An experimental study on the influence of EGR rate and fuel octanenumber on the combustion characteristics of a CAI two-stroke cycleengine
Amin Mahmoudzadeh Andwari*, Azhar Abdul Aziz, Mohd Farid Muhamad Said,Zulkarnain Abdul LatiffAutomotive Development Centre (ADC), Faculty of Mechanical Engineering, Universiti Teknologi Malaysia (UTM), 81310 Johor Bahru, Malaysia
h i g h l i g h t s
� The first study utilizing external and internal EGR on a CAI two-stroke engine.� Influence of In-EGR, Ex-EGR and octane number on combustion phasing of the engine.� Influence of In-EGR, Ex-EGR and octane number on cyclic variability of the engine.� Identification of the CAI combustion operating regions in the engine.
a r t i c l e i n f o
Article history:Received 21 March 2014Accepted 28 June 2014Available online 5 July 2014
Keywords:Controlled auto-ignitionTwo-stroke cycle engineInternal EGRExternal EGROctane Number
Having the higher power-to-weight ratio feature of two-stroke cycle engines and higher thermodynamicefficiency of controlled auto-ignition (CAI) combustion proposes a promising concept for future internalcombustion engines. The control of start of auto-ignition (CA10) and its cyclic variation are major issuesthat should be addressed. The research to date has tended to focus on the influence of internal EGR (In-EGR) utilization on the combustion characteristics of CAI two-stroke cycle engine rather than externalEGR (Ex-EGR) utilization. This experimental study aims to investigate the influence of Ex-EGR incor-porated with In-EGR utilization on the combustion characteristics and the cyclic variability of a single-cylinder CAI two-stroke cycle engine using different fuel octane number (ON). Engine's combustion-related parameters were examined statistically associated with Coefficient of Variation (COV) andStandard Deviation (STD). It is perceived that the significant CAI combustion parameters, includingHRRmax, qHRRmax, LTHR, qLTHR, CA10, CA50, MFB and TCA10 are strongly correlated to variation of In/Ex-EGRand fuel ON and they suggest an effective means to control CAI combustion phasing and cyclic variability.Furthermore, CA10 and its cyclic variability exhibited that are more sensitive to altering Ex-EGRcompared to altering In-EGR.
Global increases in fuel price combined with finite resources offossil fuel and stringent exhaust emission legislation have becomethe principle motivation of Internal Combustion Engines (ICEs)scientists to produce as efficient and clean engine as possible [1,2].As a result of unmistakable advantages of two-stroke cycle enginesincluding lightweight, fewer components, and the potential to pack
almost twice the power-density than that of a four-stroke enginehaving similar capacity, they have always been always taken intoconsideration for vehicle propulsion purpose. However their con-ventional drawbacks associated with poor combustion efficiencyand excessive uHC and CO emissions, specifically at low load, raiseissues concerning their practicality. These drawbacks can beaddressed by introducing an alternative combustion calledControlled Auto-Ignition (CAI), also known as HomogeneousCharge Compression Ignition (HCCI), suggesting high thermody-namic and volumetric efficiency; low emission and lean burncombustion operation by lower cyclic variation [3e5]. Premixedmixture of air and fuel (homogeneous) will be compressed until thebeginning of auto-ignition [6,7]. Owing to high pressure and
Engine type Single cylinder two-stroke case reed valve
Bore � Stroke 59 � 54.5 (mm)Displacement 149 (cm3)Scavenging type Schnuerle (Loop Scavenging)Scavenging port timing 117.5 CAD A/BTDCExhaust port timing 82.5 CAD A/BTDCExhaust system Expansion chamberGeometric compression ratio 11.3Cooling system Water cooledFuel supply system Electronically controlled port fuel injectionScavenging coefficients K0 ¼ �0.01456, K1 ¼ �0.84285, K2 ¼ �0.28438
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temperature of mixture combustion will start at multiple auto-ignited points in the entire combustion chamber simultaneouslyand spontaneously. Consequently, there is no longer a flame front(e.g., SI engine) and diffusion burning (e.g., CI engine) [8e10]. Inorder to achieve auto-ignition combustion, in-cylinder averagecharge temperature at exhaust port closure (Tepc or temperature atstart of effective compression) must be high [11]. In a conventionaltwo-stroke cycle engine, unlike partial load, at low and high loadsTepc is not high enough to steadily initiate auto-ignition combustion[12,13]. Study on CAI combustion began with a two-stroke cycleengine in order to improve unburned hydrocarbons (uHC) and toexamine the combustion chemistry of auto-ignition [11,14]. It wasconcluded that in order to induce auto-ignition combustion, Tepcmust be high enough to increase the concentration of the activeradical species well before self-ignition, which is crucial in theinitiation of the auto-ignition [15,16]. Throttling of the exhaust port(or exhaust pipe) was used several times in order to increase theamount of trapped residual gas [17e19]. Computational and nu-merical investigations have confirmed that an exhaust and transferport control valve can regulate themixing between residual gas andfresh intake air charge (i.e. charge stratification) which leads toimprovement of fuel economy, emission and combustion efficiency[20,21]. Correspondingly, the gradient of temperature will beincreased which accounts for commencement of the auto-ignition[22,23]. Numerous experimental studies were conducted in orderto investigate the feasibility of using auto-ignition in two-strokeengine using one special exhaust valve controlling both theexhaust port area and the exhaust port timing [24]. It wasconcluded that not only the amount of active residual gases in thecombustion chamber but also the in-cylinder pressure at the startof compression stroke (Pepc) can be controlled properly by exhaustthrottle control valve [25]. Investigations concerning the effect ofdifferent fuel type (gasoline, methanol and ethanol ether-basedfuel) on CAI two-stroke combustion demonstrated that the CAIcombustion operating range is strongly correlated with the type offuel used [26,27]. Both experimental and simulation investigationshave indicated that net heat release rate and timing of both cooland hot flames of CAI combustion are influenced by fuel octanenumber [28e30]. The effect of In-EGR utilization on the CAI com-bustion characteristics was carried out using exhaust port throt-tling and negative valve overlap (rebreathing) strategies when theengine was run in WOT condition [31]. The outcomes indicatedsubstantial reduction in exhaust emissions, extended CAI operatingregion and improved combustion efficiency [32e34].
There are three key issues related to the CAI combustion phasingcontrol which should be addressed including start and control ofthe auto-ignition and its cyclic variability [35]. High intake chargetemperatures and a considerable amount of charge dilution shouldto be provided to increase the Tepc to initiate and sustain chemicalreactions leading to auto-ignition processes [36]. Several studieshave been undertaken to investigate the effect of EGR in CAIcombustion process numerically and experimentally[31,33,34,37e40]. Generally, the overall effect of EGR applicationcan be explained as follows [7,40e45]:
i) Charge Heating Effect; Hot burned gases increase the tem-perature of the intake charge owing to their heating effect.
ii) Heat Capacity or Thermal Effect; Due to existence of somespecies in burned gases (e.g. CO2 and H2O), the overall heatcapacity of the in-cylinder charge will be higher.
iii) Dilution Effect; Substantial reduction in the air/oxygenconcentration, which caused by replacement of some inertgases. Thus, the concentration of active species will bedecreased leading to lower chemical activity and reactionrate.
iv) Chemical Effect; The chemical reactionwill be increased dueto the influence of some activated radical species.
It is perceived that ignition timing is principally influenced bycharge heating and thermal effects, and is less sensitive to thechemical and dilution effects [45]. The combined effect of thesefour factors is assumed to regulate the ignition timing of com-bustion. Accordingly, the ignition timing is advanced if the firsteffect is substantial, but will retard if the other three effects aremore dominant. The purpose of this study is to investigate theinfluence of In-EGR, Ex-EGR and fuel ON in relation to the controlof CAI combustion phasing and its cyclic variability in the case oftwo-stroke cycle engine. Even though the effect of In-EGR on CAItwo-stroke engine has been studied previously, far to little atten-tion has been devoted to study of the influence of Ex-EGR on thecyclic variability and the combustion characteristics of CAI two-stroke engines. To the best of the authors' knowledge, this exper-imental investigation is the first study undertaken to evaluate theinfluence of utilization of external EGR incorporated with internalEGR and fuel octane number changes in accordance with thecombustion phasing and cyclic variability of the CAI two-strokecycle engine.
2. Experimental engine setup, instrumentation and operatingconditions
A single-cylinder, two-stroke, naturally aspirated, water cooledengine was taken into consideration and modified in order to meetthe CAI experimental engine test rig requirements. Table 1 specifiesthe experimental engine details.
An electronic controlled port fuel injection system is used, inwhich Injector's Pulse Width (PW) regulates flow rate of theinjected fuel (control of AFR) into the engine's intake port. Thefuel injection system is also equipped with an exhaust lambdasensor (i.e. closed loop control) to monitor the engine's real timeAFR. In order to estimate the engine air consumption a pitot tubeis installed in conjunction with a micro manometer (TSI AirflowInstruments 5825). The intake charge flow temperature can beregulated via an electric heater device. Combustion burned gasescan be retained in the combustion chamber by means of exhaustport area restriction. These high temperature gases will be mixedwith the incoming fresh fuel-air mixture resulting in highertemperature and pressure after completion of the scavengingprocess. This strategy of the burned gas utilization is known asInternal Exhaust Gas Recirculation (In-EGR). Consequently oneball type valve (38 mm diameter) is installed in the exhaust pipei.e. 50 mm away from the engine's exhausting port downstreamside. In the meantime, right after the In-EGR valve, one T-jointconnection (25 mm diameter) is placed to direct a portion of theburned gases from the exhaust pipe back into the intake runner.
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This method of burned gas utilization referred to as ExternalExhaust Gas Recirculation (Ex-EGR) will result in higher intakecharge temperature and different intake charge composition. Agate type valve (25 mm diameter) through the Ex-EGR piping lineadjusts the flow rate of the Ex-EGR. All of the Ex-EGR and intakerunner piping are completely insulated to minimize for convec-tion and conduction heat transfer losses. As is illustrated in Fig. 1,K-type thermocouples (±1 �C accuracy) are located at distinctpositions to measure Tex, Tin and Tsc representing engine exhaustgas temperature, intake gas temperature and transfer port gastemperature respectively. Thus Tex is measured from 30 mmdownstream of the exhaust port and the Tsc is estimated from10 mm right before the upstream of scavenging port. Tin ismeasured after the Ex-EGR in the mixing box. A spark plug typepiezoelectric pressure transducer (KISTLER 6117B) is used to re-cord the engine's pressure history. The engine crankshaft iscoupled to a crank angle encoder (KISTLER 2613B) to measure theengine Crank Angle Degree (CAD) with 0.2 degree of resolution. Ahigh-speed data acquisition system DEWE5000 is connected tothe DEWESoft and DEWECa software for the relevant data pro-cessing. The engine is connected to an Eddy-Current brakedynamometer (30 kW, MAGTROL) via a chain and sprocketarrangement.
The fuel ON is controlled by blending of iso-octane, and n-heptane and is called Primary Reference Fuel (PRF) with an octanenumber of 0 (100% n-heptane) and 100 (100% iso-octane), in rela-tion to volume fraction [46]. Four combinations of fuel includingPRF 0, PRF 30, PRF 60 and PRF 95 are used in the experimentalstudy. The engine is set to a desired constant load and speed bymeans of injector's PW with the throttle maintained at WOT po-sition. Once the engine is fully warmed up and the CAI combustionis stable, the spark will be cut-off permanently and testing canbegin. Throughout data collection, and all variables (including In/Ex-EGR, fuel octane number, Tsc and AFR) are adjusted for eachdistinct test point to provide a stable CAI combustion state, asspecified in Table 2. In-cylinder pressure trace for 200 consecutiveengine cycles (0.2 CAD resolutions) is recorded at all of the steadystate test points, while In/Ex-EGR and the fuel ON values areadjusted.
Fig. 1. Schematic of the ex
3. In-cylinder gas exchange model and In-EGR and Ex-EGRestimation
The isothermal perfect-mixing model is assumed for scavengingprocess, where when the fresh charge enters, the cylinder will bemixed instantaneously with the cylinder charge to form a homo-geneous mixture at constant volume, pressure, and temperature[47]. Ideal gas law can be used since cylinder wall is consideredadiabatic so the charge gas has the same molecular weight withidentical and constant specific heats. A quantitative description ofthe two-stroke cycle scavenging process and its parameter termi-nology can be found in Ref. [48]. For this investigation, a semi-empirical method is used to estimate the real-time scavenging ef-ficiency (hsc) of the engine as is shown in Eq. (1) [49]. k0, k1, k2 arescavenging coefficients that have constant values and aremeasuredfor different ranges of the transfer port geometry arrangements.The value of each of the scavenging coefficients (loop scavengingwith five transfer port ducts) is shown in Table 1 [50].
hsc ¼ 1� Expðk0 þ k1Lþ k2LÞ (1)
where L is corrected delivery ratio which is the ratio of the deliv-ered mass of fresh charge (Mdel) to the mass of total gas trapped inthe cylinder (Mtr) and it is calculated from Eq. (2):
perimental test setup.
Fig. 2. HRR and MFB of a combustion phasing analysis.
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L ¼ MdelMtr
¼m�
Air � 60NS
�1þ 1
AFR
��Pepc�Vepc
R�Tepc
� (2)
where m�Air is engine's intake mass flow rate and R is specific gas
constant. Pepc, Vepc and Tepc are pressure, volume and temperatureof exhaust port closure, respectively. NS is engine speed in RPM andAFR is engine air-to-fuel ratio. Enthalpy balance Equation is used toestimate the Tepc when the piston is covered up completely theexhaust port and the effective compression is commenced, asshown in Eqs. (3) and (4) [28,29].
In this study it is assumed that all compositions of theinecylinder charge (including exhaust port closure mixture, scav-enging gas and residual gas) have an equal specific heats. Thus,Cpepc ¼ Cpsc ¼ Cpr and Eq. (4) can be derived as Eq. (5):
Tepc ¼ Tsc$hsc þ Trð1� hscÞ (5)
Residual gas temperature (Tr) is estimated as an average of bothTex and the inecylinder gas temperatures at the blow-downmoment (temperatures at the exhaust port opening, Tepo) [28,29]as shown in Eq. (6):
Tepc ¼ Tex þ Tepo2
(6)
Tepo is calculated under the assumption of adiabatic changes byusing Eq. (7) [47].
Tepo ¼ Tmax
�PepoPmax
�k�1k
(7)
where Tmax and Pmax are obtained from experimental data, and k isa computed polytropic exponent. The exponent (k) is estimated byutilizing Eq. (8) because the inecylinder mixture is highly diluted[51]. This correlation has been validated between a temperaturerange of 600e2000 K and for an AFR of 18e23:
After the completion of a scavenging process (closure of theexhaust port), a fraction of the burned gas will remain in thecombustion chamber. This residual gas can be quantified by the so-called residual gas ratio (g) according to Eq. (9):
g ¼ 1� hsc (9)
Typically, residual gas is present in all two-stroke cycle engines.It can be controlled, depending on the efficiency of scavengingprocess, but it cannot be completely removed. This amount of re-sidual gas ratio is shown by gres. And it is measured when the en-gine is operated at normal operating conditions i.e., neither In-EGRnor Ex-EGR are applied. In addition, the applied residual gas ratio(gap) is achieved when the engine is operated in the CAI mode bymeans of In-EGR or Ex-EGR application. Hence Eq. (1) can beinterpreted as follows:
The conventional operating condition (without In/Ex-EGR):
ðhscÞres ¼ 1� Expðk0 þ k1Lres þ k2LresÞ (10)
gres ¼ 1� ðhscÞres (11)
the CAI operating condition (with application of In/Ex-EGR):
ðhscÞap ¼ 1� Exp�k0 þ k1Lap þ k2Lap
�(12)
gap ¼ 1� ðhscÞap (13)
Correspondingly, after determining the residual gas ratio (g) inboth the applied and the inherent conditions, the In-EGR and theEx-EGR rates can be acquired in the same way as follows:
In� EGR ¼ �gap � gres
�� 100% (14)
Ex� EGR ¼ �gap � gres
�� 100% (15)
4. Calculation of heat release rate (HRR) and mass fractionburned (MFB)
Chemical compounds such as alkane or paraffin, including n-heptane and iso-octane will emerge in two-stages of auto-ignitionfor the CAI/HCCI combustion [52]. Two stage combustion iscomprised of low temperature heat release (LTHR) for cool flame(the 1st stage) and high temperature heat release (HTHR orHRRmax) for the main or hot flame (the 2nd stage) combustion iswell illustrated in Fig. 2. The delay between the low and hightemperature regime is associated with Negative Temperature Co-efficient (NTC) regime as depicted in figure [53].
Zero-dimensional heat release equation is used in order tocalculate the net Heat Release Rate (HRR) regarding Eq. (16) whichis in accordance with the following assumptions [54]: i) That thefirst law analysis can be applied by assuming that the engine chargedepicts ideal gas behavior; ii) That no heat transfer occurs from thecylinder iii); That there is no piston blow-by effect; iv) That com-bustion product dissociation is abandoned and v) That all ther-modynamic properties in the cylinder are uniform.
dQðqÞdq
¼�
1k� 1
�VðqÞdPðqÞ
dðqÞ þ�
kk� 1
�PðqÞdVðqÞ
dðqÞ (16)
The heat capacity ratio (k) is estimated by utilizing Eq. (8).Rassweiler and Withrow method is considered to determine theFuel Mass Fraction Burned (MFB) [55]. As shown in Fig. 2, twospecific points on the MFB curve are assigned to define the com-bustion timing characteristics. Therefore the start of auto-ignition(CA10), CA50 and the end of combustion (EOC) are denoted as
Fig. 3. Inecylinder pressure and net heat release rate (HRR) profiles for CAI operationrepresenting the upper-limit (knocking) as well as the lower-limit (misfiring), [n-heptane, rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 13e23, Tepc ¼ 530e600 K, In-EGR ¼ 7e37%, Ex-EGR ¼ 5e32% and WOT].
Fig. 4. Variation of LTHR and HTHR profiles in relation to In-EGR rate changes.
Fig. 5. Variation of LTHR and HTHR profiles in relation to Ex-EGR rate changes.
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the crank angle at 10%, 50% and 90% of MFB, respectively. Durationof combustion (DOC) is the crank angle between CA10 and EOC.CA10 is recognized as being the start of the main energy release(start of HTHR) and does not involve cool flame.
5. Results and discussion
The influence of three main variables, In-EGR, Ex-EGR and fuelON in relation to the major CAI combustion timing parameters andtheir cyclic variability are presented in the following section. Fig. 3represents the upper and lower limits of the CAI engine in terms ofinecylinder pressure and HRR profile. Throughout the datacollection, the engine was operated at the WOT setting, and allvariables, including In/Ex-EGR (5%e40%), fuel ON (0, 30, 60 and 95),Tepc (530e600 K) and m�
fuel (2.1e3.9 mg/cycle or AFR ¼ 15e20),were adjusted for each distinct test point to provide a stable CAIcombustion state. Data from each of the test points are averaged for200 consecutive cycles.
Referring to the figures shown, the engine can be run at CAImode most stably when the AFR is set between 15 until 20. Forleaner mixtures the engine was in some way unstable. This state ofengine operation is called Spark-Assisted CAI (SI-CAI). For ratherleaner settings, the engine was observed to encounter misfiring(AFR � 22).
5.1. Variation in net heat release rate (HRR)
The engine's main combustion-related parameters i.e.maximum net Heat Release Rate (HRRmax) and crank angle at netHeat Release Rate (qHRRmax), LTHR and qLTHR are taken into accountwith respect to the variation of In/Ex-EGR and fuel ON in this part ofexperiment. The influence of In-EGR changes on HHR profiles isshown in Fig. 4. It can be seen that when the concentration of In-EGR is increased, the timing of LTHR will be advanced slightly(enlarged window in the figure). In an overall look the timing ofHTHR will be advanced when In-EGR is increased. However, whenIn-EGR concentration is 15% timing of HTHR will be delayed. At thesame time, increasing of In-EGR rate will reduce the magnitude ofLTHR while the value of HTHR will increase. It is conceived that theinfluence of In-EGR on the magnitude of both LTHR and HTHR ismore considerable than their timing.
The reason for these occurrences is attributed to: In the case ofIn-EGR the most dominant effect is the charge heating effect. Bothpressure and temperature at the exhaust port closure (Pepc, Tepc)will increase since the percentage of In-EGR is raised therefore it
shortens the crank angle interval between the start of compression(qepc) and the crank angle at the temperature in which the coolflame develops (600e800 K) [29]. Due to the shorter crank angleinterval, the qLTHR appears in one earlier crank angle degree (CAD).Accordingly, the auto-ignition delay (AD) will be longer. In contrast,the advanced qLTHR will result the lower LTHR since in an earlycrank angle the piston position is far from top dead center (TDC) sothe inecylinder pressure and the LTHR will be lowered. In the caseof HTHR, the behavior can be interpreted in two ways. First, due tothe influence of the charge heating effect and the contraction of theexhaust port area by the In-EGR valve, Pepc will increase remark-ably. Consequently Tepc and HTHR will be increased. Second, oncethe auto-ignition delay is lengthened, the CA10 will be delayed.Therefore the piston can more compress the mixture before thestart of auto-ignition (CA10). Accordingly, the inecylinder pressure,temperature and HTHR will be increased [7,28,29,31,40].
In Fig. 5 the influence of Ex-EGR changes in relation to HRRprofile is illustrated. By increasing the percentage of Ex-EGR,although both the qLTHR and the qHRRmax are retarded to one laterCAD, the variations in qHRRmax are more substantial than variationsin qLTHR, which is negligible. Moreover, as Ex-EGR increases thequantity of LTHRwill increase and HTHRwill decrease, respectively.It is conceived that themost influential effect of Ex-EGR is to changethe magnitude of LTHR rather than change the qLTHR. In the case ofEx-EGR application, the most dominant effects are those of thermaland dilution effects. Adding Ex-EGR will increase the specific heatcapacity of the inecylinder charge. Even though the Tepc isincreased slightly, the mixture will take a longer time to heat up[7,28,29,33,40]. Accordingly, the crank angle interval between thestart of compression (qepc) and the crank angle at the temperature
Fig. 6. a. Variation of HRRmax in relation to In/Ex-EGR and fuel ON changes[rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K andWOT]. b. Variation ofqHRRmax in relation to In/Ex-EGR and fuel ON changes [rpm ¼ 2100, IMEP ¼ 2.2 bar,AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT].
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at which the cool flame develops becomes longer thus qLTHR will beretarded. Looking closely at the enlarged window, it can be seenthat the AD has increased thus the CA10 will be retarded accord-ingly. Furthermore, since the qLTHR is delayed, the piston is near to(TDC) so inecylinder pressure will increase and consequently themagnitude of LTHR will become larger. The reduction of HTHR isassociated with the reason, which is apart from twomost governedeffects of using Ex-EGR (i.e. thermal and charge dilution effects).Chemical effect is presumed to have a strong influence on thebehavior of the HTHR. When CA10 is too late (e.g. far from ATDC),not only the inecylinder pressure is low, but also the piston motionis downward (i.e. expansion stroke) thus HTHR will be affected.Furthermore, as matter of substitute of CO2 and H2O instead of O2(dilution effect) the overall reaction rate of combustion would besuppressed considerably.
Consequently, the principal difference between In-EGR and Ex-EGR application can be inferred that in the case of In-EGR appli-cation the charge-heating effect is more substantial since it pres-surizes the combustion chamber significantly, which leads tohigher Pepc and Tepc. Even though inecylinder charge compositionswill change (thermal and dilution effects) with In-EGR application,it is not considerable compared to changes in Tepc, Which is moresubstantial. In contrast, Ex-EGR will just mix burned gases to theintake charge leading to changes in inecylinder charge composi-tion and temperature. In this case the increasing in the mixturetemperature (i.e. Tepc) will not be considerable while the effect ofchanges in specific heats (thermal effect) and lack of oxygen(dilution effect) will be important significantly [7,28,29,32,33,40].
The most reliable criteria to characterize the combustionphasing is to examine two combustion-related parameters i.e.HRRmax and its corresponding crank angle qHRRmax [54,56]. Refer-ring to Fig. 6a, the general trend of HRRmax variation is downwardwhen fuel ON is increased. It can be also noted that HRRmax isinversely proportional with fuel ON. Furthermore, In-EGR increasesthe magnitude of HRRmax while Ex-EGR decreases that. And theslope ratio of each particular line (i.e. identical percentage ofwhether Ex-EGR or In-EGR) represents their tendency in relation tofuel ON changes. The figure also indicates HRRmax having a greaterpercentage of In-EGR is more likely to be influenced by variations offuel ON. However, when higher percentage of Ex-EGR is applied,HRRmax is less likely to be influenced by changes of fuel ON. Ingeneral, the tendency for HRRmax to vary with respect to fuel ON isvery much affected when In-EGR is applied rather than applicationof the Ex-EGR (comparing dashed ellipses A and B; and dashedrectangular C and D). It is interpreted that the state of an engine'soperating condition with lower percentage of Ex-EGR is moredependent on the changing of ON.
Combustion-related parameter qHRRmax is taken into consider-ation in conjunction with variation of the In/Ex-EGR and the fuelON as shown in Fig. 6b. It can be interpreted that when the fuel ONis increased the timing of qHRRmax will be retarded to one later CADnot only in the case of In-EGR, but also for Ex-EGR application. In ageneral manner, there is same tendency variation of qHRRmax inrelation to changes in both In-EGR and Ex-EGR rates (slope ratiosare nearly the same). In general, it can be deduced that when thepercentage of In-EGR is increased, qHRRmax will be advanced while itwill be delayed as the percentage of Ex-EGR increases.
5.2. Variation in CA10, Tmean, TCA10
As mentioned earlier, 10% of the integrated heat release of thecombustion cycle (MFB curve) is captured to identify the start ofauto-ignition (CA10). The influence of fuel ON variations on CA10is shown in Fig. 7a and b when In-EGR and Ex-EGR are applied,respectively. In Fig. 7a, as the fuel ON is increased, the CA10 will be
retarded to a later CAD. For higher concentration of In-EGR it canbe seen that the variations of CA10 due to ON changes is notsubstantial when ON is more than 20. This trend can also be un-derstood when higher rate of Ex-EGR is applied as shown inFig. 7b. The reason is attributed to this fact that at high concen-tration of In/Ex-EGR (high/over diluted) the start of auto-ignitionis not much influenced by ON variations and there is no strongcorrelation between CA10 and ON [25,28,29]. From the trendsshown, it can be deduced that the In-EGR has a retarding effect onCA10 except for the condition, which in In-EGR is at maximumconcentration (37%). The reasoning for that can be inferred asfollows: when the percentage of In-EGR is increased, LTHR andqLTHR will be decreased and advanced respectively. In the case ofthe highly diluted In-EGR, the charge heating effect will no longerbe dominant in this case to increase Tepc leading to the advance-ment of qLTHR. The thermal effect will be more considerable hereleading to one delayed qLTHR with shorter auto-ignition delay (AD)[7,26,28,29,40]. Accordingly the shortening of AD will result in theadvancement of CA10. In addition, it should be noted that when alow fuel ON is used, the tendency for the variation of CA10 (inrelation to In-EGR changes) is greater than the fuel with high ON(ellipse A and B).
Fig. 7b represents the variation of CA10 in conjunction with thefuel ON when Ex-EGR is applied. From the data shown, it can beinferred that a higher fuel ON and higher Ex-EGR will retard CA10to a later CAD. However when the maximum percentage of the Ex-EGR is performed (32%) the behavior of CA10 variation does notfollow the regular pattern. The can be understood that in the case ofhighly diluted Ex-EGR although the thermal and dilution effects aremostly governed phenomenon, the charge heating effect becomes
Fig. 7. a. Variation of CA10 in relation to In-EGR changes [PRF ¼ 0e95, rpm ¼ 2100,IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. b. Variation of CA10 inrelation to Ex-EGR changes [PRF ¼ 0e95, rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21,Tepc ¼ 530e600 K and WOT].
Fig. 8. a. Cyclic variation of CA10 due to changes in In-EGR [PRF ¼ 0e95, rpm ¼ 2100,IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. b. Cyclic variation of CA10due to changes in Ex-EGR [PRF ¼ 0e95, rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21,Tepc ¼ 530e600 K and WOT].
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considerable as well. Thus, when the AD becomes shorter the CA10will be advanced accordingly. As well, the sensitivity of CA10 versusEx-EGR is the same when ON is changed (ellipses A and B).
Standard Deviation (STD) is employed in order to examine thecyclic variation of combustion-related parameters [57,58]. In orderto analyze the cyclic variability of the CAI combustion (in terms ofauto-ignition commencement), sixteen test points with respect tocyclic variation of CA10 (STDCA10) are taken into considerationwhen the In-EGR is applied as illustrated at Fig. 8a. In the figure thetest points placed in one distinct dashed ellipse have an identicalpercentage of In-EGR but different fuel ON (denoted as cross). Theaveraged STDCA10 in each identical In-EGR is denoted as one hol-lowed circle mark while connecting dashed line through themrepresents the trend line. In addition, a further illustration of thecurve of fit of all averaged STDCA10 is drawn throughout the graph(denoted as triangles). It can be inferred that In-EGR decreases theSTDCA10. The higher the In-EGR percentage, the lower the cyclicvariation of CA10 will be. In addition, test points with In-EGR¼ 15%(ellipse A), have highest dependence on the fuel ON changes. Incontrast, test points having In-EGR ¼ 37% (ellipse B) are lesscorrelated to the variation of fuel ON. In Fig. 8b it can be also seenthat STDCA10 has increased (triangle points) when the percentage ofEx-EGR is increased. The higher the Ex-EGR percentage, the higherthe STDCA10 becomes. Furthermore, it can be interpreted that thetest points having Ex-EGR ¼ 12% (ellipse A) are less influenced bythe changing of fuel ON. On the other hand, the test points withmaximum concentration of Ex-EGR ¼ 32% (ellipse B) representhighest tendency in relation to fuel ON changes.
The effect of In/Ex-EGR variation with respect to meaninecylinder gas temperature (Tmean) history is elucidated in Fig. 9a
and b when PRF 30 is used. Additionally, the variation of temper-ature at start of auto-ignition (auto-ignition temperature, TCA10)and maximum inecylinder gas temperature (Tmax) can beperceived from the figures. As is evidently shown, In-EGR developsthe history of Tmean while Ex-EGR suppresses it. Fig. 9c and d illus-trate the variation of TCA10 on relation to In-EGR and Ex-EGRchanges respectively. As evident from the figures, TCA10 will beincreased as the percentage of In/Ex-EGR is improved. However,this increment is more substantial in the case of In-EGR applicationrather than Ex-EGR. Hence, it can be understood that the higher theIn/Ex-EGR percentage, the greater the TCA10 will become. Further-more, It should be noted that TCA10 will be decreased in both casesof In/Ex-EGR application when their concentration has reached tomaximum (In-EGR ¼ 37% and Ex-EGR ¼ 32%). The reason is as thesame as discussed concerning CA10 in Section 5.2. In addition, thefuels having higher ON present higher magnitude of TCA10 eitherwith In-EGR or Ex-EGR application.
5.3. Variation in MFB, DOC and CA50
The variation of MFB profile is demonstrated with respect to In-EGR changes in Fig. 10a. It can be seen that when the percentage ofIn-EGR increases, the MFB profiles' slope will be sharper (i.e. fastcombustion) until In-EGR ¼ 26% and for In-EGR ¼ 37% the slopewill be decreased (over-diluted In-EGR). CAI combustion phasingis not fully characterized by CA10 because different DOC can occurat the same CA10. In order to understand the cyclic variability ofCAI combustion phasing, cyclic variations of DOC must also beconsidered. Variation in DOC results when some cycles have fastcombustion (small DOC) and some cycles have slow combustion(large DOC) [54]. The variation of DOC in accordance with fuel ON
Fig. 9. a. Variation of Tmean due to In-EGR changes [PRF ¼ 30, rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. b. Variation of Tmean due to Ex-EGR changes[PRF ¼ 30 rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. c. Variation of TCA10 due to In-EGR changes [PRF ¼ 0e95, rpm ¼ 2100, IMEP ¼ 2.2 bar,AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. d. Variation of TCA10 due to Ex-EGR changes [PRF ¼ 0e95, rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT].
Fig. 10. a. Variation of MFB profiles in accordance with In-EGR changes [n-heptane,rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. b. Varia-tion of DOC in accordance with In/Ex-EGR and fuel ON changes [rpm ¼ 2100,IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT].
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changes is represented in Fig. 10b when the In-EGR and Ex-EGRare applied. In general it can be inferred that the DOC will belengthened when the fuel ON is increased indicating that thevariation of DOC is directly proportional to the variation of the fuelON. In addition, as the rate of In-EGR increases, the combustionwill be faster (lower DOC) while application of Ex-EGR increasesthe DOC (slower combustion). It is worth noting that using eitherIn-EGR or Ex-EGR effects the EOC rather than the CA10, as clarifiedin Fig. 10a.
Crank angle at which 50% of the inecylinder mass fraction isburned (CA50) represents another trustable combustion-relatedparameter, as the DOC can be different with a same timing of theCA10. This parameter suggests other reliable means instead of DOCto quantify the auto-ignition timing of CAI combustion [59,60]. Thevariation of CA50 in relation to In-EGR changes is presented inFig. 11a. Generally, the trend of the curves is downward excludingthe test points with In-EGR ¼ 37%. This is attributed to the over-dilution effect in the case of In-EGR (highly In-EGR diluted) appli-cation, which is discussed earlier in Section 5.2. Apart from themaximum rate of In-EGR, when the percentage of In-EGR improvesthe timing of CA50 will be advanced accordingly. A comparison ofthe dashed ellipses A and B indicates that the test points withlowest percentage of In-EGR (7%) are less likely to be affected bychanges in the fuel ON. Fig. 11b describes the effect of Ex-EGRchanges on the behavior of CA50. As a whole, the overall trend ofthe curves is upward. This means that, when the Ex-EGR percent-age is increased the CA50 will be delayed to a later CAD. Further-more, from the slope ratio of the curves it can be deduced that atthe higher concentration of Ex-EGR, the variation of CA50 is moresensitive to the fuel ON changes rather than lower concentration ofEx-EGR (refer to ellipses A and B).
In Fig. 12a standard deviation of the CA50 (STDCA50) is presentedin conjunction with In-EGR variations. With regard to the figure,the trend shows the higher the In-EGR percentage, the lower theSTDCA50 becomes. It means that STDCA50 is inversely proportional toIn-EGR. It is also noted that test points having lowest percentage of
Fig. 11. a. Variation of CA50 due to In-EGR changes [PRF ¼ 0e95, rpm ¼ 2100,IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. b. Variation of CA50 due toEx-EGR changes [PRF ¼ 0e95, rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21,Tepc ¼ 530e600 K and WOT].
Fig. 12. a. Cyclic variation of CA50 in relation to In-EGR changes [PRF ¼ 0e95,rpm ¼ 2100, IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT]. b. Cyclicvariation of CA50 in relation to Ex-EGR changes [PRF ¼ 0e95, rpm ¼ 2100,IMEP ¼ 2.2 bar, AFR ¼ 15e21, Tepc ¼ 530e600 K and WOT].
Fig. 13. STDCA50 versus STDCA10 representing the engine's CAI combustion regions [Inrelation to 32 test points through 200 averaged consecutive cycles].
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In-EGR¼ 7% are less likely to be influenced by fuel ON changes. Theinfluence of Ex-EGR on variations of STDCA50 is illustrated inFig.12b. It can be deduced that STDCA50 is directly proportional withthe Ex-EGR percentage. Thus, the higher the Ex-EGR percentage,the higher the STDCA50 will be. As evident from the dashed ellipses,the sensitivity of the STDCA50 in relation to fuel ON changes willincrease considerably when the Ex-EGR is at highest concentration(32%).
Standard deviation on both CA10 and CA50 can be used toidentify the engine's CAI combustion operating regions. To under-take this, STDCA50 is plotted in accordance with STDCA10 in oneparticular graph as illustrated in Fig. 13. It covers 32 test points ofengine operating condition for 200 consecutive cycles. Dependingon the type of fuel used and the level of applied In/Ex-EGR, fourdistinct zones are realized, each representing a particular state ofengine operating condition.
ZONE I demonstrates the most stable condition of the engine'sCAI combustion inwhich it can be run in the CAI mode without anyinterruption or irregularity. Accordingly here neither knocking normisfiring is observed here.
ZONE II represents knocking region, or the engine's upper limitincluding test points with most advanced combustion owing toeither lower fuel ON or excessive rate of In-EGR.
ZONE III manifests CAI-SI region in which the engine operationcan be continued by means of spark plug (called spark-assistedmode) and the combustion mode is being varied interchangeably.Here the cylinder charge will burn incompletely since Tepc is nothigh enough to reach the auto-ignition point (partial burn).
ZONE IV represents misfire region or engine's lower limitincluding test points with most retarded combustion that isattributed to either highly diluted Ex-EGR or higher fuel ON.
6. Conclusions
A detailed experimental study was carried out on a CAI two-stroke cycle engine operating at constant load and speed. Cyclicvariability and combustion phasing characteristics of the CAIcombustion were investigated in accordance with the variation ofIn-EGR, Ex-EGR and fuel ON. It is deduced that the characteristics ofCAI combustion in terms of combustion-related parameters (LTHR,HTHR, CA10, TCA10) are strongly influenced by In/Ex-EGR rate andfuel ON, which can either advance or retard the ignition timing ofCAI combustion. Furthermore they suggest significant aspectsregarding controlling the cool flame, temperature and timing ofauto-ignition (TCA10 and CA10) of CAI combustion. In general, theresults can be summarized as follows:
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1. The main influence of both In-EGR and Ex-EGR is changes inmagnitude of cool and hot flame (LTHR and HTHR) ratherchanges in than their timing (qLTHR and qHRRmax).
2. Variation of HRRmax (HTHR) in accordance with fuel ON is moresubstantial when the concentration of In-EGR is high howeverthis variation is more significant when the concentration of Ex-EGR is low.
3. Cyclic variation of CA10 (STDCA10) is more sensitive to fuel ONchanges at low In-EGR rate. However, it is more likely to beinfluenced by fuel ON in high Ex-EGR rate.
4. The timing of CA10 and its cyclic variability (STDCA10) areinfluenced more by Ex-EGR than In-EGR.
5. In-EGR will increase both the TCA10 and Tmax; however Ex-EGRwill increase TCA10 and decrease Tmax.
Acknowledgements
The authors would like to express their appreciation tothe Research Management Centre (RMC, Vote No:Q.J130000.7114.02J14) of UTM for the financial support. The tech-nical assistance rendered by the laboratory technicians of theAutomotive Development Centre (ADC) throughout the entireperiod of the experimental work is very much appreciated.
References
[1] M. Cerit, H.S. Soyhan, Thermal analysis of a combustion chamber surroundedby deposits in an HCCI engine, Appl. Therm. Eng. 50 (2013) 81e88.
[2] S.E. Hosseini, A.M. Andwari, M.A. Wahid, G. Bagheri, A review on green energypotentials in Iran, Renew. Sustain. Energy Rev. 27 (2013) 533e545.
[3] C. Cinar, €O. Can, F. Sahin, H.S. Yucesu, Effects of premixed diethyl ether (DEE)on combustion and exhaust emissions in a HCCI-DI diesel engine, Appl.Therm. Eng. 30 (2010) 360e365.
[4] S. Saxena, I. Bedoya, Fundamental phenomena affecting low temperaturecombustion and HCCI engines, high load limits and strategies for extendingthese limits, Prog. Energy Combust. Sci. 39 (5) (October 2013) 457e488.
[5] R.H. Thring, Homogeneous-charge Compression-ignition (HCCI) Engines,1989. SAE Paper 892068.
[6] Q. Fang, J. Fang, J. Zhuang, Z. Huang, Influences of pilot injection and exhaustgas recirculation (EGR) on combustion and emissions in a HCCI-DI combustionengine, Appl. Therm. Eng. 48 (2012) 97e104.
[7] H. Zhao, HCCI and CAI Engines for the Automotive Industry, Woodhead Pub.,2007.
[8] B. Bahri, A. Aziz, M. Shahbakhti, M. Said, Analysis and modeling of exhaust gastemperature in an ethanol fuelled HCCI engine, J. Mech. Sci. Technol. 27 (2013)3531e3539.
[9] B. Bahri, A.A. Aziz, M. Shahbakhti, M.F. Muhamad Said, Understanding anddetecting misfire in an HCCI engine fuelled with ethanol, Appl. Energy 108(2013) 24e33.
[10] M.T. García, F.J.J.-E. Aguilar, T.S. Lencero, J.A.B. Villanueva, A new heat releaserate (HRR) law for homogeneous charge compression ignition (HCCI) com-bustion mode, Appl. Therm. Eng. 29 (2009) 3654e3662.
[11] S. Onishi, S.H. Jo, K. Shoda, P.D. Jo, S. Kato, Active Thermo-atmosphere Com-bustion (ATAC) e a New Combustion Process for Internal Combustion Engines,1979. SAE Paper 790501.
[12] M.A. Amin, A.A. Azhar, Homogenous charge compression ignition (HCCI)technique: a review for application in two-stroke gasoline engines, Appl.Mech. Mater. 165 (2012) 53e57.
[13] A.M. Andwari, A.A. Aziz, M.F. Muhamad Said, Z. Abdul Latiff, Controlled auto-ignition combustion in a two-stroke cycle engine using hot burned gases,Appl. Mech. Mater. 388 (2013) 201e205.
[14] M. Noguchi, Y. Tanaka, T. Tanaka, Y. Takeuchi, A Study on Gasoline EngineCombustion by Observation of Intermediate Reactive Products during Com-bustion, 1979. SAE Paper 790840.
[15] Y. Ishibashi, M. Asai, A Low Pressure Pneumatic Direct Injection Two-strokeEngine by Activated Radical Combustion Concept, 1998, 790840 980757.
[16] K. Tsuchiya, S. Hirano, M. Okamura, T. Gotoh, Emission Control of Two-strokeMotorcycle Engines by the Butterfly Exhaust Valve, 1980. SAE Paper 800973.
[17] M. Asai, T. Kurosaki, K. Okada, Analysis on Fuel Economy Improvement andExhaust Emission Reduction in a Two-stroke Engine by Using an ExhaustValve, 1995. SAE Paper 951764.
[18] P. Duret, A new generation of two-stroke engines for the future?, in: Pro-ceedings of the International Seminar Held in Rueil-Malmaison, France,November 29-30, 1993, �Editions Technip, 1993.
[19] P. Duret, A. Ecomard, M. Audinet, A New Two-stroke Engine withCompressed-air Assisted Fuel Injection for High Efficiency Low EmissionsApplications, 1988. SAE Paper 880176.
[20] P. Duret, J.-F. Moreau, Reduction of Pollutant Emissions of the IAPAC Two-stroke Engine with Compressed Air Assisted Fuel Injection, 1990. SAE Paper900801.
[21] P. Duret, S.p. Venturi, Automotive Calibration of the IAPAC Fluid DynamicallyControlled Two-stroke Combustion Process, 1996. SAE Paper 960363.
[22] P. Duret, A new generation of engine combustion processes for the future?, in:Proceedings of the International Congress, Held in Rueil-Malmaison, France,November, 26e27, 2001, Editions Technip, 2002.
[23] P. Duret, J.-C. Dabadie, J. Lavy, J. Allen, D. Blundell, J. Oscarsson,G. Emanuelsson, M. Perotti, R. Kenny, G. Cunningham, The Air Assisted DirectInjection ELEVATE Automotive Engine Combustion System, 2000. SAE Paper2000-01-1899.
[24] Y. Ishibashi, M. Asai, Improving the Exhaust Emissions of Two-stroke Enginesby Applying the Activated Radical Combustion, 1996. SAE Paper 960742.
[25] Y. Ishibashi, Basic Understanding of Activated Radical Combustion and itsTwo-stroke Engine Application and Benefits, 2000. SAE Paper 2000-01-1836.
[26] N. Iida, Combustion Analysis of Methanol-fueled Active Thermo-atmosphereCombustion (ATAC) Engine Using a Spectroscopic Observation, 1994. SAEPaper 940684.
[27] N. Iida, Y. Yamasaki, S. Sato, K. Kumano, Y. Kojima, Study on Auto-ignition andCombustion Mechanism of HCCI Engine, 2004. SAE Paper 2004-32-0095.
[28] K. Goto, A. Iijima, K. Yoshida, H. Shoji, Analysis of the Characteristics of HCCICombustion and ATAC Combustion Using the Same Test Engine, SAE Inter-national, 2004.
[29] A. Iijima, K. Yoshida, H. Shoji, A Comparative Study of HCCI and ATAC Com-bustion Characteristics Based on Experimentation and Simulations Influenceof the Fuel Octane Number and Internal EGR on Combustion, SAE Interna-tional, 2005.
[30] N. Jeuland, X. Montagne, P. Duret, Engine and Fuel Related Issues of GasolineCAI (Controlled Auto-ignition) Combustion, 2003. SAE Paper 2003-01-1856.
[31] Y. Zhang, H. Zhao, M. Ojapah, A. Cairns, CAI combustion of gasoline and itsmixture with ethanol in a 2-stroke poppet valve DI gasoline engine, Fuel 109(2013) 661e668.
[32] T. Takei, N. Iida, Study on Auto-ignition and Combustion Completion of N-Butane in a Two-stroke Homogeneous Charge Compression Ignition (HCCI)Engine, 2002.
[33] Y. Zhang, H. Zhao, M. Ojapah, A. Cairns, Effects of injection timing on CAI oper-ation in a 2/4-stroke switchable GDI engine, SAE Int. J. Engines 5 (2011) 67e75.
[34] Y. Zhang, H. Zhao, M. Ojapah, A. Cairns, Experiment and Analysis of a DirectInjection Gasoline Engine Operating with 2-stroke and 4-stroke Cycles ofSpark Ignition and Controlled Auto-ignition Combustion, SAE International,2011.
[35] G.P. Merker, C. Schwarz, R. Teichmann, Combustion Engines Development:Mixture Formation, Combustion, Emissions and Simulation, Springer, 2011.
[36] P.G. Aleiferis, A.G. Charalambides, Y. Hardalupas, A.M.K.P. Taylor, Y. Urata,Autoignition Initiation and Development of N-heptane HCCI CombustionAssisted by Inlet Air Heating, Internal EGR or Spark Discharge: an OpticalInvestigation, SAE International, 2006.
[37] H. Machrafi, S. Cavadias, J. Amouroux, Influence of fuel type, dilution andequivalence ratio on the emission reduction from the auto-ignition in anhomogeneous charge compression ignition engine, Energy 35 (2010)1829e1838.
[38] N. Sarabchi, R. Khoshbakhti Saray, S.M.S. Mahmoudi, Utilization of waste heatfrom a HCCI (homogeneous charge compression ignition) engine in a tri-generation system, Energy 55 (2013) 965e976.
[39] A. Tsolakis, A. Megaritis, D. Yap, Application of exhaust gas fuel reforming indiesel and homogeneous charge compression ignition (HCCI) engines fuelledwith biofuels, Energy 33 (2008) 462e470.
[40] H. Zhao, Z. Peng, N. Ladommatos, Understanding of controlled autoignitioncombustion in a four-stroke gasoline engine, Proc. Institution Mech. Eng. PartJ. Automob. Eng. 215 (2001) 1297e1310.
[41] M.T. Garcia, F.J. Jim�enez-Espadafor Aguilar, J.A. Becerra Villanueva,E.C. Trujillo, Analysis of a new analytical law of heat release rate (HRR) forhomogenous charge compression ignition (HCCI) combustion mode versusanalytical parameters, Appl. Therm. Eng. 31 (2011) 458e466.
[42] A. Khaliq, S.K. Trivedi, I. Dincer, Investigation of a wet ethanol operated HCCIengine based on first and second law analyses, Appl. Therm. Eng. 31 (2011)1621e1629.
[43] H. Liu, P. Zhang, Z. Li, J. Luo, Z. Zheng, M. Yao, Effects of temperature in-homogeneities on the HCCI combustion in an optical engine, Appl. Therm.Eng. 31 (2011) 2549e2555.
[44] I. Sher, D. Levinzon-Sher, E. Sher, Miniaturization limitations of HCCI internalcombustion engines, Appl. Therm. Eng. 29 (2009) 400e411.
[45] H. Zhao, N. Ladommatos, Engine Combustion Instrumentation and Di-agnostics, Society of Automotive Engineers, Warrendale, PA, 2001, p. 842.
[46] X.-c. Lü, W. Chen, Y.-c. Hou, Z. Huang, Study on the Ignition, Combustion andEmissions of HCCI Combustion Engines Fueled with Primary Reference Fuels,SAE International, 2005.
[47] J.B. Heywood, E. Sher, S.o.A. Engineers, The Two-stroke Cycle Engine: itsDevelopment, Operation, and Design, Taylor & Francis, 1999.
[48] K. Nishida, H. Sakuyama, T. Kimijima, Improvement of Fuel Economy Using aNew Concept of Two-stroke Gasoline Engine Applying Stratified-charge Auto-ignition, The Automotive Research Association of India, 2009.
[49] G.P. Blair, S.P.A.S.P. Committee, Advances in Two-stroke Cycle Engine Tech-nology, Society of Automotive Engineers, 1989.
A. Mahmoudzadeh Andwari et al. / Applied Thermal Engineering 71 (2014) 248e258258
[50] G.P. Blair, The Basic Design of Two-stroke Engines, Society of AutomotiveEngineers, 1990.
[51] J. Chang, O. Güralp, Z. Filipi, D. Assanis, T.-W. Kuo, P. Najt, R. Rask, New HeatTransfer Correlation for an HCCI Engine Derived from Measurements ofInstantaneous Surface Heat Flux, SAE International, 2004.
[52] S. Tanaka, F. Ayala, J.C. Keck, J.B. Heywood, Two-stage ignition in HCCI com-bustion and HCCI control by fuels and additives, Combust. Flame 132 (2003)219e239.
[53] M. Yao, Z. Zheng, H. Liu, Progress and recent trends in homogeneous chargecompression ignition (HCCI) engines, Prog. Energy Combust. Sci. 35 (2009)398e437.
[54] J.B. Heywood, Internal Combustion Engine Fundamentals, McGraw-Hill, 1988.[55] G.M. Rassweiler, L. Withrow, Motion Pictures of Engine Flames Correlated
with Pressure Cards, SAE International, 1938.[56] N. Ozdor, M. Dulger, E. Sher, Cyclic Variability in Spark Ignition Engines a
Literature Survey, SAE International, 1994.[57] H.S. Soyhan, H. Yasar, H. Walmsley, B. Head, G.T. Kalghatgi, C. Sorusbay,
Evaluation of heat transfer correlations for HCCI engine modeling, Appl.Therm. Eng. 29 (2009) 541e549.
[58] H. Yasar, H.S. Soyhan, H. Walmsley, B. Head, C. Sorusbay, Double-wiebefunction: an approach for single-zone HCCI engine modeling, Appl. Therm.Eng. 28 (2008) 1284e1290.
[59] H. Liu, Z. Zheng, M. Yao, P. Zhang, Z. Zheng, B. He, Y. Qi, Influence of tem-perature and mixture stratification on HCCI combustion using chem-iluminescence images and CFD analysis, Appl. Therm. Eng. 33e34 (2012)135e143.
[60] M. Shahbakhti, C.R. Koch, Characterizing the cyclic variability of ignitiontiming in a homogeneous charge compression ignition engine fuelled with n-heptane/iso-octane blend fuels, Int. J. Engine Res. 9 (2008) 361e397.
Glossary
A/BTDC: after/before top dead centerAD: auto-ignition delay
CO2: carbon dioxideCAI: controlled auto-ignitionCI: compression ignitionCOV: coefficient of variationDOC: duration of combustionEx-EGR: external exhaust gas recirculationHCCI: homogeneous charge compression ignitionIn-EGR: internal exhaust gas recirculationK0, K1, K2: scavenging coefficientsk: heat capacity ratioLap: applied corrected delivery ratioLres: irremovable corrected delivery ratiom�
fuel: fuel mass flow rateNOx: nitric oxidesNTC: negative temperature coefficientON: octane numberPW: pulse widthPepc: inecylinder pressure at exhaust port closurePmax: maximum inecylinder pressureSTD: standard deviationCA10: start of auto-ignitionSI: spark ignitionTex: exhaust gas temperatureTepc: inecylinder gas temperature at exhaust port closureTsc: scavenging gas temperatureTr: residual gas temperatureuHC: unburned hydrocarbonWOT: wide-open throttleqHRRmax: crank angle at HRRmaxqLTHR: crank angle at LTHRqPmax: crank angle at Pmaxgres: irremovable residual gas ratiogap: applied residual gas ratio(hsc)res: irremovable scavenging efficiency(hsc)ap: applied scavenging efficiency
