation.Consequently, the smoke emissions increase more rapidly and alarger smoke peak is observed compared to the low load test case.

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Nevertheless, the in-cylinder excess ratio is sufficiently lean at thesmoke peak such that additional EGR rates can be applied to enableLTC where sharp drop of smoke is observed. When the engine loadis further increased to 12 bar IMEP, the in-cylinder excess ratioapproaches 1.0 as the EGR ratio is increased. The lack of oxygen(low in-cylinder excess ratio) at the smoke peak prohibits furtherEGR application. It must be noted however, that other researchershave achieved LTC by operating in the fuel-rich region (smokelessrich combustion) [40]. Fuel rich operation is not attempted in thecurrent work in the interest of maintaining acceptable combustionefficiency and limiting the maximum smoke peak during thetransition.

In the lower plot of Fig. 7, the EGR sweep test data at the threeengine load levels is marked onto the IDL and EGR ratio contoursgenerated from parametric calculations. The IDL is presented inthe form of colored contours against the fuelling amount and thein-cylinder excess ratio. Lines of fixed EGR ratio are also overlaidon the chart. At lower fuelling rates, the in-cylinder excess ratiois higher and thus a larger intake dilution can be attained byincreasing the EGR amount. However, as the fuelling amountincreases, the window of lean operation shrinks rapidly, thusrequiring a lower air–fuel ratio to attain the same IDL. It may alsobe noted that at the lower fuelling rates, a significantly higher EGRratio is necessary to achieve a certain level of intake dilutionwhereas when the fuelling rate increases, the same IDL can beattained at a lower EGR ratio.

In Fig. 8, the indicated NOx and smoke emissions are plotted forthree EGR sweeps to demonstrate the NOx–smoke trade-off andthe enabling of LTC in the single intake diesel combustion strategy.To summarize the NOx–smoke trade-off and the enabling of LTC inthe single injection diesel combustion strategy, the indicated NOxand smoke emissions are plotted in Fig. 8 for three EGR sweeptests. The EGR sweeps are conducted at three engine load levelswherein the intake boost and fuel injection pressure are concur-rently increased at the higher engine load. The emissions data isplotted on a log scale for easy readability. Dotted lines of constantIDL are marked onto the data to highlight the test points with sim-ilar IDL. As highlighted in Fig. 6, the indicated NOx emissions arelargely insensitive to the load level but primarily depend on theIDL. The smoke emissions exhibit an increasing trend as the engineload increases, even though the intake boost and fuel injectionspressures are raised. In the 5 bar and 10 bar IMEP test cases, theIDL can be increased to enable LTC, whereas in the 16 bar test case,further increase in EGR is restricted excessive smoke emissions.(6 FSN at 28% IDL)

Fig. 8. NOx–smoke trade-off and LTC enabling for single injection dieselcombustion.

4.2. Diesel injection with port fuelling

Diesel fuel’s low volatility and its strong tendency for auto-ignition (high cetane number) limit the ability to enable partiallypre-mixed combustion at increased load levels. The dual-fuel strat-egy has been identified as a promising solution to extend theengine load level in the low NOx and smoke combustion regime.This strategy includes the injection of a high volatility, low cetanefuel at the intake port followed by the in-cylinder high pressureinjection of diesel fuel. Under the dual-fuel combustion strategy,the application of EGR is necessary for NOx emission reductionwhereas the pre-mixed portion of the cylinder charge may reducethe smoke penalty associated with the increase in the EGR amount.

EGR sweep tests conducted with three high volatility fuels inthe dual-fuel mode are presented in Fig. 9. The high volatility fuel(ethanol, butanol, or gasoline) is port-injected during the inductionstroke followed by the near top dead center (TDC) injection of die-sel. The ratio of the two fuels is adjusted such that 45–50% of thetotal fuel energy is contributed by the port injected fuel. Testresults for the diesel-only combustion configuration at the sameoperating conditions are also plotted for reference. Compared tothe diesel-only results, when diesel is partially replaced with theport injected fuel, NOx emissions decrease, even at 0% IDL.The impact of the dual-fuel strategy on NOx emissions canbe explained from two main aspects. A larger fraction of thecombustion is lean and pre-mixed in the dual-fuel mode causinga reduction in NOx emissions. In addition, the evaporation ofthe port-injected fuel during the compression stroke causes anoticeable reduction in the compression-end temperature, henceresulting in a lower flame temperature. The effect of the dual-fuel strategy on NOx emissions is the largest in the gasoline–dieseltest case among the cases presented in Fig. 9. Analysis of thecylinder pressure data reveals that the combustion duration inthe gasoline–diesel test case is the longest compared to the otherdual-fuel test cases. The long combustion duration contributes toa lower temperature rise which may result in the lower NOx.

As shown by the overall trend of all four presented cases, NOxemissions drop substantially as the IDL is increased at higherEGR rates. Therefore, the application of EGR is still the primaryenabler of ultra-low NOx emissions in either diesel-only or dual-fuel combustion. When the NOx emission reduction is presentedrelative to the peak NOx emissions at 0% IDL, all the test cases fol-low a similar trend. This highlights the correlation between intake

Fig. 9. NOx reduction versus intake dilution level for dual-fuel combustion.

Fig. 10. Smoke emission trends for dual-fuel combustion.

Fig. 11. Heat release traces for selected test points in the dual-fuel combustionmode.

Fig. 12. NOx–smoke trade-off and LTC enabling in diesel–ethanol, dual-fuelcombustion.

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dilution and NOx reduction, which can be applied to both thediesel-only and dual-fuel strategies.

The smoke emissions, however, do not exhibit a consistenttrend when different fuels are used in dual-fuel combustion. Whenapproximately 50% of the diesel fuel is replaced by the port injec-tion of ethanol, a noticeable reduction in the smoke emissions isobserved throughout the EGR sweep test. This effect however isnot seen when butanol or gasoline is used at a similar fuel ratio.In fact, with the application of EGR, the smoke emissions tend toincrease more sharply in the diesel–gasoline and diesel–butanoltesting cases (see Fig. 10).

The heat release rate traces for a representative case of each ofthe three dual-fuel combustion tests are shown in Fig. 11 toexplain the trends in the smoke emissions. Ethanol has a strongtendency to resist auto-ignition even on the high compression ratioengine due to its high octane number. Thus, the combustion of thepremixed ethanol-air charge is essentially initiated after the dieselfuel delivery has commenced. On the other hand, when butanol orgasoline is port injected, the low auto-ignition resistance of thefuel (relative to that of ethanol) causes a portion of the portinjected fuel to auto ignite prior to the diesel injection. Thus, the

ignition delay of the diesel is substantially shortened and the dieselundergoes diffusion type combustion, resulting in a significant risein the smoke emissions.

These test results suggest that with the high compression ratio,the dual-fuel combustion strategy using a relatively low octane(e.g. butanol or gasoline) still encounters the NOx–smoke trade-off. However, the auto-ignition of port injected fuels indicates thepossibility of enabling HCCI type combustion, which is beyondthe scope of this work and more details can be found in theauthors’ previous work [41,42].

Nevertheless, the test results of ethanol–diesel combustionhave clearly shown the benefits of implementing a dual-fuel strat-egy in addition to the EGR application for the reduction of smokeand NOx emissions. The results of the ethanol–diesel combustionat higher ethanol fractions are summarized in Fig. 12. Under thesame engine boundary conditions, EGR sweeps are performed atfour different fuel ratios. The NOx and smoke emissions are plottedon the log scale. The ethanol substitution not only supresses thetendency to produce smoke at elevated EGR rates, but the NOxemissions are also reduced as the ethanol fraction is increased sug-gesting that a lower EGR may be sufficient for achieving the sametarget NOx emissions. At very high ethanol fractions, e.g. 80%, thesmoke emissions remain ultra-low, and the increase in the EGRrate produces a consistent reduction in the NOx emissions.

5. EGR and thermal efficiency

Previous research has suggested that the use of EGR bears anegative impact on the specific fuel consumption in diesel engines[4,43,44]. The reduction in the thermal efficiency is attributed toseveral factors that can be broadly categorized as the combustionand system level factors. The pumping work associated with a pos-itive exhaust-to-intake manifold pressure difference (required todrive sufficient EGR across the EGR path) is the primary contribu-tor to the system level reduction in the efficiency. The combustionrelated factors, on the other hand, comprise of the reduced cyclework and deteriorated combustion efficiency. The reduction inthe cycle work is primarily caused by the retarded combustionphasing with EGR application if the diesel injection timing is fixed,whereas, the reduction in the combustion efficiency is caused bythe increase in the exhaust HC and CO emissions. Other factorssuch as prolonged combustion duration, lowered air-excess ratioand reduced combustion temperature also contribute to the effi-ciency loss. This work mainly focuses on the engine efficiency from

Fig. 13. Effect of EGR on combustion phasing control by diesel injection command.

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the combustion perspective as discussed in the subsequentsubsections.

Fig. 14. Effect of EGR on thermal efficiency for diesel combustion and diesel–ethanol dual-fuel combustion.

5.1. Combustion phasing impact on efficiency

The ignition delay is typically prolonged as EGR is increasinglyapplied. When the injection timing is fixed, the start of combustionpostpones in accordance to the longer ignition delay and the entirecombustion event can be delayed into the expansion stroke. Enginetest results are presented to demonstrate the combustion phasingretard caused by EGR application and its effect on the engine ther-mal efficiency. The results of two EGR sweeps are presented inFig. 13. These two test sets are conducted under the same engineconditions except the fuel injection timing. The CA50 and thermalefficiency are plotted against the NOx emissions so that the impactin achieving the NOx reduction is highlighted. The solid dots repre-sent the results where the diesel injection timing was held con-stant throughout the EGR sweep test, while the hollow dotspresent test results wherein the fuel injection timing was adjustedto maintain the CA50 within 1 �CA of the baseline CA50.

As the EGR rate is increased while the fuel injection timing isfixed, the ignition delay increases and the CA50 shifts later. Whenthe same EGR sweep test is repeated with a fixed CA50, majority ofthe thermal efficiency penalty is recovered. It is noted that a grad-ual reduction in the thermal efficiency is observed as EGR isincreased, even when the CA50 is fixed. The thermal efficiencyreduction is more prominent at high IDL applied to achieve NOxemissions lower than 0.2 g/kW h. This reduction in the thermalefficiency may be caused by the reduction in combustion effi-ciency, increase in combustion duration, a lower combustiontemperature.

5.2. Combustion efficiency impact

In the diesel-only combustion, the exhaust HC and CO emis-sions tend to increase sharply when heavy EGR is applied, whereasin the dual-fuel combustion these incomplete combustion prod-ucts remain at high levels across the EGR range. The increasedHC and CO emissions lead to lower combustion efficiency and inturn reduce the engine thermal efficiency. The EGR and dual-fuelstrategy impacts on the combustion efficiency are investigated inthe current section. Representative results from two EGR sweeptests are presented in Fig. 14. The diesel-only fuelling strategyand the ethanol–diesel strategy are tested at 10 bar IMEP. In boththe test cases, the CA50 is maintained at 368–369 �CA by adjustingthe diesel injection timing to counteract the combustion phasingshift from the use of EGR. The smoke emissions, combustion effi-ciency and thermal efficiency are plotted against NOx emissions.

In the diesel-only case, the initial reduction in NOx (from highlevel to 0.3 g/kW h) by increasing the EGR rate does not incur anoticeable penalty in the smoke or the combustion efficiency,although the thermal efficiency shows a minor decrease, consistentwith trends reported in Section 5.1. This decrease may beattributed to the marginal increase in combustion duration andthe reduction in the combustion temperature. A further increasein EGR reveals the NOx–smoke trade-off and the thermal efficiencycontinues to decrease. A gradual reduction in the combustionefficiency also contributes to the thermal efficiency reduction atthis point. As the intake dilution is increased further to attainsimultaneous ultra-low NOx and smoke, the thermal efficiencypenalty increases substantially because of a sharp rise in theincomplete combustion products. It is noted that ultra-low NOx(>0.2 g/kW h) can be achieved without a significant thermal effi-ciency penalty, but further EGR addition is necessary to attainultra-low smoke emissions in LTC where significant efficiencydegradation is observed.

Under the ethanol–diesel combustion, the NOx reduction byEGR can be obtained with a smaller rise in smoke. When a highethanol fraction is used, ultra-low smoke emissions can beachieved as a larger portion of the cylinder charge is already pre-mixed (recall Fig. 12). The high homogeneity also results inreduced combustion efficiency, as seen in Fig. 14. However, inthe dual-fuel mode, the thermal efficiency is marginally higherthan that of the diesel-only mode. The combustion inefficiency iscompensated by the thermal efficiency gain from the charge cool-ing effect of the port injected ethanol [45]. It may be noted thatalthough EGR is necessary for NOx abatement in both the combus-tion strategies presented in Fig. 14, the low smoke in the dual-fuelmode reduces the IDL necessary for achieving LTC (simultaneouslylow NOx and smoke). Thus the rapid reduction in thermal effi-ciency at LTC conditions is not observed in the dual-fuel strategy.

6. EGR control considerations

The engine test results presented in this work indicate that inthe diesel-only mode the NOx–smoke emission trade-off remainsa major challenge in deciding the applicable EGR amount. LTCmay be enabled with excessively high rates of EGR. This requires

Fig. 15. Contour map of intake boost pressure and MAF at varying engine loads anda fixed IDL.

Fig. 16. Engine full load demonstrated with diesel ignited ethanol LTC.

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significantly larger intake boost amounts to maintain sufficientlyhigh in-cylinder excess ratio while attaining the desired intakedilution. Furthermore, a large thermal efficiency penalty isobserved when excess amounts of EGR are applied to attain theLTC conditions. On the contrary, in the dual-fuel mode the dis-placement of diesel with the port injected ethanol supresses thesmoke emissions when a large ethanol fraction is used. Even inthe dual-fuel mode, sufficient intake dilution by EGR is necessaryfor NOx reduction and a high in-cylinder air-excess ratio isdesirable.

While the EGR flow displaces the fresh intake air, maintaining alean in-cylinder charge requires adequate control over the intakeboost and the EGR flow rate. As the EGR flow rate is typically notmeasured, it is necessary to control the intake boost and the freshmass air flow (MAF) rate. In this regard, a contour map of intakeboost and MAF is presented in Fig. 15. The map is generated fromthe parametric EGR analysis discussed in Section 3. The analysis isconducted at a fixed IDL (for the most effective NOx reduction,recall Figs. 5 and 9) over a range of boost, MAF and fuel amounts.The color contours represent the IMEP levels corresponding tothe increased fuelling amounts. Iso-lines of EGR (solid blue) andin-cylinder excess ratio (dotted orange) are overlaid on the samecontour plot.

The trends presented in Fig. 15 can be understood as follows.While maintaining a fixed IDL at a low engine load level (bottomright region), a high cylinder excess ratio can be attained byincreasing the intake boost level. However, the EGR rate necessaryfor this condition rapidly increases which may be achieved byincreasing the intake boost pressure and maintaining a fixedMAF. On the contrary, at the high load condition (top right region),the same IDL is achieved at a much lower EGR ratio. The intakeboost pressure and the MAF have to be increased simultaneouslyas the load level increases. If a fixed IDL is desired, in combinationwith a fixed in-cylinder excess ratio, a linear relationship isobserved between the intake boost level, fresh air mass flow rateand the fuelling amount. The boost, MAF and fuelling correlationis similar to the EGR and air excess ratio correlation presented byNakayama et al. [46]. While the EGR quantities necessary to main-tain a fixed dilution and an in-cylinder excess ratio may vary withthe engine load, a closed loop control over intake boost and thefresh air flow may be designed to achieve the NOx reduction with-out compromising the in-cylinder excess ratio.

A full load test point operated in the diesel–ethanol mode isoverlaid on the contour map to further explain the utility of such

a representation. The cylinder pressure and heat release trace forthis test point are presented in Fig. 16. The full load condition atthe desired IDL is achieved by operating close-to-stoichiometricconditions at 2.5 bar abs intake pressure. At this test condition, fur-ther increase in the intake boost is prevented by the peak cylinderpressure limit of the test engine.

The operating window for the fixed IDL shrinks as the engineload increases. Moreover, significantly higher levels of intake boostpressure may be necessary to maintain the dilution and in-cylinderexcess ratios at high engine load points. In reality, the ability toincrease the intake boost is limited by the turbocharging hardwareas well as the peak cylinder pressure limit of the engine. Neverthe-less, this EGR analysis provides guidelines for better operating theEGR and turbocharging systems when pre-defined limits for intakedilution and in-cylinder excess ratio are available.

7. Conclusions

Empirical studies are conducted to develop a comprehensiveunderstanding of the use of EGR for achieving ultra-low NOx emis-sions. The impact of the EGR rate on the engine performanceparameters is investigated for two fuelling strategies, the diesel-only strategy and the dual-fuel strategy. An EGR analysis is devel-oped to account for the different fuel molecular structures anddual-fuel applications. The research results are summarized asfollows:

1. Using the EGR analysis and test data, an engine operatingparameter IDL is defined as an indicator to gauge the effective-ness of EGR on NOx reduction independent of engine operatingconditions and fuelling strategies.

2. In the dual-fuel combustion mode, three port-injection fuels areinvestigated, ethanol, butanol, and gasoline, while a moderateamount of diesel e.g. 12–50% is used as the direct-injection fuel.Although lower NOx emissions are observed when a larger frac-tion of the port-injected fuel is used, EGR is still necessary toachieve the ultra-low NOx emissions.

3. The smoke emissions tend to increase as the intake dilutionlevel increases for both diesel-only and dual-fuel combustionstrategies. However, in the dual-fuel combustion, the ethanol–diesel combustion results in a reduction of this smoke penalty.

4. In addition to the conventional air excess ratio, the in-cylinderexcess ratio is used to explain the trends of the smoke emis-sions associated with the increase in the EGR. A larger in-cylinder excess ratio is beneficial for the lowering of the smokeemissions and for achieving LTC in the diesel-only combustionmode.

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5. The reduction in the thermal efficiency associated with theincrease in the EGR rate can be attributed to the combustionoff-phasing and the combustion efficiency degradation.

6. The combustion-off phasing caused by EGR increase can becompensated by adjusting the diesel injection timing in boththe diesel-only and the diesel–ethanol dual-fuel strategies.

7. At 10 bar IMEP, a heavy use of EGR is necessary to attain simul-taneously low smoke and NOx emissions in the diesel-onlycombustion mode, at which point the combustion efficiencydecreases sharply. In the diesel–ethanol dual-fuel combustionmode, heavy EGR application is not necessary as the smokeemissions are ultra-low when a high ethanol fraction is used.By avoiding the heavy use of EGR, the large reduction in com-bustion efficiency is avoided.

8. The intake dilution and in-cylinder excess ratio control byintake boost and fresh air flow regulation is suggested for theattainment of the desired NOx reduction with a smaller smokepenalty. The quantification of EGR in this can facilitate to over-come the challenges associated with the EGR flow adjustment.

Acknowledgements

The research is supported by NSERC CRD, Discovery, CREATEprograms; the NCE AUTO21 and BioFuelNet programs; the FordMotor Company, and the University of Windsor.

Appendix A. EGR analysis for dual-fuel mode

When the dual-fuel combustion mode is considered, EGR anal-ysis can be complicated due to the different fuel compositions andthe varying fuel quantities. The EGR analysis for the dual-fuel sce-nario can be greatly simplified by defining an equivalent hypothet-ical fuel C1HbOc. The hypothetical fuel produces the same molenumbers of the primary exhaust gas components when it replacesthe two test fuels.

The two fuels used for the current dual-fuel analysis are repre-sented as follows.

1. C1Hb1Oc12. C1Hb2Oc2

In the dual-fuel mode, an energy ratio, w, is used to evaluate therelative amounts of the two fuels used.

w ¼ nf2LHV2nf1LHV1 þ nf2LHV2 ðA:1Þ

where LHV1 and LHV2 (represented in energy per moles of fuel) arethe lower heating values of the primary and secondary fuels. Amolar fuel ratio, w0, is defined to evaluate the fuel mole numberratio of the secondary fuel to the total moles of fuel.

w0 ¼ nf2nf1 þ nf2 ðA:2Þ

Combustion for the dual-fuel mode can be written as,

nf1ðC1Hb1Oc1Þ þ nf2ðC1Hb2Oc2Þ þ nO2O2 þ nN2N2 þ nH2OH2Oþ nCO2CO2 ! ðnCO2 þ nf1 þ nf2ÞCO2þ nH2O þ

b12

nf1 þ b22 nf2� �

H2Oþ nN2N2

þ nO2 þc12

nf1 þ c22 nf2 � nf1 � nf2 �b14

nf1 � b24 nf2� �

O2 ðA:3Þ

By comparing Eq. (A.3) to Eq. (1), the following can be derived

nf ¼ nf1 þ nf2 ðA:4Þ

b ¼ b2 � w0 þ b1ð1� w0Þ ðA:5Þ

c ¼ c2 � w0 þ c1ð1� w0Þ ðA:6ÞUsing the equivalent fuel formulation, if the fuel ratio is known,

the EGR analysis approach developed in Section 3 including the airexcess ratio definitions apply to the dual-fuel mode.

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Energy efficiency impact of EGR on organizing clean combustion�in diesel engines1 Introduction2 Experimental method3 EGR analysis3.1 Analytical approach3.2 Quantification of EGR amount3.3 Intake dilution level3.4 Air–fuel ratio considerations

4 EGR versus NOx and smoke emissions4.1 Single injection diesel combustion4.2 Diesel injection with port fuelling

5 EGR and thermal efficiency5.1 Combustion phasing impact on efficiency5.2 Combustion efficiency impact

6 EGR control considerations7 ConclusionsAcknowledgementsAppendix A EGR analysis for dual-fuel modeReferences