Experimental thermal analysis of cylinder block and head of a bi-fuel turbocharged engine

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MeccanicaAn International Journal of Theoreticaland Applied Mechanics AIMETA ISSN 0025-6455Volume 47Number 8 Meccanica (2012) 47:1987-2004DOI 10.1007/s11012-012-9569-7

Experimental thermal analysis of cylinderblock and head of a bi-fuel turbochargedengine

J. Ghorbanian & M. Ahmadi

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Meccanica (2012) 47:1987–2004DOI 10.1007/s11012-012-9569-7

Experimental thermal analysis of cylinder block and headof a bi-fuel turbocharged engine

J. Ghorbanian · M. Ahmadi

Received: 9 June 2010 / Accepted: 20 June 2012 / Published online: 26 September 2012© Springer Science+Business Media B.V. 2012

Abstract The beauty and application of thermal anal-ysis concept leads researchers to one of the main stepsfor thermomechanical design during engine develop-ment. The durability and output potential of such en-gines is strongly linked to the operating temperatureof certain key components. Thus, accurate tempera-ture predictions are an essential pre-requisite to thecontinuing engine evolution. From material sciencepoint of view, temperature field in engine main com-ponents like cylinder head and block is required toevaluate component functionality under specific loadand conditions. Moreover, need for more power andless weight is one of the most important targets in en-gine design, especially in the case of alternative fuelengines. In order to look at this issue, authors zeroin on wide-ranging experimental and analytical studyto investigate temperature fields in cylinder head andblock of a recently developed turbocharged bi-fuel en-gine. A bi-fuel turbocharged engine (CNG and gaso-line) were equipped with more than 40 sets of ther-mocouples and a comprehensive thermal survey wascarried out on the fired engine in the various condi-tions. Thermocouples were installed on different posi-tions of the cylinder block and head to measure mate-rial temperatures. Experiments were done both in nat-ural gas and gasoline mode to compare the results. An

J. Ghorbanian (�) · M. Ahmadi6th Km Karaj Makhsous Road, Tehran, Irane-mail: [email protected]

analytical comparison was made between natural gasand gasoline modes to understand root and effect ofheat transfer differences. The presented thermal anal-ysis could be helpful to know material requirementsto design and develop turbocharged natural gas enginewhich is a good candidate as an alternative fuel en-gine.

Keywords Heat transfer · Experimental thermalanalysis · Natural gas engine · Thermal survey ·Bi-fuel turbocharged engine

NomenclatureaTDC After top dead centerbTDC Before top dead centerCNG Compressed natural gasTC Coolant temperature at engine outletmf Fuel mass flow ratemAir Inlet air mass flow rate�TAir Intake air temperature difference due to

vaporization of gasolineIC engine Internal combustion engineλ lambdaHV Latent heat of vaporizationLHV Low heating valueNG Natural gasNA engine Naturally aspirated engineCP(Air) Specific heat at constant pressure(A/F)S Stoichiometric air/fuel ratioTDC Top dead centerTC engine Turbocharged engine

Author's personal copy

mailto:[email protected]

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TSolid Material temperatureWOT Wide open throttle

1 Introduction

In recent years using natural gas (NG) in automo-tive engines shows its potential as a good alternativefuel for future engines. Mass produced automobilesequipped with NG dedicated engines, commercial bi-fuel manufactured engines, and growing number ofconverted gasoline vehicles working with natural gasrepresent global prone to use NG in powertrains. Onthe other hand, technical weak points of these engineslike relatively lower durability life, additional equip-ments for fuel storage, lower specific power and mi-nor natural gas lubricity prompt run more investiga-tions. Importance of this issue is highlighted for au-tomotive industry when considering worthwhile ben-efits of natural gas engines like inherent potential forlower pollution and excellent knock resistance due tothe higher octane number which promises cleaner andmore efficient engines. One of the most imperativeweak points of NG engines is lower specific powercompared to gasoline ones which subsequently leadsto a lower mileage in the vehicle. In comparison withgasoline fuel, natural gas has lower heating value perengine charged mass at stoichiometric mixture, andvolumetric efficiency is lower in natural aspirated NGengines due to gaseous form and lack of latent heat[1–3]. Some solutions have been suggested, tested, andemployed to improve NG engines performance, dura-bility and reliability. Lapetz et al. [4] developed a FordCNG bi-fuel pickup truck. To control emissions andinsure safety, they modified the base vehicle’s config-uration for conversion to bi-fuel CNG operation. Nat-ural gas has a lower flame speed than gasoline. Thiscauses the total combustion duration prolonged com-pared with gasoline fuel [5]. In designing a turbulenteffect in order to increase the natural gas flame speed,Johansson and Olsson [6, 7] developed ten differentgeometries of combustion chamber. The results showa high correlation between in cylinder turbulence andrate of heat release in combustion process. However,the results also showed that geometries that gave thefastest combustion would also gave the higher NOconcentrations. Evans et al. [8] investigated combus-tion chamber design for fast burning of natural gas.Their studies based on the principle of using squish

motion to generate a series of jets directed towards thecenter of the chamber just prior to ignition. Swain etal. [9] have been studied the effects of hydrogen addi-tion on the natural gas engine operation. According totheir results, adding hydrogen into the CNG/air mix-ture had adverse effect on the combustion delay andincreased the burning rate.

Zuo and Zhao [10] developed a quasi-dimensionalmodel for analysis of combustion processes in SI pre-chamber natural gas engine. They have used two sub-models to simulate turbulence intensity in cylinder andmodeling of jet orifices in pre-chamber. They verifiedtheir simulation code with experimental data. Mani-vannan et al. [11] studied lean burn strategy for reduc-ing emissions of natural gas spark ignition engines.They considered performance and emissions charac-teristics of a SI lean burn natural gas engine. Also,they studied effects of fuel composition, combustionchamber geometry, combustion modeling, and burningrate models, pre-chamber and after treatment on theseengines. Chiodi et al. [12] have investigated mixtureformation and combustion process in a CNG engineby using a fast response 3D CFD simulation. Volpatoet al. [13] studied engine management for multi-fuelplus compressed natural gas vehicles. Aslam et al. [14]have retrofitted a conventional 1.5L, 4-cylinder ProtonMagma gasoline engine for running with CNG. Theytested the bi-fuel engine for CNG and gasoline fuelsand a comparative study of emissions has been madefor both fuels.

In this paper turbocharging bi-fuel or dedicated NGengines introduced as a remarkable feat which is help-ful to achieve acceptable performance level in both gasand gasoline modes. However, thermal and mechan-ical loads of pistons, valves, bearings, cylinder headand block will be higher than naturally aspirated en-gine and it may force redesigning some componentsor changing the material. So the authors focus on in-vestigating temperature fields in cylinder head andblock of a recently developed turbocharged bi-fuel en-gine.

1.1 Engine heat transfer

The importance of engine heat transfer is clear due toits role in engine efficiency, emission, endurance, fric-tion, and lubrication. In a typical IC engines, sourceof heat generation is burning of air-fuel mixture in-side the cylinder(s). In these engines less than a half of


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the generated heat in the working fluid produces powerand other part is dissipated from different ways in en-gine via three modes: conduction, convection and radi-ation. It is therefore, a complex process which dependsupon numerous parameters like: properties of workingfluid and combustion parameters, engine geometry andmaterial specifications, cooling and lubrication systemdesign and characteristics of employed fluids insidethem. The complexity would be greater when consid-ering that heat transfer takes place under conditions ofvarying gas pressure, temperature and the surface areaof the combustion chamber through the cycle.

Engines have evolved rapidly in the last decadelargely as a result of emissions legislation, but heattransfer modeling techniques have not kept pace withthese developments. The commonly used models topredict the heat flux arising from combustion gases,namely those of Annand [15], and Woschni [16], arebased on outdated engine technology. At the timewhen these methods were proposed, most of the mod-ern features already mentioned in the case of new en-gines did not exist or were not widely used. Thesemodels predict the heat transfer coefficient very goodaround the combustion stroke. However, at the othercrank angles specifically in expansion stroke they havedeviation from the actual data [17]. Furthermore, theWoschni and Annand models are not able to predictthe local convective heat transfer in a spark ignitionengine very well [18, 19].

These facts suggest a necessity for updated predic-tions of thermal conditions, in particular in-cylindertemperatures and temperature profiles in the engineblock and cylinder head [20, 21]. The potential ben-efits obtainable from better thermal predictions in-clude enhanced cooling systems (smaller and lighterpumps and heat exchangers) [22–25], reduced ther-mal distortion (lower friction and optimized piston-ring assembly), and higher power [26]. Improved pre-dictions of engine temperatures are also importantfor the application of computational methods (CFDand FEA) [23–27]. The accuracy of computer pro-grams relies, along with the availability of computerpower, on the specification of boundary conditions.Therefore, the application of these promising tech-niques will benefit from improvements in the methodsto predict thermal boundary conditions. Ultimately,these benefits are likely to contribute to decreasedfuel consumption and lower emissions in motor vehi-cles.

In this paper thermal analysis of cylinder head andblock of a newly developed turbocharged bi-fuel en-gine is presented because it is important to knowpeak temperature location and magnitude in the com-ponents. In addition, temperature gradients in solidshould be known to avoid intolerable thermal stress.Temperature field in both CNG and gasoline modes incylinder head and block are drawn and compared an-alytically. It is one of the most important steps towardreliable material selection by considering thermome-chanical limits; it also, seems essential to (re)designengine cooling system. Moreover, heat transfer pro-cesses are compared by considering temperature dis-tribution in an engine which is working in two modes(CNG and gasoline) and all engine geometrical vari-ables are similar in both modes. The most significantdifference is type of fuel that results in dissimilarity ofcombustion phenomenon in two modes. Because en-gine geometry, components material, cooling and lu-brication circuits design are same, effects of the men-tioned parameters in the analysis are not considered.Understanding the process in both modes and prepar-ing a base to compare them is helpful to design such anengine more reliable and efficient. Although completeanalysis needs advanced measurements and calcula-tions, temperature field analysis in the engine compo-nents will be very useful for engine material selectionand cooling passages design. From this approach, lo-cation and condition of peak temperature in solid aremeasured and discussed as well as temperature gradi-ents. This also, will applicable to develop, calibrate,and validate heat transfer and thermal stress calcula-tion codes. As well, knowing the quality and quan-tity of material thermal loads in the turbocharged bi-fuel engine is necessary to determine critical pointsfor defining engine durability and functional test meth-ods.

2 Thermal survey of engine

2.1 Turbocharged bi-fuel engine

In this research temperature distribution in cylinderhead and block of a recently developed bi-fuel (CNGand gasoline) turbocharged engine has been measured.Engine has been designed and developed by Irankho-dro Powertrain Co. (IPCO) and FEV Motorentechnikin turbocharged version. This engine is designed based


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Fig. 1 Engine characteristics; WOT, TC = 90 [°C]

Table 1 Bi-fuel turbocharged engine specifications

Bore × stroke [mm] 78.6 × 85

Displacement [l] 1.7

Compression ratio 10

Fuel injection system MPFI (CNG and gasoline)

Number of cylinder 4

Valve train system DOHC

Number of valve 16

Cylinder block material Cast iron

Cylinder head material Aluminum alloy

on dedicated CNG fuel concepts. In other words, en-gine features like compression ratio, cooling and lubri-cation circuits, and material selection are based on theassumption in which CNG is the main fuel and engineshould be durable and reliable while working perma-nently in natural gas mode. Specification of the men-tioned engine is given in Table 1 and engine torque andpower are graphed in Fig. 1. It shows engine perfor-mance in both modes are same after 2300 rpm wherethe boost produced by turbocharged are enough to re-cover lose of power in CNG mode.

2.2 Engine preparation

A bi-fuel TC engine, with specifications given in Ta-ble 1 and Fig. 2, was prepared to carry out thermalsurvey test in which solid temperatures of all engineoperating points were measured.

2.2.1 Instrumentation and test setup

24 sets of thermocouple were installed on cylinderblock and 18 sets on cylinder head. Positions of thetemperature sensors were chosen based on the experi-ence in a manner that almost all of critical points ofcombustion chamber and cylinder liner were covered.Figures 2 and 3 schematically show sensors’ positionson cylinder block and Fig. 4 shows location of ther-mocouples in cylinder head from two different views.To carry out a true measurement, sensors should bein direct contact with solid material and installed asclose as possible to the free surfaces which combus-tion gases touch. Maximum distance between sensors’tips from fire-side wall was 1 [mm]. Moreover, sen-sors’ tips should be insulated from engine coolant and


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Fig. 2 Positions of thermocouples on cylinder block in (a) crossdirection; (b) longitudinal direction

Fig. 3 Positions of thermocouples on cylinder block, top view

oil to avoid error in reading solid material tempera-ture.

24 sets of thermocouples were installed on engineblock in the following positions:

• 1st piston ring position in TDC at thrust and anti-thrust sides (cross direction)

• 1st piston ring position in TDC at flywheel and free-end sides (longitudinal direction)

• 1/4 and 1/2 stroke positions at thrust and anti-thrustsides (cross direction)

• 1/4 and 1/2 stroke positions at flywheel and free-end sides (longitudinal direction)

Figure 5 demonstrates positions where tempera-ture sensors were installed in the cylinder head ma-terial from a closer view as the sensors’ tips are just1 mm beneath the combustion chamber wall. Numberswhich are indicated in the figure represent the follow-ing locations:

• Between intake and exhaust valves (#1 and 2)• Between intake valves (#7 and 8)• Between exhaust valves (#3, 4 and 5)• Near spark plug (#6)

Making a good contact between sensors’ tips andsolid, a special glue was used which had very highconduction coefficient and is resistant to high temper-atures and vibration.

2.3 Testing methodology

Prepared TC engine was installed on a dynamometertest bench equipped with various devices and sensorsto measure temperature, pressure and flow rate of en-gine coolant, fuel (CNG and gasoline), inlet air, lubri-cating oil, and blow-by gases. Solid temperatures wererecorded in idle, and all part and full load points inboth CNG and gasoline modes. Engine coolant tem-perature (at engine outlet), air pressure and tempera-ture (at engine inlet), and exhaust back pressure werecontrolled by external devices to make comparableboundary conditions for various tests. Engine speedsare swept from idle to maximum and all the mea-surements carried out in part load and full load forboth fuel modes and two coolant temperatures (TC ):90 °C and 110 °C. Specifications of the CNG andgasoline which are used in the tests are listed in Ta-ble 2.

3 Test results

3.1 Cylinder block

The highest temperatures in cylinder block mate-rial were achieved in the vicinity of the rated powerpoint (5500 rpm, full load) in both gasoline and CNGmodes. At each operation condition, sensors installedin TDC position recorded maximum temperature in


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Fig. 4 Location of thermocouples in cylinder head from three different views

Fig. 5 Measuring points in cylinder head material

comparison with other positions. Figure 6 shows mea-sured temperatures in the cylinder block in gasolinemode, and Fig. 7 shows them in CNG mode. In thecase of TC = 90 °C, temperatures which were higher

Table 2 Specification of CNG and gasoline fuel used in thetests

Fuel type Gasoline

LHV 42.5 [MJ/kg]

Molecular weight 99.2 [g/mol]

Heat of vaporization 325 [kJ/kg]

(A/F)S 14.3

Fuel type Natural gas

LHV 46.3 [MJ/kg]

Molecular weight 17.9 [g/mol]

Heat of vaporization -

(A/F)S 15.8

than 280 °C are listed in Tables 3 and 4 in gasoline andCNG modes correspondingly.

Maximum temperature difference (between CNGand gasoline mode) was 25 °C at cylinder #2, TDC,free-end side, 5000 rpm. Generally, in both modes


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Fig. 6 Cylinder block temperature values: WOT, gasoline mode, TC = 90 [°C]

Table 3 Maximum temperatures of cylinder block in gasolinemode, WOT, TC = 90 [°C]

Position Speed[rpm]

Temperature[°C]

Cyl. #2, TDC, free-end side 5500 293.5


Cyl. #2, TDC, flywheel side 5500 284.8



temperature increasing rate was greater than zero inall points except 6000 rpm. In comparison with gaso-line, block temperature in engine speeds greater than2300 rpm were higher in CNG mode only at TDCpositions, while were lower at other speeds and po-sitions. Performing a direct comparison between thetwo modes, Figs. 8 and 9 show the most significantdifferences between CNG and gasoline temperaturesat TDC position of cylinder #2 and #4 respectively atcoolant temperature of 90 °C.

Table 4 Maximum temperatures of cylinder block in CNGmode, WOT, TC = 90 [°C]

Position Speed[rpm]

Temperature[°C]










3.2 Cylinder head

Highest temperatures in cylinder head material mea-sured at point #4 (between exhaust valves) aroundrated power (5500 rpm, full load) in both gasolineand CNG modes. Figures 10 and 11 show material


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Fig. 7 Cylinder block temperature values: WOT, CNG mode, TC = 90 [°C]

temperatures of the cylinder head in gasoline andCNG mode respectively at TC = 90 °C. Table 5 liststop five temperatures of cylinder head in each fuelmode. At the rated power condition where the cylin-der head material experiences maximum temperature,sensor #4 of cylinder #4 recorded 15.5 °C higher inCNG mode compared to gasoline. The highest tem-perature difference between CNG and gasoline modetook place at sensor #7 and #8 (between intake valves)which its value was around 30 [°C]. Trends of tem-perature differences and increase rate are similar tothose mentioned in the paragraphs in the previous sec-tion.

Regarding to the high number of graphs and lesscritical situation, part load results are not presented.

4 Thermal analysis

To reduce complexity, at first temperature fields incylinder block and head in each fuel mode are ana-

Table 5 Top five temperatures of cylinder head, WOT, TC = 90[°C]

Position Speed[rpm]

Temperature[°C]

Operationmode

Cyl. #4, sensor #4 5500 222.5 Gasoline

Cyl. #2, sensor #4 5500 219 Gasoline




Cyl. #4, sensor #4 5500 238 CNG

Cyl. #2, sensor #4 5500 237 CNG

Cyl. #2, sensor #4 6000 234.2 CNG

Cyl. #2, sensor #4 5000 229.1 CNG

Cyl. #4, sensor #4 5000 228.8 CNG

lyzed and then a comparison has been made betweenmodes. Before proceeding to the analyses, some en-gine parameters have been graphed: Fig. 12 showsspark timing and engine lambda variation measured


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Fig. 8 Temperature differences between CNG and gasoline mode at TDC position of cylinder #2 at coolant temperature of 90 °C

before exhaust catalyst converter, Fig. 13 displays to-tal energy flow rate (total energy flow rate equals tofuel mass flow rate × lower heating value) to engineand overall engine efficiency (overall engine efficiencyequals to brake power/energy flow rate), and Fig. 14show coolant temperature difference over engine inletand outlet.

4.1 Gasoline mode

Peak temperature in almost all positions occurs atrated power point (5500 rpm, WOT). Sensors installedbetween cylinders (free-end and flywheel sides ofcylinder #2, and free-end side of cylinder #4), wherethe coolant flow is almost nothing near TDC and islower at 1/4 and 1/2 stroke in comparison with othersides (intake and exhaust sides of cylinder #2 and in-take, exhaust, and flywheel sides of cylinder #4) indi-cate higher temperatures. To clarify this issue, CFDcoolant flow analysis has been done and the resultof cylinder block is depicted in Fig. 15. It should be

pointed out that the coolant enters the engine from thefree-end side and leaves from the flywheel side. Theresult shows higher coolant flow rate around the fly-wheel side of cylinder #4 compare to that of free-endside. In other words, free-end side of cylinder #4 hashigher temperature in comparison with its flywheelside. Generally, sensors at TDC show higher tempera-tures than sensors at 1/4 and 1/2 stroke at the sameside, mainly due to position of highest temperatureattained within the cylinder which typically is about15° aTDC for SI engines [1]. Exhaust Valve Bridge(sensors #4 and 5) is the hottest point of the cylinderhead in all operation points because of hot gas flowthrough exhaust ports. Like the block, cylinder headsensors show peak temperature at engine rated powerpoint.

As engine speed and load increases, total energyflow to engine increases and total efficiency decreases(Fig. 13). Decreasing in overall efficiency occurs be-cause a considerable fraction of additional gasolineis consumed to lower the exhaust gas temperature.


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Fig. 9 Temperature differences between CNG and gasoline mode at TDC position of cylinder #4 at coolant temperature of 90 °C

Preparing and inducing a richer mixture is a knownmethod to limit exhaust gas temperature and con-sequently to avoid damaging exhaust components.Therefore, as Fig. 12 shows, lambda decreases asspeed and load increase. Simulation investigationsshow as engine speed and load increase, the averageheat transfer per unit time increases and heat transferper cycle is a maximum at λ = 1.0 and decreases forricher and leaner mixtures [1]. From the other side theburned gas temperatures are decreased as spark tim-ing is retarded because combustion occurs later whenthe cylinder volume is larger [1]. Moreover, exhaustvalve temperature which is mainly affected by ex-haust gas temperature increases as spark timing is re-tarded.

Total effect of these engine variables results thesolid temperature in cylinder head and block increasewhen speed rises from 1000 up to 5500 rpm andthen drop partly at 6000 rpm. Studying the recordedtemperatures in gasoline mode reveals temperaturechange rate declines from a maximum value about 20

[°C/1000 rpm] at the lowest engine speed to about −5[°C/1000 rpm] at the highest speed. It demonstratesthe significance of lambda effect on thermal load ofdifferent components.

Increasing coolant temperature rises the tempera-ture of components which cooled by the coolant di-rectly [1]. In the case of TC = 110 °C engine vol-umetric efficiency and consequently engine power isreduced before 2300 rpm where the produced boostby turbocharger is negligible. Changing TC from 90to 110 °C in gasoline mode loses 3.5 % and 4.3 %of engine power at 1000 and 2000 rpm respectively.From the other point of view, the warmer cylinder headand liner wall results in preparing a better air-fuel mix-ture and combustion propagation with higher fuel con-version efficiency. Besides, oil temperature variationwhich in the case of self-setting is dependent uponcoolant temperature affects friction level of compo-nents.

Combination of these effects causes variation in thesolid temperature difference of a typical position due


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Fig. 10 Cylinder head temperature values: WOT, gasoline mode, TC = 90 [°C]

Fig. 11 Cylinder head temperature values: WOT, CNG mode, TC = 90 [°C]

to the engine speed variation. Nevertheless, coolant

temperature difference is constant. Figures 16 and 17

show solid temperature differences at TC = 110 °C

from that of TC = 90 °C in cylinder block and head

respectively. It is reduced as engine speed increases

mainly because at the higher engine speeds there is


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Fig. 12 Spark advance and engine lambda variation, WOT, TC = 90 [°C]

Fig. 13 Total energy flow rate to engine (fuel mass flow rate × QLHV) and overall engine efficiency (brake power/energy flow rate),WOT, TC = 90 [°C]

a lesser time for heat transfer. By virtue of chang-

ing coolant temperature, fuel conversion efficiency

and heat transfer will be different and maximum en-

gine power happens at 5000 rpm instead of 5500 rpm.Hence a peak in solid temperature difference graph isexpected.


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Fig. 14 Coolant temperature difference over the engine, WOT,TC = 90 [°C]

4.2 CNG mode

Although the influence of some engine variables de-pends upon the fuel type but overall trends are similar.Effect of total energy flow, overall efficiency, sparktiming, and lambda is same in fundamental but notin magnitude. The most important differences will bediscussed in the next section.

4.3 Comparison between CNG and gasoline mode

In order to compare temperature fields in CNG andgasoline two ranges have been recognized:

1. Engine speeds lower than 2300 rpm in which theturbocharger does not produce considerable boost.

2. Engine speeds equal and greater than 2300 rpm.

In these regions temperature field is influenced byfollowing parameters:

– Spark advance that is so advanced in CNG mode– Lambda which equals to the actual air fuel ratio di-

vided by stoichiometric air fuel ratio– Overall engine efficiency that is lower in gasoline

mode compare to CNG mode– Energy flow rate that is higher in gasoline mode

with respect to CNG mode at all speeds– Coolant temperature differences over the engine

that is lower in CNG mode at engine speeds lowerthan 2300 rpm and is indistinguishable from gaso-line mode at engine speeds equal and greater than2300 rpm

– Vaporization heat of gasoline which makes air-fuelmixture chilled

– In-cylinder pressure peak and its distribution

In the first division (engine speeds lower than2300 rpm) all sensors installed at TDC have recordedhigher temperatures in gasoline mode. It is due to en-ergy flow rate and efficiency of conversion which playsdominant role in these speed ranges. Figure 1 showsthat before 2300 rpm, where the turbocharger doesnot produce considerable boost, engine brake powerin CNG mode is 39 % and 28 % lower than those

Fig. 15 CFD coolant flow analysis, coolant flow velocities in cylinder block


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Fig. 16 Solid temperature difference at TDC, Cyl #4, due to coolant temperature difference (TSolid @ 110 − TSolid @ 90 [°C]coolant temperature), WOT

Fig. 17 Solid temperature difference at cylinder head Cyl. # 4, due to coolant temperature difference (TSolid @ 110 − TSolid @ 90[°C] coolant temperature), WOT

of gasoline mode at 1000 and 2000 rpm respectively.Also, Fig. 13 shows that energy flow rate is 41 % and

47 % lower in CNG mode at the same points. Fur-thermore, as shown in Fig. 14 coolant temperature dif-


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ference over the engine is highest at those speeds ingasoline mode and lowest in CNG mode. Therefore,more average heat flux and higher solid temperature atblock and head is expected in gasoline mode.

In the second section, combination of some ofmentioned parameters leads to higher temperatures inCNG mode in comparison with gasoline mode. Beforediscussing about reasons, it should be pointed out thatthere are four main ways for consumption of fuel en-ergy entering the combustion chamber:

1. Brake power2. Heat transfer rate to the cooling medium3. The exhaust enthalpy loss due to incomplete com-

bustion4. The exhaust enthalpy loss due to exhaust gases

With this very brief look at the consumption waysof fuel energy, we could turn to the reasons now.

– As Fig. 12 shows spark timing of engine in CNGmode is always about 20 degree more advancedthan that of gasoline mode, which leads a higherpeak pressure and consequently higher temperaturein CNG mode. Hence, there will be enough time forpremixed phase combustion in CNG mode which isthe most important reason for lesser unburned hy-drocarbon and higher overall engine efficiency inCNG mode compare to gasoline.

– Figure 14 shows that for the speeds greater than2300 rpm, the heat rejection is same in both modes.

– Regarding to lambda strategies in gasoline mode, asudden decrease in lambda occurs at 2300 rpm andas engine speed increases, λ decreases while it isalmost constant in CNG mode (Fig. 12). In spiteof this temperature control strategy, the temperatureof exhaust gases is higher in gasoline mode whichindicated about more energy losses. This issue hasbeen showed in Table 6.

– Figure 13 shows overall engine efficiency is higherin CNG mode in contrast to gasoline mode. It meansthat although the energy flow rate in gasoline modeis higher than that of CNG but the larger part ofCNG fuel converts to break power and the energyloss will be lower.

– Latent heat of gasoline vaporization is another rea-son. Making an estimation of its effect, a typicalsimple calculation is presented based on the fuelspecifications (see Table 2) and test bed data at5500 rpm, full load condition (see nomenclature forabbreviations):

Table 6 Temperature of exhaust gases, WOT, TC = 90 [°C]

Speed [rpm] Gasoline CNG

1000 617 491

2000 838 673

3000 886 820

4000 922 886

5000 937 941

5500 965 960

6000 953 949

HV = 3.25 [kJ/kg], (A/F)S(G) = 14.3

λ(G)@5500 [rpm] = 0.755

mf (G) = 38.92 [kg/h]mAir = 425.4 [kg/h]�TAir = mf (G)HV

mAirCP(Air)= 29.7 [°C]

Note that the calculation is based on air and gasolineproperties at 25 °C and the assumption that all injectedgasoline would be vaporized. It’s also worth mention-ing that the maximum temperature difference in cylin-der head is observed at intake Valve Bridge (sensors#7 and 8) because gasoline is injected in the port nextto the intake valves and heat is transferred from valvesand solids nearby.

Since the engine designed to produces same powerand torque for each speed and all engine geometry,components material, cooling and lubrication circuitsdesign are same, the combination of above reasonsleads to a higher temperature in CNG mode for speedsgreater than 2300 rpm.

Temperature on 1/4 and 1/2 stroke mainly dependson pressure distribution inside cylinders. On the otherhand, as Fig. 12 shows spark timing of engine in CNGmode is always about 20 degree more advanced thanthat of gasoline mode, which leads peak pressure get-ting closer to TDC in CNG mode. Also, methane ismore resistant to knock; thus spark timing could ad-vance to get closer to MBT point. This is a main pa-rameter which affects solid temperature values in 1/4and 1/2 stroke positions. Peak pressure position ingasoline mode is closer to 1/4 and 1/2 stroke andtemperature in all operating points will be higher thanthat of CNG mode. Therefore, temperature gradient incylinder block is more severe in CNG mode.

As indicated in Figs. 16 and 17, effect of chang-ing coolant temperature is same in CNG and gasoline


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Fig. 18 Temperature distribution in cylinder head at coolant temperature of 90 °C and WOT in gasoline mode

modes both in quality and quantity. That is to say, ef-fect of coolant temperature change is independent offuel type.

4.4 Application of test results in CAE

Cylinder block and head temperatures should not ex-ceed certain limits to guarantee for durability andfunction. For this purpose, it’s necessary to investi-gate component surface temperatures under worst caseconditions and theoretical results (CAE) should be val-idated by experimental data. Figure 18 shows the tem-perature distribution in cylinder head at coolant tem-perature of 90 °C and WOT in gasoline mode. It canbe seen that CAE results have a good adoption withtest results in cylinder head. For example, CAE resultfor Cyl. #2, sensor #4 is about 2 % lower than test re-sult.

Figure 19 shows the temperature distribution incylinder block at coolant temperature of 90 °C andWOT in gasoline mode. The results show 70 °C dif-ference in temperature for Thermocouple at Cyl. #2,TDC, free-end side which was due to consideringzero thermal resistance for gasket. After correcting the

value of thermal resistance of gasket, the temperaturedifference changed to 3.5 °C.

5 Conclusion

A complete thermal survey carried out on a recentlydeveloped bi-fuel turbocharged engine as an importantstep towards design and material verification. Gen-erally, TDC positions in cylinder block and exhaustValves Bridge in cylinder head were hottest points atall operation points and modes. Sensors installed inTDC positions and cylinder head recorded about 9 %higher temperature in CNG mode which could be crit-ical for some materials and designs. Temperature gra-dient in CNG mode is more severe; hence, more stressand deformation is expected in that mode. This wouldbe resulted in more wear, blow-by gas flow and oilconsumption in cylinder block and less sealing be-tween block and head [28]. Also, cylinder head wouldplace in more critical situation due to low cycle fatigueand material creep [29, 30].

To design and develop a new bi-fuel engine withthe equal performance in CNG and gasoline modes, it


Meccanica (2012) 47:1987–2004 2003

Fig. 19 Temperature distribution in cylinder block at coolant temperature of 90 °C and WOT in gasoline mode

should be noted that engine components would attainhigher peak temperature and larger temperature gradi-ent in CNG mode. It would influence design featuresof cylinder block and head and gasket.

Although the distribution of temperatures in thecylinder head is highly tri-dimensional, the data ob-tained on this engine provided valuable informationsuitable to be used in the definition of boundary con-ditions for CFD and FEA models.

Acknowledgements Authors would like to thank Mr. FarshidJahangiri Moghadam and Mr. Majid Karbasi for their invaluableguidance.

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Experimental thermal analysis of cylinder block and head of a bi-fuel turbocharged engine

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