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Experimentalandsimulationinvestigationofthecombustioncharacteristicsandemissionsusingn-butanol/biodieseldual-fuelinjectiononadieselengine

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Energy 74 (2014) 741e752

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Energy

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Experimental and simulation investigation of the combustioncharacteristics and emissions using n-butanol/biodiesel dual-fuelinjection on a diesel engine

Haifeng Liu a, Xin Wang a, Zunqing Zheng a, *, Jingbo Gu a, Hu Wang b, Mingfa Yao a

a State Key Laboratory of Engines, Tianjin University, Tianjin 300072, Chinab Engine Research Center, University of WisconsineMadison, Madison, WI 53705, USA

a r t i c l e i n f o

Article history:Received 31 December 2013Received in revised form24 June 2014Accepted 16 July 2014Available online 11 August 2014

Keywords:n-ButanolBiodieselDual-fuel injectionCombustionEmissions

* Corresponding author. Tel.: þ86 22 27406842x80E-mail address: zhengzunqing@tju.edu.cn (Z. Zhen

http://dx.doi.org/10.1016/j.energy.2014.07.0410360-5442/© 2014 Elsevier Ltd. All rights reserved.

a b s t r a c t

The combustion and emissions of n-butanol/biodiesel dual-fuel injection were investigated on a dieselengine based on experiments and simulations. n-Butanol was injected into the intake port, while soy-bean biodiesel was directly injected into the cylinder. Three different premixed ratios (rp) were inves-tigated, including 80%, 85% and 90%. The injection timings of biodiesel were adjusted to keep the 50%burn point (CA50) between 2� CA and 10� CA after top dead center for achieving stable operation. TheEGR (exhaust gas recirculation) rates were changed from 35% to 45%. Results demonstrate that the sameCA50 can be achieved by the early or late-injection of biodiesel. For both early- and late-injection, theauto-ignition is triggered by the biodiesel reaction. Increasing premixed ratios can retard the combustionphasing and reduce the pressure rise rate, while the indicated thermal efficiency (ITE) reduces by about0.6% as increasing rp to 90%. The early-injection has lower NOx emissions compared to the late-injectiondue to lower combustion temperature. The soot emissions are comparable for both early- and late-injection. With the increase of EGR, the NOx and soot emissions decrease, while the HC (hydrocar-bons) and CO (carbon monoxide) emissions increase. The ITE reduces by 1e2% as increasing EGR to 45%.

© 2014 Elsevier Ltd. All rights reserved.

1. Introduction

Diesel engines are widely used in transportation, agricultureand engineeringmachinery due to its reliability and high efficiency.However, they contribute significantly to carbon dioxide andharmful emissions and consume large amounts of fossil oil.Therefore, the development of diesel engines faces the challengesof energy and environment in the future. Advanced combustiontechniques and the application of biofuels are promising ways tomeet these challenges.

Biofuels derived from renewable resources are considered as thesustainable alternative to conventional fossil fuels [1,2]. At present,biodiesel is the primary alternative to diesel due to their similar fuelproperties [3]. Investigation on engine test demonstrated thatbiodiesel-fueled engines could reduce emissions of hydrocarbons(HC), carbon monoxide (CO) and particulate matter (PM), howevernitrogen oxides (NOx) emissions and the brake specific fuel con-sumption were increased slightly [4e6]. Meanwhile, the

13; fax: þ86 22 2738 3362.g).

fundamental spay and flame propagation of biodiesel fuel has alsobeen studied on constant-volume chamber to deeply understandthe mixing and combustion process at different ambient conditions[3,7,8]. Due to the wide application of biodiesel on engines, thechemical kinetic models have been developed rapidly in past tenyears. The whole development trend is from small alkyl esters (upto 5 carbons) to large alkyl esters (up to 19 carbons) and from singlesurrogate to a mixture of surrogates [9,10]. For example, the kineticmodel of small alkyl ester of methyl butanoate (C5H10O2) wasdeveloped by Fisher et al. [11]. After that, Metcalfe et al. [12] andDooley et al. [13] improved the kinetic model of methyl butanoate.Then, with the further development on models, large alkyl estershave been developed in recent years, such as methyl hexanoate(C7H14O2) [14], methyl decanoate (C11H22O2) [15], methyl stearate(C19H38O2) andmethyl oleate (C19H36O2) [16]. Apart from the singlesurrogate, some mixture have been proposed to represent the realbiodiesel, such as the blends of n-heptane and methyl butanoate[17] and the mixture of methyl stearate (C19H38O2), methyl oleate(C19H36O2), methyl linoleate (C19H34O2), methyl linolenate(C19H32O2) and methyl palmitate (C17H34O2) [18]. Another class ofbiofuel used in diesel engines is alcohols, such as ethanol and n-butanol [19,20]. Studies indicated that the blends of alcohol and

H. Liu et al. / Energy 74 (2014) 741e752742

diesel could reduce emissions of NOx, soot and CO simultaneously[21e23], while HC emissions the brake specific fuel consumptionwere increased [24,25]. In comparison with ethanol, n-butanol is amore promising alcohol for application in diesel engines due to itshigher heating value, good intersolubility with diesel and nocorrosion to existing fuel pipelines [26,27]. Due to the above ad-vantages in fuel properties, the fundamental spay and flamepropagation of the blend of n-butanol and diesel fuel has also beenstudied on constant-volume chamber and optical engines to deeplyunderstand the mixing and combustion process [28,29]. Mean-while, many different chemical kinetic model of n-butanol has beenproposed in recent years, including Refs. [30e32], and the kineticmodels of diesel and n-butanol blends have also been developed bySaisirirat et al. [33] andWang et al. [34]. In addition, to improve theintersolubility between diesel and alcohols, some studies considerbiodiesel as an additive to stabilize ethanol/diesel and n-butanol/diesel blends [35] and the combustion and emissions as fuellingbiodieselealcoholediesel blends have been investigated by Sukjitet al. [36] and Yoshimoto et al. [37]. Alcohol/biodiesel blend werealso investigated in diesel engines because the disadvantage ofworse cold flow behavior and higher viscosity for biodiesel and thelower cetane number for alcohols could be offset and properties ofblends are similar to conventional diesel fuel [38,39]. Accordingly,the kinetic model based on biodieselealcohol has also beendeveloped by Togbe et al. [40].

In the recent 10 years, some advanced combustion modes, suchas homogeneous charge compression ignition (HCCI), premixedcharge compression ignition (PCCI), low temperature combustion(LTC), etc., have obtained tremendous attention. These new com-bustion modes can provide both good fuel economy and very lowemissions ofNOx andPM, but the auto-ignition timing is veryhard tocontrol and the operating range is limited [41]. To control the igni-tion timing and extend the operating range, many methods havebeen proposed, including advanced injection strategies, intaketemperature control, higher boost pressure, variable valve actuationand varying fuel properties [42e44]. Among these, fuel property is aquite important control parameter to these advanced combustionmodes. Many fuels have been used to control the auto-ignitiontiming and extend the load range in advanced combustion modes,including diesel [42], gasoline [43], dimethyl ether [44], biodiesel[5], ethanol [45], n-butanol [27], 2,5-dimethylfuran [46], etc.

However, in the view of ignition control and load extension, ahigh cetane number fuel is suitable for low loads and cold start dueto its easier auto-ignition ability, while a high octane number fuel ispreferable for high loads due to its anti-knock ability [47].Furthermore, the optimum octane number which can achieve thehighest thermal efficiency needs to be varied as the engine loadchanges [47]. To achieve the above targets, an effective method isthe use of a dual-fuel injection system. In the dual-fuel system, twoselected fuels have the opposite auto-ignition characteristics. Forexample, a high cetane number fuel is used to improve auto-ignition characteristics, while a high octane number fuel is usedto suppress auto-ignition. Then, the needed fuel properties can beachieved by changing the ratio of two fuels according to thedifferent engine loads. Some previous studies have shown that theHCCI combustion process can be flexibly controlled and the oper-ating range can be extended by dual-fuel port-injection systems[48e50]. The dual-fuel injection system can also be composed ofdirect-injection and port-injection. One fuel with high cetanenumber is directly injected into the cylinder, while the other liquidfuel with high octane number and low boiling point or a high oc-tane number gas fuel is injected into the intake port, which isnamed as reactivity controlled compression ignition (RCCI) [51,52],premixed compression ignition (PCI) [53], or just called dual-fuelcombustion system [54,55]. In addition, some studies have

investigated the opposite combination in fuel properties in dual-fuel system, such as injecting high cetane number fuel at theintake port followed by the direct-injection of a high octanenumber fuel [56] or just using a single fuel ofgasolineegasoline þ DTBP (di-tert butyl peroxide cetane improver)[57]. In these dual-fuel injection systems, it includes not only thevariation of fuel properties but also the variation of charge strati-fication as the injection timing is changed. The charge stratificationhas been seen as an effective method to control the ignition timingand extend the operating range [58,59]. Therefore, the dual-fuelstrategy consisting of port- and direct-injection has more advan-tages than those of single-fuel or dual-fuel port-injection owing tothe cooperated control of fuel properties and charge stratification.In fact, the dual-fuel injection system has been widely studied inconventional mechanical pump diesel engine using the port-injection methanol [60] or gasoline [61e63] or two-stroke cycleengines [64] to reduce soot and NOx emissions and increase ther-mal efficiency. For the current common-rail injection system, it canoffer flexible injection strategies and thus form the needed chargestratification. Therefore, the current dual-fuel system should havemore advantages than the conventional mechanical pump dual-fuel system and the combustion and emission characteristicsneed to be a new awareness in the current engine technology.

In previous studies, Chen et al. [65] investigated the effect of n-butanol volume fractions (0e65%) and EGR (exhaust gas recircu-lation) rates (15% and 45%) on combustion and emissions on a dual-fuel systemwith port-injection of n-butanol and direct-injection ofdiesel fuel. They found that n-butanol fractions and EGR have acoupled impact on combustion process, and the dual-fuel systemcould simultaneously reduce both NOx and soot emission to a verylow level. Soloiu et al. [66,67] investigated the effects of port fuelinjection of n-butanol and direct-injection of biodiesel on com-bustion and emissions at idling and low loads (<5 bar IMEP (indi-cated mean effective pressures)). They found that soot and NOx

emissions reduced by ~90% and ~50%, respectively, and clean idlingtechnology could be developed based on this dual-fuel system. Inthe current study, the effects of different control parameters such asn-butanol premixed ratios, injection timings, and EGR rates oncombustion, emissions and performance were investigated on adual-fuel system with port-injection of n-butanol and direct-injection of diesel fuel. Meanwhile, a reduced chemical kineticmodel of n-butanol/biodiesel dual-fuel was coupled into thecomputational fluid dynamics (CFD) model to reveal the mecha-nism of combustion and emissions.

2. Experimental setup and methods

A six-cylinder diesel engine was modified to operate in onecylinder only. This arrangement gave a robust and inexpensivesingle-cylinder engine, but at the cost of the reliability of the brakespecific results. With a pressure transducer, the gross indicatedmean effective pressure during the compression and expansionstrokes only was calculated, which means that the effect of super-charging on the gas exchange process was absent. The detailedspecifications are shown in Table 1. Fig. 1 illustrates the experi-mental setup. The intake air was provided by the externalcompressor and air-conditioning system. The intake temperaturewas kept at 25 �C and intake absolute pressure was kept at0.18 MPa. The rates of exhaust gas recirculation (EGR) werecontrolled by adjusting the EGR valve and EGR rates were changedfrom 35% to 45%. Under steady operating conditions, the EGR ratewas calculated by the concentrations of carbon dioxide in intakeand exhaust gas. The in-cylinder pressure was measured with apressure transducer (Kistler 6125B). A charge amplifier (Kistler5018 A1003) was connected with the pressure transducer to

Table 2Fuel properties of gasoline, diesel, n-butanol, and biodiesel [16,20].

Gasoline Diesel n-Butanol Biodiesel

Molecular formula C4eC12 C12eC25 C4H9OH C12eC24Cetane number 0e10 40e55 25 47Octane number 80e99 20e30 96 e

Oxygen content (wt. %) e e 21.6 10Density (g/mL) 0.72e0.78 0.82e0.86 0.81 0.885Auto-ignition

temperature (�C)~300 ~210 385 363

Flash point (�C) atclosed cup

�45 to �38 65e88 35 166

Lower heating value(MJ/kg)

42.7 42.5 33.1 37.5

Boiling point (�C) 25e215 180e370 117.7 262e359Stoichiometric ratio 14.7 14.3 11.21 12.5Latent heating (kJ/kg)

at 25 �C380e500 270 582 200

Viscosity (Pa s) at 20 �C 0.5e0.6 2.8e5.0 3.64 4.11(at 40 �C)

Table 1Engine specifications.

Bore � stroke 105 � 125 mmDisplacement 1081.8 mLRated engine speed 2500 r/minConnecting road length 210 mmNumber of valves 4Compression ratio 16:1Swirl ratio 1.6Combustion chamber Bowl in pistonBowl volume 61.6 mLIntake valve open timinga 343 �CA ATDCIntake valve close timinga 133 �CA BTDCExhaust valve open timinga 125 �CA ATDCExhaust valve close timinga 343 �CA BTDC

a 0 �CA is taken to be top dead center compression.

H. Liu et al. / Energy 74 (2014) 741e752 743

convert charge to voltage. The cylinder pressure was recorded inhalf crank-angle increments, triggered by an optical shaft encoder(Kistler 2614A4). At each operating point, 100 pressure cycles wererecorded.

The dual-fuel injection was composed of port-injection of n-butanol and direct-injection of soybean biodiesel. The fuel prop-erties are listed in Table 2. An electronic port fuel injector (Delphi)was installed in the intake port. An electronic controller was usedto control both injection timing and injected fuel mass flow. Theport-injection timing was maintained at intake valve close to pro-vide the homogeneous charge. Soybean biodiesel was directlyinjected into the cylinder by a common-rail injection system(Bosch). The injection timing was varied to form the differentcharge stratifications in the cylinder. The flow-rate for port- anddirect-injection fuel was measured by fuel consumption meters(AVL 733S) with gravity scale. The specifications of the port andcommon-rail injection system are listed in Table 3.

Gaseous emissions were measured by a gas analyzer (HORIBAMEXA 7100DEGR), which measured total hydrocarbon by a methodof hydrogen flame ionization, CO and CO2 by non-dispersiveinfrared and NOx by a chemiluminescent analyzer. Soot wasmeasured by a filter paper smoke meter (AVL 415S). The indicatedspecific dry soot (unit: g/kW h) was calculated through thefollowing formula [68]:

Fig. 1. Engine setup. 1, Compressor; 2, bypass valve; 3, air flow meter; 4, air tank; 5,intake cooler; 6, EGR valve; 7, EGR cooler; 8, one-way valve; 9, port injector; 10, directinjector; 11, pressure transducer; 12, charge amplifier; 13, encoder; 14, backpressurevalve; 15, smoke meter; 16, exhaust analyzer.

soot ¼1.0:405� 5:32� FSN� e0:3062�FSN � 0:001

��mair þmfuel

�.ð1:2929� PiÞ

(1)

where FSN is the filter smoke number,mair andmfuel denote air andfuel mass flow (kg/h), respectively, and Pi denotes indicated power(kW).

In this study, all tests were conducted at the engine speed of1500 r/min. The cooling water and lubricating oil temperature werekept at 85 ± 2 �C and 95 ± 2 �C, respectively. Other uncertainties ofthe measurement instruments have been shown in Table 4. At eachtested point, the engine was run for several minutes until thecontrolled and measured parameters were stable. Then, the resultsof combustion pressures, emissions and performance were recor-ded for the off-line analysis.

3. Description of chemical kinetics and computational fluiddynamics (CFD) model

The multi-dimensional CFD (KIVA-3vr2 code) was used toinvestigate the combustion and emissions mechanisms. Thereduced chemical kinetic model of n-butanol/biodiesel dual-fuelwas coupled into the CFD model. In this reduced kinetic model,the methyl decanoate (MD) was used as the biodiesel surrogate.The previous study has showed that the MD was a good biodieselsurrogate due to the similar characteristics in combustion andemissions [69]. Then, the detailed MD kinetic model proposed byHerbinet et al. [15] was reduced by the method of direct reactiongraph, path flux analysis and sensitivity analysis. A skeletal mech-anism including 116 species and 517 reactions was achieved and thedetailed reduced process and validation can be found in Ref. [69].

Table 3Specifications for port and common-rail injection system.

Port-injection Number of holes 4Included spray angle 15�

Steady flow-rate 700 mL/minInjection pressure 0.3 MPa

Direct-injection Number of holes 8Included spray angle 150�

Hole diameter 0.15 mmSteady flow-rate at 100 bar 500 mL/30 sInjection pressure 100 MPaElectronic control unit EDC7 (BOSCH)

Table 4Uncertainties of the measurement instruments.

Instrument Uncertainties Resolution/sensitivity

Gaseous analyzer (HORIBA MEXA7100DEGR)

±0.5% full scale 1 ppm

Smoke meter (AVL 415S) 0.005 FSN þ 3% ofmeasured value

0.001 FSN

Inecylinder pressure (Kistler 6125B) <±1% �16 pC/barAir flow meter (vortex shedding flow

meter)<±1% 0.1 m3/h

Fuel flow meter (AVL 733S) <±1% 0.01 kg/hIntake gas pressure (pressure

transmitter)±1 kPa 0.1 kPa

Intake gas temperature (k-typethermocouple)

±1 �C 0.1 �C

Fig. 2. The geometry of combustion chamber and computational mesh at �30� CAATDC.

H. Liu et al. / Energy 74 (2014) 741e752744

The reduced kinetic mechanism for n-butanol and polycyclic aro-matic hydrocarbons (PAH) proposed byWang et al. [34] was used inthe current study. Finally, the reduced dual-fuel kinetic mechanismincluding 157 chemical species and 641 elementary reactions andthe detailed reduced process and validation can be found inRef. [70]. The turbulence model of RNG keε was used to simulatethe turbulent characteristics in the cylinder. Spray droplet breakupwas modeled by KelvineHelmholtz and RayleigheTaylor model.For the emission's models, NOx emissions were simulated by areduced NOx mechanism that was derived from the Gas ResearchInstitute NO mechanism [71], while soot emissions were simulatedby a multi-step phenomenological soot model and the polycyclicaromatic hydrocarbons were used as the soot precursor [72].

The geometry of combustion chamber and computational gridsused for the simulations are shown in Fig. 2. The fuel injector iscenter-located and has eight holes. In order to improve thecomputational efficiency, a sector of 45� was calculated from intakevalve close to exhaust valve open. The cell numbers of the engine atbottom dead center were about 9500. Fig. 3 shows the comparisonof cylinder pressure and heat release rate between experiment andsimulation at early-injection (�30� CA ATDC) and late-injection(�9� CA ATDC). It can be seen that the CFD simulations can repro-duce the experimental combustion pressures and heat release rates.

4. Results and discussion

4.1. Effects of premixed ratios and injection timings on combustioncharacteristics

In this paper, the premixed ratio (rp) was defined as the ratio ofcycle energy of premixed fuel to total energy which included pre-mixed fuel and directly injected fuel. The rp can be calculated usingthe following formula:

rp ¼ Qp

Qt� 100% ¼ _mpLHVp

_mpLHVp þ _mdLHVd � 100%(2)

where _m is the fuel mass flow-rate and the unit is mg/cycle, andLHV represents the lower heating value of the fuel and the unit is J/mg. The subscripts p and d denote premixed and directly injectedfuel, respectively. The previous dual-fuel study has showed thatlower NOx and soot emissions were obtained at higher rp condition.Meanwhile, a higher rp could also reduce the pressure rise rate,which is beneficial to extend the engine load [51]. Therefore, threehigh premixed ratios, 80%, 85% and 90% were controlled, while theEGR rate was kept at 35% in this part. It should be noted that thecorresponding premixed fuel mass ratios were 84%, 88% and 92%due to the lower heating value of n-butanol, while the calculated rp

based on Eq. (2) were used in this work. The overall energy injectedper cycle was constant for different injection strategies and themass flow of total fuel was kept at 60 mg of equivalent biodiesel.The equivalent biodiesel means that the mass flow of n-butanol inthe total fuel is converted to biodiesel mass flow according to thelower heating value. The total mass flow of equivalent biodiesel canbe calculated using the following formula:

Total mass flow ¼ _mpLHVp

LHVdþ _md (3)

The symbolic meaning and unit in Eq. (3) is the same as Eq. (2).Accordingly, the direct-injection of biodiesel mass was 12 mg, 9 mgand 6 mg. Due to the same energy input in each cycle, the indicatedmean effective pressures (IMEP) were roughly kept at 0.95 MPa(56% load of the original engine) for tested cases.

Fig. 4 shows the effects of premixed ratios and injection timingson the combustion phase, maximum pressure rise rate (MPRR) andcoefficient of variation (COV) in IMEP. The combustion phasing isrepresented as the 50% burn point (CA50). The location of CA50greatly influences the performance, emissions, mechanical load ofdiesel engines. In this study, the CA50was controlled between 2 �CAand 10 �CA after top dead center (ATDC) by varying injection tim-ings. It can be seen that the CA50 is initially advanced, and then isdelayed as the injection timing retards. A same CA50 can be ach-ieved by the early or late-injection of biodiesel, so the injectionstrategy can be divided into early-injection and late-injection ac-cording to the different injection timings. The discontinuous curvescan be seen as the rp decreasing to 80e85%, which is due to thelimitation of MPRR. For example, as the injection timing is delayedfrom �43 �CA to �35 �CA under the rp of 80%, the MPRR increasesfrom0.4 to 1.0MPa/�CA. For this experimental engine, the limitationon themaximumallowablepressure rise rate is 1.0MPa/�CA, beyondwhich combustion tends to become “knocky”. Therefore, the higher-pressure rise rate constrains the range of injection timings. How-ever, as the injection timing is later than�10� CA ATDC, theMPRR islower than the limitation of 1.0 MPa/�CA again and the MPRR re-duces with the delay of injection timing. The reduced pressure riserate should be attributed to the retard of CA50. With greater CA50retard, the rate of expansion due to piston motion increases, whichresults in the lower combustion temperature and the slowed pres-sure rise rate. However, even though retarding CA50 is an effectivemethod to reduce pressure rise rates andextendoperating loads, theretard of CA50 is limited by poor cycle-to-cycle stability. In thisstudy, as the CA50 is delayed over 10 �CA ATDC, the COV increasesand results in potentially unstable combustion. This result is

Fig. 3. Comparison between the measured and computed cylinder pressure and heat release rate under early-injection (�30� CA ATDC) and late-injection (�9� CA ATDC). Thepremixed ratio of n-butanol was 85% and the EGR rate was 35%.

H. Liu et al. / Energy 74 (2014) 741e752 745

consistent to the previous study which reported that the combus-tion phasing could not be retarded beyond 10e15 �CA ATDC [41].Therefore, tomaintain theCA50between2 �CAand10 �CAATDC, theinjection timings of biodiesel are limited by the COV and MPRR forthis dual-fuel combustion system.

Fig. 5 shows the effects of injection timings on the in-cylinderpressure, apparent heat release rate (AHRR), pressure rise rate(PRR) and mean gas temperature. As to the early-injection, Fig. 5ashows that the combustion process presents a single-stage hightemperature heat release (HTHR). For early-injection cases, thekinetic simulation in Fig. 6a shows that biodiesel was not consumeddirectly after injection. With the piston moving up, the in-cylindertemperature and pressure increases and biodiesel fuel starts to beconsumed and the H2O2 starts to form at about �17 �CA ATDC.Subsequently, the mole fraction of biodiesel continually reducesand is consumed completely at �2 �CA ATDC. Meanwhile, theobvious heat release can be seen in Fig. 5a and a large amount of OH

Fig. 4. The combustion phase, maximum pressure rise rate (MPRR), and coefficient ofvariation (COV) under different premixed ratios and injection timings, the EGR ratewas kept at 35%. (a) Early-injection. (b) Late-injection.

is formed. After, n-butanol starts to be consumed at about �2 �CAATDC and the whole consumption of n-butanol is very fastbetween �2 �CA and 5 �CA ATDC, which results in the intense heatrelease process as shown in Fig. 5a. In addition, it can also be seenthat with the retard of injection timing, the ignition timing ad-vances, the peak of in-cylinder pressure, PRR and mean gas tem-perature increase. As injection timing retards, more stratifiedmixture can be formed due to the shorter mixing time. The previ-ous study has reported that the stratification increased the localequivalence ratio and formed more fuel-rich regions, which madethe auto-ignitionmuch easier under the relative homogeneous leanconditions [73]. Therefore, the auto-ignition timing was advancedwith the increase of mixture stratification.

As to the late-injection, Fig. 5b shows that the combustionprocess presents a two-stage high temperature heat release (HTHR)and the ignition timing is retarded with the delay of injectiontiming. The peaks of in-cylinder pressure, PRR and mean gas tem-perature decrease with the retard of injection timing. The kineticsimulation in Fig. 6b shows that biodiesel was consumed directlyafter injection due to the high pressure and temperature near thetop dead center under late-injection conditions. Subsequently, themole fraction of biodiesel reduces much quickly than that of early-injection case and is consumed completely at 0 �CA ATDC. Mean-while, the peak of first-stage heat release as shown in Fig. 5b waslocated at 0 �CA ATDC and n-butanol just starts to be consumed attop dead center. Therefore, the first-state stage HTHRmainly comesfrom the combustion of biodiesel. After, the first-stage of HTHRtriggers the remainder mixture to burn and results in a highersecond stage of HTHR. Further, it can be noted that the first peak ofAHRR is nearly consistent with different injection timings, whilethe second peak of AHRR decreases with the retard of injectiontiming. As the biodiesel mass flow decreases, the first-state HTHRreduces as shown in Fig. 7, which again demonstrates that the first-stage HTHR should mainly come from the combustion of biodiesel.In addition, compared to the early-injection, the pressure rise ratesof late-injection demonstrate two peaks. Furthermore, the secondpeak of PRR is higher than that of first one, which indicates that theMPRR of late-injection is caused by the fast heat release of pre-mixed n-butanol fuel during the second stage HTHR. Based on theabove discussion, it can be found that the biodiesel is consumedfirstly for both early and late-injection. Then, n-butanol starts totake part in the reaction as biodiesel is nearly fully consumed.Furthermore, as the OH radical starts to be formed, the consumedn-butanol is quite limited. Therefore the auto-ignition at the cur-rent condition was trigged by the reaction of biodiesel.

Fig. 7 shows the effects of premixed ratios on the in-cylinderpressure, AHRR, PRR and mean gas temperature. The combustion

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Fig. 5. The in-cylinder pressure, heat release rate, pressure rise rate, and mean gas temperature under different injection timings, the EGR rate was kept at 35%. (a) Early-injection(�30� CA ATDC). (b) Late-injection (�9� CA ATDC).

H. Liu et al. / Energy 74 (2014) 741e752746

phasing is retarded with the increase of rp due to the lower globalfuel reactivity. Then, the retarded combustion phasing reduces thelevel of constant-volume combustion, which results in the lowercombustion temperature, pressure and AHRR under higher pre-mixed ratios conditions. The relatively slow heat release prolongsthe combustion duration and decreases MPRR for higher rp case. Tofurther reveal the effect of premixed ratios on the combustionprocesses, Fig. 8 shows the in-cylinder pressure and AHRR underearly-injection and late-injection conditions with the same CA50. Itshows that the combustion pressure and AHRR at the rp of 85% arenearly consistent to those of 80% case. However, as the rp increasesto 90%, the combustion characteristics are different to those oflower premixed ratios. At the rp of 90%, the heat release rate isslowed, the peak value of AHRR is reduced and the combustionduration is prolonged. The higher premixed ratio results in more n-butanol in the cylinder, which reduces the fuel reactivity and re-action rate, leading to the longer combustion duration.

4.2. Effects of premixed ratios and injection timings on combustionefficiency and indicated thermal efficiency (ITE)

Fig. 9 shows the effects of premixed ratios and injection timingson the combustion efficiency and ITE (indicated thermal efficiency).

-30 -20 -10 0 10 20 30

10-4

10-3

10-2

Crank Angle (oCA ATDC)

MDNBOHH2O2

Mol

e fr

actio

n

(a) Early-injection (-30°CA ATDC)

Fig. 6. The mole fractions of fuels, OH radical and H2O2 at differe

The combustion efficiency is evaluated from the exhaust gascomposition. The gross indicated thermal efficiency is evaluated bymeasuring the fuel flow and the indicated mean effective pressureduring the compression and expansion strokes only. This meansthat the effect of supercharging caused by the external aircompressor on the gas exchange process is absent.

It can be seen that the combustion efficiency and ITE increasefirstly and then decrease with the retard of injection timing. Theearlier or later injection timings result in the delay of CA50, thelower combustion temperature and the lower degree of constant-volume combustion, and thus the combustion efficiency and ITEare reduced. Compared to the lower rp, the higher rp leads to moren-butanol entering into the crevices where the fuel cannot be fullyoxidized. Therefore, the combustion efficiency decreases with theincrease of rp. The ITE is comparable for both 80% and 85% of rp,while the ITE reduces by about 0.6% as increasing rp to 90%. Inaddition, Fig. 9 also shows that the early-injection has a little higherITE compared to the late-injection.

4.3. Effects of premixed ratios and injection timings on emissions

Fig.10 shows the effects of premixed ratios and injection timingson the NOx, soot, HC and CO emissions. The NOx emissions increase

-30 -20 -10 0 10 20 30

10-4

10-3

10-2

Mol

e fr

actio

n

Crank Angle (oCA ATDC)

MDNBOHH2O2

(b) Late-injection (-9°CA ATDC)

nt injection timings (MD, methyl decanoate; NB, n-butanol).

2

4

6

8

10

12 Injection timing/rp -7o / 80% -7o / 85% -7o / 90%

Hea

t rel

ease

rate

J/

o CA

Mea

n ga

s te

mpe

ratu

re (K

)

e tare si r

e rus se rP(M

Pa/ o C

A)

e russer predn ilyc -n I(M

Pa)

080160240320400

-0.4

0.0

0.4

0.8

Crank Angle (oCA ATDC)-15 -10 -5 0 5 10 15 20 25 30

800

1000

1200

1400

1600

Fig. 7. The in-cylinder pressure, heat release rate, pressure rise rate, and mean gastemperature under different premixed ratios, the EGR rate was kept at 35%.

91

92

93

94

95

96

97

Indi

cate

d th

erm

al e

ffici

ency

(%)

Injection timing of biodiesel ( CA ATDC)

)%(

ycneiciff enoit sub

moC

rp=80%rp=85%rp=90%

-45 -40 -35 -30 -25 -20 -15 -10 -5 045.0

45.5

46.0

46.5

47.0

47.5

48.0

Fig. 9. The combustion efficiency and indicated thermal efficiency under differentpremixed ratios and injection timings, the EGR rate was kept at 35%.

H. Liu et al. / Energy 74 (2014) 741e752 747

with the retard of injection timing under early-injection conditions.Retarding injection results in the increase of charge stratificationand the combustion temperature, thus the NOx emissions increase.However, the NOx emissions decrease with the delay of injectiontiming under late-injection conditions, which is attributed to the

2

4

6

8

10

12

Hea

t rel

ease

rate

J/

o CA

erus serPredn ilyC-nI

( MPa

)

Injection timing / rp -43o / 80% -36o / 85% -23o / 90%

Hea

t rel

ease

rate

J/

o CA

erusserPrednilyC-nI

(MPa

)

Injection timing / rp -5o / 80% -6o / 85% -8o / 90%

CA50=8.5oCA

080160240320400

-15 -10 -5 0 5 10 15 20 25 30

2

4

6

8

10

12

Crank Angle (oCA ATDC)

CA50=9.0oCA

080160240320400

Fig. 8. The in-cylinder pressure and heat release rate under early-injection and late-injection conditions with the same CA50, the EGR rate was kept at 35%.

retard of combustion phasing and the lower combustion temper-ature. Compared to the late-injection, the early-injection can ach-ieve lower NOx emissions, which is due to the fact that the morehomogeneousmixture formed in early-injection results in less localhigh temperature zones as shown in Fig. 11. Furthermore, thesimulation results in Fig. 11 also show that the regions of high NOx

emissions are agreement with the high temperature zones.Compared to early-injection, the late-injection has wider hightemperature zones in the cylinder and thus the higher NOx

emissions.

-45 -40 -35 -30 -25 -20 -15 -10 -5 03

6

9

12

15

Injection timing of biodiesel ( CA ATDC)

2345670.000

0.0020.0040.0060.0080.010

Soot

(g/k

W.h

)

0.00.40.81.21.62.0

rp=80%rp=85%rp=90%

CO

(g/k

Wh)

HC

(g/k

W.h

)N

Ox

(g/k

Wh)

Fig. 10. The NOx, soot, HC and CO emissions under different premixed ratios and in-jection timings, the EGR rate was kept at 35%. The equivalence ratio was approximately0.478 at this case.

Fig. 11. The simulation results of NOx and temperature distribution in the cylinder at different injection timings of biodiesel.

H. Liu et al. / Energy 74 (2014) 741e752748

The quite low soot emissions (<0.01 g/kWh) are achieved in thisstudy. This result can be attributed to the following reasons. On onehand, the dual-fuel combustion system offers more homogeneousmixture, which leads to the decrease of the local high temperatureand high equivalence ratio zones. On the other hand, the oxygen inn-butanol and biodiesel can suppress the soot formation. The sootemissions of late-injection are comparable with those of early-injection cases, although the late-injection has stronger chargestratification. The simulation results of soot and equivalence ratiodistribution and soot formation and oxidation processes are shownin Fig. 12. It can be seen that the early-injection has lower sootformation due to the more homogeneous distribution of equiva-lence ratios. For the late-injection, it results in higher soot forma-tion due to the more local fuel-rich regions. However, the highercombustion temperature increases the soot oxidation under late-injection case, which makes the soot emissions under late-injection conditions quite low either.

Fig. 12. The simulation results of soot and equivalence ratio distribution in the cylinder an

The HC and CO emissions decrease firstly and then increasewiththe retard of injection timing. The earlier or later injection timingsresult in the retard of CA50 and the lower degree of constant-volume combustion and thus the lower combustion temperature,which is disadvantageous to the oxidation of HC and CO. Comparedto the lower rp, the higher rp leads to more n-butanol entering intothe crevices where the fuel cannot be fully oxidized and thus hasthe higher HC and CO emissions. Fig. 13 shows the simulation re-sults of CO distribution in the cylinder. Even though the time of COformation is different for early and late-injection, the main COemissions come from the near wall region due to the lower tem-perature in these regions. Furthermore, compared to early-injection, the late-injection has higher CO emissions. The distri-bution of biodiesel fuel concentrates on the center of combustionchamber for late-injection case, which results in that the radicalsprovided by biodiesel combustion cannot react with n-butanolcompletely. Therefore, more n-butanol near wall regions cannot be

d the soot formation and oxidation process at different injection timings of biodiesel.

Fig. 13. The simulation results of CO distribution in the cylinder at different injectiontimings of biodiesel.

0

6

12

18

24

30

36

42

Com

bust

ion

dura

tion

(o CA

)

Injection timing of biodiesel ( CA ATDC)

yalednoitingI

(o CA

)

EGR=35% EGR=40% EGR=45%

-45 -40 -35 -30 -25 -20 -15 -10 -5 010

12

14

16

18

20

22

24

Fig. 14. The ignition delay and combustion duration under different EGR rates andinjection timings, the premixed ratio was kept at 85%.

H. Liu et al. / Energy 74 (2014) 741e752 749

oxidized completely and leads to higher CO emissions at late-injection case.

As a whole, the early-injection has more advantages than thoseof late-injection due to its lower NOx emissions and the comparableor lower HC, CO and soot emissions. For the effects of premixedratios under early-injection conditions, the NOx, HC and CO emis-sions increasewith the increase of rp at a given CA50, while the sootemissions are comparable with different rp cases. Therefore,considering the compromise between engine performance andemissions, the early-injection has more advantages than the late-injection, while the rp of 85% has better results compared toother premixed ratios.

91

92

93

94

95

96

97

Indi

cate

d th

erm

al e

ffici

ency

(%) EGR=35%

EGR=40% EGR=45%

Injection timing of biodiesel ( CA ATDC)

)%( ycneiciffe noitsub

moC

-45 -40 -35 -30 -25 -20 -15 -10 -5 045.0

45.5

46.0

46.5

47.0

47.5

48.0

Fig. 15. The combustion efficiency and indicated thermal efficiency under differentEGR rates and injection timings, the premixed ratio was kept at 85%.

4.4. Effects of EGR on combustion, emissions and load extension

In this section, the premixed ratio was kept at 85% and the totalequivalent biodiesel mass was also kept at 60mg per cycle. The EGRrates were changed from 35% to 45%. The CA50 was also controlledbetween 2 �CA and 10 �CA ATDC. Fig. 14 shows the ignition delayand combustion duration at different EGR rates and injection tim-ings. At early-injection conditions, different EGR rates have littleeffect on ignition delay, while EGR rates have larger effects onignition delay at late-injection conditions and higher EGR ratesresult in the longer ignition delay. With the increase of EGR rates,the combustion reaction rate reduces and the combustion durationincreases. Fig. 15 shows the combustion efficiency and gross indi-cated thermal efficiency (ITE). With the increase of EGR rates, thecombustion efficiency reduces. Increased EGR rates results in thedecrease of combustion temperature and thus partial fuel cannot beoxidized completely. The ITE is comparable for both 35% and 40%EGR, while the ITE reduces by 1e2% as increasing EGR to 45%. Onone hand, the decreased combustion efficiency results in thedecrease of ITE. On the other hand, the combustion duration isprolonged under higher EGR rates as shown in Fig. 14, which re-duces the degree of constant-volume combustion.

Fig. 16 shows the NOx, soot, HC, and CO emissions at differentEGR rates and injection timings. It can be seen that the NOx

emissions of the late-injection decrease with the increase of EGRrates. At 45% EGR, the NOx emissions of late-injection are compa-rable with the early-injection. The soot emissions decreasewith theincrease of EGR rates. The decrease of soot emissions should be dueto the longer ignition delay under higher EGR rates as shown inFig. 14, which can improve the mixing process of biodiesel andreduce the over-rich zones. However, HC and CO emissions increasewith the increase of EGR rates. The increasing EGR rates results inthe decrease of combustion temperature, which is disadvantageousto the oxidation of HC and CO. The higher HC and CO emissionsresult in the lower combustion efficiency under higher EGR rates asshown in Fig. 15.

To investigate the potential of high load extension, the injectedfuel mass flow was increased in this part until the combustionachieving to knocking limit. The injection timing was controlled atlate-injection due to it has larger anti-knocking based on previousstudies. Fig.17 shows themaximum gross IMEP, CA50, soot and NOx

-45 -40 -35 -30 -25 -20 -15 -10 -5 04

8

12

16

20

Injection timing of biodiesel ( CA ATDC)

2

5

8

11

140.0000.0020.0040.0060.0080.010

Soot

(g/k

W.h

)

0.00.40.81.21.62.0

EGR=35% EGR=40% EGR=45%

CO (g

/kW

h)

HC ( g

/kW

.h)

NOx

(g/k

Wh)

Fig. 16. The NOx, soot, HC and CO emissions under different EGR rates and injectiontimings; the premixed ratio was kept at 85%. The equivalence ratio was approximately0.478, 0.511, 0.551 at 35%, 40% and 45% EGR rates, respectively.

H. Liu et al. / Energy 74 (2014) 741e752750

emissions at different fuel mass flow. Themaximum IMEP is chosenfrom the results of different CA50 and the matching CA50 is alsoshown in Fig. 17. To keep the MPRR below the allowable limitationof 1.0 MPa/�CA, the CA50 achieving to the maximum IMEP retardswith the increase of the equivalent fuel mass flow. The largestachievable load is 13.48 bar at 40% EGR, which is equal to 80% load

55 60 65 70 75 80 85 90 95 1000.00.51.01.52.02.5

Injection mass of total fuel (mg)

0.00

0.01

0.02

0.030246810

CA

50 (o C

A)

91011121314

EGR=35% EGR=40%

NO

x (g

/kW

h)

Soot

(g/k

W.h

)IM

EP (b

ar)

Knocking point

Fig. 17. Large load extension with higher fuel mass flow, the premixed ratio was fixedat 85%.

of the original engine. However, the NOx and soot emissions arerelatively higher at this knocking limit point compared to those oflower fuel mass flow. If consider the emissions meanwhile, it can befound that the mass flow at 90 mg has better results which theIMEP can reach to 12.88 bar (76% load) and the soot and NOx

emission are 0.009 g/kW h and 0.88 g/kW h, respectively. At 35%EGR, the largest achievable load is 12.19 bar (72% load) and the sootand NOx emission is 0.016 g/kW h and 0.77 g/kW h, respectively.Compared to the 35% EGR, the case at 40% EGR can endure a higherfuel mass flow and keep the low soot emissions. This is due to thehigher EGR rates can prolong the ignition delay and improve themixing process. The NOx emissions are also affected by the com-bustion phasing. It can be seen that the case at 35% EGR has lowerNOx emissions than those of 40% EGR due to the retarding of CA50at the same fuel mass flow.

Therefore, it can be concluded that the higher EGR rate hasmoreadvantages on extending operating load and reducing NOx and sootemissions, while the combustion efficiency and indicated thermalefficiency decrease using higher EGR rate. Therefore, to obtain thehigh efficiency and clean combustion process, the cooperatedcontrol is necessary among direct-injection timing, premixed ratioand EGR rate.

5. Conclusions

The detailed combustion characteristics and emissions of n-butanol/biodiesel dual-fuel injection systemwere investigated on adiesel engine based on experiments and simulations. n-Butanolwas injected into the intake port to form premixed charge, whilesoybean biodiesel was directly injected into the cylinder. Thedifferent premixed ratios (rp) and EGR rates were investigated. Theinjection timings were adjusted to keep the 50% burn point (CA50)between 2� CA and 10� CA after top dead center for achieving stableoperation. Several conclusions can be drawn from this study.

1. A same CA50 can be achieved by the early or late-injection ofbiodiesel. For both early and late-injection, biodiesel isconsumed firstly and triggers the auto-ignition, then n-butanolstarts to take part in the reaction as biodiesel is nearly fullyconsumed. As to the early-injection, the combustion processpresents a single-stage high temperature heat release (HTHR).With the retard of injection timing, the ignition timing ad-vances, the peak of in-cylinder pressure and pressure rise rate(PRR) increase. As to the late-injection, the combustion processpresents a two-stage HTHR and the first-state HTHR mainlycomes from the combustion of biodiesel fuel. With the retard ofinjection timing, the ignition timing delays, the peak of in-cylinder pressure and PRR decrease. Increasing the premixedratios can retard the combustion phasing and reduce the cyl-inder pressure and PRR.

2. The combustion efficiency and indicated thermal efficiency (ITE)increase first and then decrease with the retard of injectiontiming. The ITE is comparable for both 80% and 85% of rp, whilethe ITE reduces by about 0.6% as increasing rp to 90%. The early-injection has a little higher ITE compared to the late-injection.

3. Very low NOx and soot emissions can be achieved simulta-neously by the dual-fuel combustion system. The early-injectionhas lower NOx emissions compared to the late-injection due tothe lower combustion temperature. The soot emissions arecomparable for both early- and late-injection. For early-injection, more homogeneous charge results in lower soot for-mation. For late-injection, more over-rich regions results inhigher soot formation, but the higher combustion temperaturepromotes the soot oxidation. The higher premixed ratio leads tothe higher HC and CO emissions.

H. Liu et al. / Energy 74 (2014) 741e752 751

4. With the increase of EGR rates, the NOx and soot emissionsdecrease, while the HC and CO emissions increase. The ITE iscomparable for both 35% and 40% EGR, while the ITE reduces by1e2% as increasing EGR to 45%. By controlling the combustionphase and EGR rate, the engine load of 12.88 bar IMEP (76% loadof the original engine) can be achieved and keep quite low NOx

and soot emissions.

Acknowledgement

The authors would like to acknowledge the financial supportsprovided by National Natural Science Found of China (NSFC)through its project of 51320105008 and 51206120.

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