Fuel Processing Technology 113 (2013) 97–108

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Fuel Processing Technology

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Conversion of a heavy duty truck diesel engine with an innovative powerturbine connected to the crankshaft through a continuously variabletransmission to operate compression ignition dual fuel diesel-LPG

Alberto BorettiUniversity of Ballarat, PO Box 663, Ballarat, 3353 VIC, Australia

E-mail address: a_boretti@yahoo.com.

0378-3820/$ – see front matter © 2013 Elsevier B.V. Allhttp://dx.doi.org/10.1016/j.fuproc.2013.03.018

a b s t r a c t

a r t i c l e i n f o

Article history:Received 13 December 2011Received in revised form 5 December 2012Accepted 14 March 2013Available online 11 April 2013

Keywords:Diesel enginesLPGAlternative fuelsPower turbine

This paper considers the option to convert the diesel engine to LPG retaining the diesel-like behaviour. LPG isan alternative fuel with a better carbon to hydrogen ratio permitting reduced carbon dioxide emissions. Itflashes immediately to gaseous form even if injected in liquid state for a much cleaner combustion almostcancelling some of the emissions of the diesel. Within Australia, LPG permits a much better energy securityand the refueling network is widespread. In this paper, a latest heavy duty truck diesel engine with a novelpower turbine connected through a continuously variable transmission to the crank shaft and fitted with aby-pass is modified to accommodate a second direct injector for the LPG fuel and operate full load with 5%diesel and 95% LPG. Results of engine performance simulations with diesel and diesel-LPG are presented.The engine retains the diesel performances while permitting the advantages of LPG in terms of particulate,carbon dioxide emissions and diversification of fuel supplies and energy security.

© 2013 Elsevier B.V. All rights reserved.

1. Conversion of diesel truck engines to LPG

As long as the price of crude oil increases and the threat of globalwarming and air pollution remain, there will be a need for alternativefuels. Liquefied Petroleum Gas (LPG) has emerged as one of the solu-tions to these challenges because it is inexpensive, clean burning andrelatively low in carbon content. LPG has however also some issuesassociated with global LPG availability, distribution infrastructure,handling, fuel-system changes, lubricity, ignition quality, and compo-sitional variability. Diesel engines have been the backbone of thetransportation industry because of their power, fuel economy anddurability. However, diesel engines are heavy emitters of particulatematter (PM) and oxides of nitrogen (NOx). By converting diesel enginesto run on LPG we can significantly reduce the problem of diesel pollu-tion whilst also improving emissions of greenhouse gases (GHG). Suchconversions are however not a simple matter of changing the fuel,many technical problems present particularly with availability of spe-cific fuel injectors, fuel injection control and engine optimisation toensure that the engine performance is maintained and the exhaustemissions are minimized.

Road transport is particularly relevant to Australia [1]. All the state-ments below are taken from [1]. Upuntil the early 1960s, railways dom-inated all but the shortest land-based freight task in Australia. Sincethen, vast improvements in road vehicle productivity and road infra-structure quality, the gradual removal of regulations restricting roadfreight carriage and the exponential growth in interstate trade has

rights reserved.

broadened the range of freight tasks for which road is better suitedthan rail. The Australian domestic freight task measured 521 billiontons kilometers in 2007, with 35% carried by road. The domestic freighttask has doubled in size over the past 20 years, with an average growthof 3.5% per annum. Projections suggest this trendwill continue, althoughwith slightly slower growth into the future, growing by approximately3.0% per annum until 2030. Over this period, road freight volumes areprojected tomore than double, with domestic demand formanufacturedgoods sustaining much of the growth. Energy transport accounts for the14.5% of the national carbon dioxide equivalent annual emissions inAustralia.

Life-cycle emissions analysis of alternative fuels for heavy vehicleshas shown the potential of gaseous fuels for heavy duty vehicles inAustralia [2]. All the statements below are taken from [2]. Therehave been major advances in natural gas engines in recent yearsthat mean that the present generation of natural gas vehicles has sig-nificantly lower emissions than the present generation of diesel vehi-cles. The emissions based on use in Original Equipment Manufacturer(OEM) vehicles are lower in all categories, GHG, important criteriapollutants, and air toxics. The lower particulate emissions and noiselevels compared with diesel make it particularly attractive for urbanareas. The major uncertainty relates to upstream and in-service leak-age, on the other hand already sufficiently reduced in the presentgeneration of OEM compressed or liquid natural gas (CNG or LNG) vehi-cles, and the lack of sufficient refueling stations. The extra weight ofCNG fuel tanks leads to slightly higher fuel consumption, or loss ofpayload in the case of buses, but this is less of a problemwith LNG vehi-cles due to the higher energy density. Similarly, a dedicated LPG bus

98 A. Boretti / Fuel Processing Technology 113 (2013) 97–108

produces significantly lower emissions of important criteria pollutants,and lower exbodied emissions of GHG. Toxic air pollutants from tailpipeemissions of LPG vehicles are much lower than those of diesel vehicles,but the greater upstream emissions of air toxics results in the exbodiedemissions of air toxics from LPG being much the same as those fromdiesel. LPG HD5 has minimum propane content of 90% whereas theratio of propane and butane varies widely in auto gas LPG. When com-pared with auto gas, HD5 LPG emits more NOx but less particulatematter. Emissions of hydrocarbons are similar. The main benefit of HD5compared with auto gas is that the compression ratio can be altered tosuit this higher-octane fuel. Therefore, the lower particulate emissionsand lower noise levels compared with diesel make it attractive for usein urban areas. Themajor disadvantage of LPG is the lack of market pen-etration of LPG dedicated heavy vehicles.

In Australia LPG is used in automotive applications in cars (almostall taxis) and special applications in close proximity to workers inconfined or enclosed spaces such as fork lift trucks, concrete mixersand similar plant because of the inherent low emissions of the fuelrelated to its gas phase and small molecule size relative to conven-tional fuels. There have been few serious attempts to convert heavyduty diesel engines to LPG. In many countries there is poor availabil-ity of both LPG and CNG outlets, and in many others where LPG isused for heating and cooking there is an excess of demand over sup-ply. The engine manufacturers are therefore reluctant to devote effortto developing engines for a niche or non-existent market. Australia isunique in having an established LPG infrastructure with over 3500retail outlets for LPG distributed across the continent well in excessof those in the whole of North America, where they are largely con-fined to West and East Coast regions. Furthermore Australia producesabout 30%more LPG than it consumes and its availability will increaseas more LNG plants come on stream as LPG is a liquid by-productfrom natural gas liquefaction. Inherently as a by-product LPG hasonly a minor energy production cost compared with gasoline and die-sel where significant increase in production energy have occurred tomeet the recently introduced Australian fuel quality standards reduc-ing sulphur levels and meeting composition and test requirementswith strict quality control. Moreover the GHG emissions associatedwith LPG production according to both GREET and SIMA-PRO LCAanalyses are only about two thirds of those of low sulphur dieseland about 60% of gasoline production. More important is that LPGhas a 12% reduction in CO2 emission (because of its lower C/H ratio),and whilst not as low as NG (Methane), it requires negligible compres-sion energy comparedwith NG gas compression (~7–11% of gas energy)to CNG or liquefaction (~17%) to LNG.

The challenge for further use of LPG is to produce an engine thathas equal or near equal thermal efficiency to diesel to thus obtainthe benefits of fuel composition (12%) and (8–20%) less productionand distribution energy. Use of LPG in truck engines is covered in afew papers written over the last 60 years [3–13] that demonstratethe not-that-great interest paid so far by the automotive industry toLPG. These papers cover both applications diesel-like, gasoline-likeand mixed diesel/gasoline-like. The option considered in this paperis to use a dual fuel injector with a pilot diesel followed by a mainLPG as it has been done since the 1980's with diesel and NG[14–17]. Dedicated diesel and LPG dual fuel injectors are currentlyunavailable on the market, but considering their design followingthe Westport HPDI concept [16,17] should not be difficult, the paperassumes these injectors could be made available and investigatesthe performance a latest heavy duty diesel engine could have withthese injectors by using simulations [18].

In addition to this design, the Author is also considering the optionto convert a diesel engine to LPG replacing the diesel injector with aLPG injector and the glow plug with a jet ignition injector accommo-dating another LPG injector and a glow or a spark plug. Both solutionsneed parts not available off-the-shelf, being the dual fuel diesel-LPGand the LPG only direct injectors currently unavailable.

2. Dual fuel diesel–gas diesel-like combustion

The direct injection of gas and diesel within the cylinder is not anovelty. Since the early 1980's, several researches have been madeon the subject, with [14,15] just to name a few. A small amount of die-sel fuel was injected before a main injection of natural gas to start thecombustion of the engine. With this technology, diesel engines whereable to run on natural gas without compromises in performance andfuel efficiency [16,17]. The major advantage of the Direct Injectiondiesel engine is the bulk, lean, diffusion combustion in a high com-pression ratio highly boosted engine with reduced heat losses to theliner and head walls. The engines using the stoichiometric, premixedcombustion controlled by a spark discharge suffer the disadvantage ofthe higher heat losses to the liner and head walls. The combustion iswall initiated, and the flame has to reach all the walls for burning allthe homogeneously distributed fuel. Besides, the compression ratio isreduced to avoid the abnormal combustion phenomena. In the leanburning diesel the load may be controlled by the quantity of fuelinjected, while the stoichiometric gasoline requires throttling theintake to vary the load and this bring a further advantage of the dieselable to perform with smaller penalties from already higher efficien-cies reducing the load. In case of dual fuel operation of a diesel enginewith LPG as the second fuel, it seems therefore logical to inject boththe diesel and the LPG fuel within the cylinder to achieve the sameas diesel efficiencies both full load and part load, with the loadbeing controlled by the quantity of fuel injected.

The dual fuel injector [16,17] has the advantage to be easily incor-porated into an existing diesel engine with minimal or no modifica-tions to the engine cylinder head. A two injectors' option wouldrequire a new design of the cylinder head however possible onnovel engines. Late-cycle, high-pressure direct injection may ensurediffusion type combustion for the LPG and therefore retains the highpower, torque, and efficiency of the diesel engine with possibly theadvantages of higher power densities running much closer to stoichi-ometry. Not considered here, more flexible injection strategies mayalso permit different combustion strategies for operation not onlydiesel-like but also gasoline-like or homogeneous charge compres-sion ignition (HCCI)-like. Homogeneous charge compression ignitionmay be achieved directly injecting the LPG early during compression.Diesel-like operation may be achieved preceding a late LPG main in-jection with a diesel pre injection. Gasoline-like operation may beachieved injecting the diesel after the LPG. Mixedmodes may be finallyobtained performing simultaneously the two injections. Furthermore,continuous operation of the glow plug could possibly permit the autoignition of the LPG without the diesel injection. All these additionalopportunities are neglected in the present work but certainly deservingfurther investigations. The injector has to manage the high volumetricflow rates required for the LPG fuel much less dense than diesel andthe carbonizing and thermal fluctuations that can result from thereduced cooling capacity of the LPG. The major disadvantage of theLPG–diesel fuel injector is the complex fuel routing for two fuels withinone injector. The injector has to provide a small diesel pilot spray and amuch larger LPG gas spray. The diesel should averages only about 5% ofthe total energy input and it is used to start themain combustion of thehigh-pressure directly-injected LPG gas.

To model the operation of the engine with the dual fuel injector,the model of the diesel only operation has to be discussed first. Theactual fuel injection strategy is quite complicated and mane of multi-ple injections. Typically, at low loads the fuel injection profile is madeof a pilot, a pre and a main injection. At medium to low loads, an afterinjection is added following the main injection. At medium to highloads, the after injection is removed. At high loads, there is only apre and a main injection. The reasons why these injection profilesare adopted are not only because of fuel economy, but also becauseof other optimisation criteria including formation of pollutants andnoise. These complex injection profiles produce a subsequent complex

99A. Boretti / Fuel Processing Technology 113 (2013) 97–108

combustion evolution. In the following section, these complex injectionprofiles and the subsequent complex combustion evolution is approxi-mated by a single equivalent injection and a much simpler combustionprocess. Today's engine performance models like GT-POWER [18] orWAVE [19] are not able to represent these complicated injection andcombustion phenomena, but they can tackle reasonably well the singleinjection or the pilot and main combustion events producing the heatrelease rates shown in Fig. 1. For the purpose of fuel conversion effi-ciency computations only, the more complex injections may then bereplaced by the much simpler equivalent pilot and main injectiononly, or even by a single equivalent injection event. Obviously, manyof the engine optimisation criteria, namely soot, smoke, NOx, HC andCO emissions are this way not predictable with accuracy or are notpredictable at all. This is demonstrated by the extensive validation ex-ercise proposed in Appendix A for a 1.6 liter passenger car TDI engine.

Comparison of computed and measured values of full loadpower output and fuel conversion efficiency and brake specific fuelconsumptions (BSFC) at different engine speeds and loads usually

Fig. 1. Modeled heat release rate profile vs. crank angle

provides an accuracy within a few percents. Appendix A presentsthe experimental and the computational results obtained on a partic-ular engine. The experimental results considered in the validation ex-ercise include power, mass of fuel injected in all the injection eventsand timings and durations of these events, common rail pressure,fuel flow rate, air flow rate, boost pressure, maximum pressure withinthe cylinder, pumping mean effective pressure, temperature andpressure in selected location within the gas exchange line includingentry and exit of turbine and compressor, exhaust gas recirculationstroke, EGR temperature, speed of turbine and compressor, brakespecific fuel consumption, smoke, combustion noise, brake specificemissions of CO, soot, NOx and HC, fuel-to-air and air-to-fuel ratio,pressure and temperature of water and oil for 180 different mappoints. The pictures in Appendix A present some interesting compar-isons both full load and part load of some of these parameters. Thefuel conversion efficiency (i.e. the BSFC) for a given BMEP and speedis well predicted with the equivalent single injection and the simpli-fied combustion model all over the speed and load range.

3. Details of the modelling of injection and combustion

Results in Appendix A prove that despite the very simplified representation of the complex multi event injection, an equivalent single injec-tion of all the fuel and a semi predictive model may approximate the combustion evolution quite well for what concerns the pressure build-upand the heat transfer to the cylinder walls. The differences in measured and computed brake specific fuel consumptions BSFC (or brake enginethermal efficiencies η = CF/BSFC being CF a fuel related constant) are generally a few percents over the load range all over the range of enginespeeds, with differences increasing reducing the loads to show a less accurate modelling of combustion with higher air-to-fuel ratios as well asthe less accurate modelling of the gas exchange processes far from design points and ultimately the increased relative weight of frictionmean effec-tive pressure from a simple correlation vs. the computed indicated mean effective pressure. Worth of note is that the start and the duration of theequivalent injection are not related to actual multi event injection parameters. Once a combustion model has been defined, some tuning is neededin principle to determine the start and duration of the equivalent injection. To get the results of Appendix A, this tuning has been very limited, withboth parametersmade linearly proportional to load and speed. Result of this tuning is an equivalent injection event close to the actualmain injectionevent but with a totalmass injected inclusive of the other injections contributions. FromAppendix Awemay therefore conclude that for the purposeof global heat release and pressure build-up computation only, an equivalent single injection of the same amount of diesel fuel injected in themultiple events may therefore replace a complicated injection strategy made up of pre, pilot, main and post injections as the one shown in Fig. 1.

Considering the diesel-like behaviour of dual fuel LNG-diesel engines having a pilot diesel injection followed by a main LNG injection exper-imentally verified in many circulating heavy duty trucks converted from diesel [14–17], it seems therefore reasonable to assume that for the samepurposes, an equivalent single injection of LPG fuel may replace a complicated injection strategy made up of a pilot diesel injection and a main LPGinjection. As previously stated, the Westport HPDI injector [14–17] has been developed to operate dual fuel with diesel and NG, being NG a fuel ofmuchmore interest than LPG in the North Americanmarket. There is no special reason why a similar injector could not be developed for diesel andLPG, and there is no special reason why a modified diesel engine should work with diesel pilot and LPG main in a drastically different way from theengines working with diesel pilot and LNG main. Obviously, some refinement of the injector, for example the size and the number of the injectionholes should be made different to better work with a different fuel. Further experiments should certainly be conducted to give improved insightsinto the effects of fuelling with LPG rather than LNG using the dual fuel direct injection technology developed for LNG. However, the properties ofLPG are not that far from those of LNG [20] to make unreliable the assumption. Propane has a research octane number of 110 smaller than the re-search octane number of NG of 130. The Cetane number is therefore slightly higher for propane than NG even if obviously much smaller than theone of the diesel. The auto ignition temperature of propane is 723 K vs. the 755–905 K of NG while the diesel has an auto ignition temperatureabout 590 K. The upper and lower flammability limits in volume are 2.2 and 9.5% for propane and 5.3 and 15% for NG while they are 1 and 6 forthe diesel. The other fluid properties for methane and propane fluids at 298 K and 200 bar (from [21]) presented in Table 1 do not suggest thepossibility of major issues with the fuel replacement.

with diesel pilot and main diesel or gas injections.

Table 1Properties of propane and methane (from [21]) at 200 bar and 298 K.

Fluid Propane Methane

Critical temperature (Tc) 369.825 K 190.564 KCritical pressure (Pc) 42.4766 bar 45.992 barCritical density (Dc) 220 kg/m3 162.66 kg/m3

Acentric factor 0.1524 0.01142Normal boiling point 231.06 K 111.667 KDipole moment 0.084 Debye 0.0 DebyeDensity (kg/m3) 532.85 157.24Volume (m3/kg) 0.001877 0.00636Internal energy (kJ/kg) 236.8 601.35Enthalpy (kJ/kg) 274.34 728.54Entropy (J/g · K) 1.1315 3.473Cv (J/g · K) 1.6793 1.8558Cp (J/g · K) 2.4549 3.5717Sound spd. (m/s) 975.38 525.01Joule–Thomson (K/bar) −0.03089 0.15838Viscosity (Pa · s) 0.000129 1.94E−05Therm. cond. (W/m · K) 0.11066 0.061976Phase Liquid Supercritical

100 A. Boretti / Fuel Processing Technology 113 (2013) 97–108

Starting from measured pressure traces for the different map points with the diesel, the heat release analysis may provide the combustionmodel parameters of the different map points. Then, after some tuning on the start and duration of the equivalent injection, the enginemodel may then provide maps of brake specific fuel consumption vs. BMEP and speed very close to the experimental ones. In what follows,the assumption made is that with diesel only and diesel pilot and LPG main, the only thing that changes in between the two models are thefuel properties, but the combustion model parameters and the start and duration of the injection set up for the diesel are also valid for the dieselpilot and the LPG main at same speed and load. This way the operation of a converted heavy duty truck diesel engine to run diesel and LPG withan injector to be specifically developed following the Westport HPDI concept can be investigated. This is done in the next section.

The main injection combustion is modelled with the GT-POWER DI Wiebe model [18]. This model imposes the burn rate using a three-termWiebe function (the superposition of three normalWiebe curves). TheseWiebe curves approximate the “typical” shape of a DI compression ignitionburn rate. The purpose of using three functions is to make it possible tomodel pre-ignition (the large initial spike) and larger tail. The injection profiledoes not influence the burn rate except if, at any instant, the specified cumulative combustion exceeds the specified injected fuel fraction. The inputsof the Wiebe equations are SOI =Start of Injection, ID = Ignition Delay, DP = Premix Duration, DM =Main Duration, DT = Tail Duration, FP =Premix Fraction, FT = Tail Fraction, EP = Premix Exponent, EM = Main Exponent, ET = Tail Exponent and CE = Fraction of Fuel Burned (alsoknown as “Combustion Efficiency”). The calculated constants of these equations are FM = Main Fraction, WCP = Wiebe Premix Constant, WCM =Wiebe Main Constant andWCT = Wiebe Tail Constant given by equation below.

FM ¼ 1−FP−FTð Þ WCP ¼ DP

2:3021= EPþ1ð Þ

−0:1051= EPþ1ð Þ

264

375− EPþ1ð Þ

WCM ¼ DM

2:3021= EMþ1ð Þ

−0:1051= EMþ1ð Þ

264

375− EMþ1ð Þ

WCT ¼ DT

2:3021= ETþ1ð Þ

−0:1051= ETþ1ð Þ

264

375− ETþ1ð Þ

With θ = Instantaneous Crank Angle, the burn rate is finally given by the equation below:

Combustion θð Þ ¼ CEð Þ FPð Þ 1−e− WCPð Þ θ−SOI−IDð Þ EPþ1ð Þ� �

þ CEð Þ FMð Þ 1−e− WCMð Þ θ−SOI−IDð Þ EMþ1ð Þ� �

þ CEð Þ FTð Þ 1−e− WCTð Þ θ−SOI−IDð Þ ETþ1ð Þ� �

The cumulative burn rate is calculated, normalised to 1.0. Combustion starts at 0.0 (0.0% burned) and progresses to 1.0 (100% burned).The combustion model uses a (diesel) Wiebe function of parameters tabled vs. load and speed for the diesel only operation. The diesel

parameters are used for the diesel and LPG operation same percentage of maximum load and speed.It is worth of note that diesel-like operation of engines fueled with gaseous fuels has been achieved so far not only with the dual fuel option

diesel-LNG of Westport [14–17] but also with the injection of only the gaseous fuel [16,22–24] having some aids to the start of combustion like acontinuously operated glow plug [16,22–24].

The assumption of start and duration of injection as being the same at the same speed and load of the diesel and the dual fuel injector is notsupported by measurements done on off-the-shelf injectors, being the dual fuel diesel-LPG injector a novel injector to be developed. Possiblyonly changing the percentage of diesel fuel injected during the pilot injection vs. the diesel-LNG operation, there is no doubt the diesel-LPGengine could behave similarly to what has been verified with LNG, CNG and hydrogen.

4. Engine results and discussion

Engine performance simulations have been performed with theGT-SUITE code [18] for a 12.8 litre in-line six cylinder turbo chargeddirectly injected diesel engine. The engine geometry (baseline) is

presented in Table 2. IVO is the intake valve opening, IVC the intakevalve closure, EVO the exhaust valve opening and EVC the exhaustvalve closure. IVO of 314° means the intake valves open 46° beforeTDC. IVC of 602° means the intake valves close 62° after BDC. EVOof 100° means the exhaust valves open 80° before BDC. EVC of 400°

Table 212.8 litre in-line six cylinder turbo charged directly injected diesel engine.

Displacement per cylinder [l] 2.13Number of cylinders 6Engine layout L-6Compression ratio 18Bore [mm] 131Stroke [mm] 158Connecting rod length [mm] 300Wrist pin offset [mm] 0Clearance vol. [l] 0.125Engine type C.I.No. of intake valve per cylinder 2Intake valve dia. [mm] 46.3Intake valve max. lift [mm] 12.1No. of exhaust valve per cylinder 2IVO [deg] 314(−46)IVC [deg] 602 (+62)Exhaust valve dia. [mm] 41.5Exhaust valve max. lift [mm] 13.4EVO [deg] 100 (−80)EVC [deg] 400 (+40)

101A. Boretti / Fuel Processing Technology 113 (2013) 97–108

means the exhaust valves close 40° after TDC (this set-up is due to thepower turbine). A power turbine downstream of the turbocharger tur-bine is used to recover a percentage of the thermal energy that wouldnormally be lost through the engine's exhaust. The first turbine in theexhaust gas path is the turbocharger turbine driving the turbochargercompressor. The second turbine in the exhaust gas path is the powerturbine, supplementing the engine power at the crankshaft through acontinuously variable transmission. With reference to a fixed gearratio transmission, the continuously variable transmission has the dis-advantage of a worse mechanical efficiency in the best point of opera-tion, but the advantage of much better fluid dynamic efficiencies apartfrom this point by using the optimal speed of rotation of the turbine forevery speed-load operating point, and the opportunity to even disen-gage the turbine from the crankshaft when needed.

Fig. 2 presents a sketch of the turbocharger, engine cylinders andpower turbine. The toroidal variator and the clutch permit to decou-ple the speed of the power turbine from the speed of the crankshaft,otherwise fixed with a simple couple of gears. The toroidal variatorand the clutch are the typical component of the continuous variabletransmission coupling the flywheel and the drivetrain in mechanicalkinetic energy recovery systems [26].

Engine simulations have then been performed with and withoutthe power turbine. A by-pass of the power turbine is also included.This permits to operate the power turbine when the turbine producesmore power than the power loss for the back pressures while permit-ting temperatures to the downstream after treatment system highenough. The continuously variable transmission permits operationof the power turbine at the more efficient speeds. The engine has atarget performance (with diesel fuel) of 2600 Nm torque 1000–

Fig. 2. Sketch of the turbocharger, engine cylinders and power turbine. The toroidalvariator and the clutch permit to decouple the speed of the power turbine from thespeed of the crankshaft, otherwise fixed with a simple couple of gears.

1450 rpm, and of about 400 kW of power 1450–1900 rpm, withBSFC values around 190 g/kWh, corresponding to brake efficienciesof 44%. (The Volvo D13 and D16 engines are now a clear benchmarkfor truck manufacturers, delivering up to 2600 Nm of torque and upto 400 kW of power the 13 l, and up to 3150 Nm of torque and515 kW of power the 16 l [25]). The (diesel) engine is compliantwith EURO 5 emission standards. The additional cooling of the ex-haust gases through the power turbine may in principle reduce theeffectiveness of the exhaust after treatment systems. In principle,this may require more active regenerations for particulate filter, orless use of cooled EGR. The additional cooling may certainly reducethe time when NOx systems are effective (LNA, SCR, or LNC). Howev-er, using the power turbine for high loads and speeds only, this is notan issue, cause temperatures for these loads and speeds will be other-wise much higher than those of medium to low loads and speeds.

The diesel fuel is considered a liquid hydrocarbon of compositionC = 13.5 and H = 23.6, fuel heating value at 298.15 K of 43.25 MJ/kgand stoichiometric air-to-fuel ratio of 14.33. The LPG fuel is consideredpropane liquid of composition C = 3 H = 8, fuel heating value at298.15 K of 46.35 and stoichiometric air-to-fuel ratio of 15.58. TheLPG is a much simpler liquid hydrocarbon that vaporises and mixeswith air faster than the diesel. As in the diesel engine model, also inthe LPG model only the main injection event is considered. In bothcases, the details of the previous pilot and pre injection with the dieselor the pilot diesel with the LPG are neglected. Then, we suppose thesame injection durations and the same combustion evolution willapply for the diesel and the LPG.

Fig. 3 presents the brake efficiency, mean effective pressure andtotal and turbine power for diesel-only and propane-diesel enginesat minimum λ of 1.45 for the diesel and 1.25 for the propane.Especially in engines with a power turbine downstream of the turbo-charger turbine as in this case, running lower λ with propane mayproduce benefits in both the power output and the fuel conversionefficiency. The power turbine works much better with propane anda lower λ. The top fuel conversion efficiency (total) is a 44.3% forthe propane engine and 43.8% for the diesel, but the total output is480 kW with propane and just 400 kW with the diesel.

Fig. 4 presents the brake efficiency and the air-to-fuel equivalentratio λ for the diesel and the propane engines at fixed value ofbrake mean effective pressure 1 to 5. The power turbine is nowby-passed and disconnected. The efficiency is about the same, whilethe equivalent air-to-fuel ratio is marginally lower with propane.

Fig. 5 presents the brake efficiency and the air-to-fuel equivalentratio λ for the diesel and the propane engines at fixed value of brakemean effective pressure 10 to 25. A minimum λ of 1.45 is consideredfor the diesel, while a smaller minimum λ of 1.25 is considered for thepropane. The diesel is a mixture of heavier liquid hydrocarbons, andvaporisation andmixing with air takes longer than when injecting pro-pane, that especially if the injection holes should bemade small enoughshould not take too much to vaporise in a high temperature and highpressure environment. The efficiency is higher with the propane,while the equivalent air-to-fuel ratio is marginally lower. The powerturbine operates all over these loads over the full speed range. Theoperation at 25 bar BMEP is permitted with propane (but with alower λ) and not permitted with diesel.

A minimum value of λ of 1.25 is used for the LPG engine, while aminimum value of 1.45 is used for the diesel engine. This compen-sates for the much faster mixing of the LPG flashing immediatelyafter being injected and the therefore much quicker mixing with air.It is at higher loads and speeds that the LPG fuel permits to improveall the performances in both fuel conversion efficiency and poweroutput also thanks to the power turbine. At low speeds and loads,the LPG would perform similarly to the diesel engine.

The power turbine does not help too much to increase the torqueor the power output except than at very high speeds. It is therefore ef-fective for this criterion only over a very narrow portion of the speed

Fig. 3. Brake efficiency, mean effective pressure and power for diesel-only andpropane-diesel engines at minimum λ of 1.45 for the diesel and 1.25 for the propane.

102 A. Boretti / Fuel Processing Technology 113 (2013) 97–108

range about maximum speed. This is because using the exhaust ener-gy in the turbine turbocharger permits to introduce much more airwithin the cylinder that translates in much more fuel within the

cylinder operating with a fixed air-to-fuel ratio. However, sometimesthe turbocharger turbine does not use the available energy about themaximum speed and maximum load operating point because of therequirements to properly cover all the speeds and loads range, andthere is therefore space for further improvements there. The powerturbine helps considerably in increasing the efficiency, with increasesof almost 4 percentage points in the brake efficiency. These figuresare slightly better than those already provided by other researchers.The reason is the operation of the power turbine in a better speedrange thanks to the continuously variable transmission.

Thanks to the continuously variable transmission and the by-pass,the power turbine is operated close to the area of high efficiencywhen the conversion of the energy available produces more benefitsthan the increased back pressure and the temperature downstreamof the power turbine remains high enough for the after treatment.When exhaust gases pass through the turbine, the pressure and tem-perature drops as energy is extracted and because of the inevitablelosses. The power taken from the exhaust gases with the power tur-bine is larger than without. To make this possible the pressure inthe exhaust manifold is higher and this increases the pump workthat the pistons have to do. The power turbine permits better efficien-cies at high loads and high speeds, when the energy available down-stream of the turbocharger turbine is significant, while at low tomedium speed and low to medium loads the effect is opposite.Decoupling the speed of rotation of the power turbine from thespeed of the crankshaft and by passing the power turbine when it isnot convenient for after treatment or fuel conversion efficiencies,turbo compounds may still be attractive. This option could maketurbo compounds appealing also for passenger cars application,where so far almost all of the major Original Equipment Manufac-turers (OEM) have tested this opportunity but decided not to movefurther. Additional benefits may arise from designing specific lowpressure ratio turbines that would maximize the trade-off in betweendrop of pressure and temperature within the turbine and workextracted.

The computational result is due to the same Wiebe function pa-rameters (i.e. combustion timings and duration) of smaller amountsof diesel and larger amounts of LPG vs. their stoichiometric values.This assumption is only preliminary and it is based on the literature'sWestport HPDI experience with Diesel and LNG plus the opportunityto tune the amount of LPG injected before and after the diesel injec-tion ignition event to change the premixed and diffusion combustioncomponents. The pressure build-up is therefore supposed to be muchcloser to isochoric in the LPG rather than the diesel operation, beingthe liquid propane a fuel much easier to vaporise and burn than thediesel made up of a mixture of higher alkanes. With a smaller λ, thetemperature increases within the cylinder, but in the bulk ratherthan close to the walls. This reduces the heat transfer penalties withinthe cylinder while keeping the advantages in terms of energy avail-able at the power turbine. Clearly this result must be confirmed atthe dynamometer.

With reference to the particulate matter (PM) emission, the com-putational procedure does not permit to determine this parameter forevery operating point. Experiments done on a truck engine convertedto LPG with upstream manifold injection of LPG, a turbulence plate tolower the compression ratio and enhance in-cylinder turbulence anda spark plug to control ignition have shown very low values of PM vs.the diesel [27]. It is expected that the direct injection of LPG will pro-vide marginally worse results than the upstream injection and thehomogeneous combustion.

5. Conclusions

Dual fuel LPG–diesel internal combustion enginesmay operatemoreefficiently and with increased power density adopting direct injectionwith high pressure and fast actuating injectors. The key development

5 bar BMEP

2.5 bar BMEP

1 bar BMEP

Fig. 4. Air-to-fuel equivalence ratio λ and brake efficiency for diesel-only and propane-diesel engines light loads with power turbine by-passed and disconnected.

103A. Boretti / Fuel Processing Technology 113 (2013) 97–108

for these engines is the novel injectors able to deal with LPG and dieselindependently with high flow rates and high speed actuation.

The major advantages of the dual fuel LPG–diesel operation arethe reduced CO2 emissions, for the better C\H ratio of the LPG fuel,the reduced smoke and particulate emissions thanks to the muchsimpler liquid hydrocarbon having a faster vaporisation and mixingand combustion, and finally the better energy security for Australiareplacing foreign diesel with local LPG.

The dual fuel LPG–diesel engine may have top brake engine thermalefficiency approaching 45% as in the original diesel-only engine andsimilarly reduced penalties in efficiency reducing the load. The dualfuel LPG–diesel engine may however permit up to 10% larger poweroutputs thanks to themuch closer to stoichiometric operationwith LPG.

Even if the engine may possibly also operate diesel only or dieselpilot and LPG main, the availability of a diffused refueling networkmay suggest adoption of a small diesel fuel tank and a large LPGfuel. Availability of both diesel and LPG directly injected independentlycertainly leaves theflexibility to exploremanymoremodes of operationthan the classic diesel-like.

LPG in Australia especially may consist of a wide range of (typically)propane/butane blends. Simulating LPG as propane is certainly ques-tionable for Australia. However, engine performance simulations areusually performedwith reference rather than actual fuels, and the valid-ity of the results is not infringed by the change of fuel maintaining thepump LPG vs. the pump diesel the most part of the advantages propanehas vs. the reference diesel.

25 bar BMEP

20 bar BMEP

15 bar BMEP

10 bar BMEP

Fig. 5. Air-to-fuel equivalence ratio λ and brake efficiency for diesel-only and propane-diesel engines medium to heavy loads with power turbine operational.

104 A. Boretti / Fuel Processing Technology 113 (2013) 97–108

105A. Boretti / Fuel Processing Technology 113 (2013) 97–108

The major downfall of alternative and renewable fuel engines ingeneral is the limited amount of moneys available to support theirspecific research and development. This is true especially for LPG,where products by OEM as well as by their suppliers are either nonavailable or still very far from their potentials. Without any support,this research has shown that diesel engines converted to rundiesel-LPG dual fuel with only a specially developed dual injectormay be of interest to Australia. To move further some support forexperiments and development of components is certainly needed.

NomenclatureBDC Bottom dead centreBMEP Brake mean effective pressureBSFC Brake specific fuel consumptionC CarbonCNG Compressed natural gasCVT Continuously variable transmissionDI Direct injectionEGR Exhaust gas recirculationEVO Exhaust valve openingEVC Exhaust valve closureGHG Green house gasesH HydrogenIVO Intake valve openingIVC Intake valve closureLHV Lower heating valueLNG Liquefied natural gasLPG Liquefied petroleum gasNG Natural gasNOx Oxides of nitrogenPM Particulate matterTC TurbochargedTDC Top dead centre

Appendix A. Combustion model validation for a 1.6 liter, fourcylinder, TDI engine

Simulations have been performed for a 1.6 liters, four cylinder,and turbo direct injection (TDI) engine with the WAVE code [19].The main parameters of the engine are given in Table A1. IVO is theintake valve opening, IVC the intake valve closure, EVO the exhaustvalve opening and EVC the exhaust valve closure. IVO of 350°means the intake valves open 10° before TDC. IVC of 550° meansthe intake valves close 10° after BDC. EVO of 160° means the exhaustvalves open 20° before BDC. EVC of 360° means the exhaust valves

Table A11.6 litre I4 TDI engine geometry for computational purposes.

Displacement per cylinder [m3] 4.023E−04

Number of cylinders 4Engine layout Inline four (I4)Compression ratio (CR) 16.5Bore B [m] 8.0E−02Stroke S [m] 8.0E−02Connecting rod length [m] 1.45E−01Wrist pin offset [m] 0Clearance volume [m3] 2.594E−05Engine type Compression ignition (C.I.)No. of intake valve per cylinder 2Intake valve diameter [m] 2.16E−02Intake valve maximum lift [m] 7.44E−03No. of exhaust valve per cylinder 2IVO [deg] 350 (−10)IVC [deg] 550 (+10)Exhaust valve diameter [m] 1.92E−02Exhaust valve maximum lift [m] 7.44E−03EVO [deg] 160 (−20)EVC [deg] 360 (0)

close at TDC. Simulations have been performed for 13 engine speeds,4500, 4000, 3750, 3500, 3000, 2750, 2500, 2250, 2000, 1750, 1500,1250, and 1000 rpm, and 14 loads for each speed, for a total of 182points wheremeasured values are available. Simulations are performedreducing the load by increasing the air-to-fuel ratio starting from thefull load experimental values. The equivalent injection duration is sim-ply takenproportional to themass offluid injected. The start of injection(SOI) is adjusted to optimize the BMEP for same mass of fluid injected,i.e. to optimize the brake fuel conversion efficiency. A trial and errorprocedure is necessary, so runs have to be repeated until the requiredBMEP is delivered for every speed. The real injection and combustionprocesses differ considerably from those simulated. However, beingthe total mass of fuel injected about the same, the equivalent start of in-jection may be adjusted to provide a similar pressure build up process.The same equivalent single injection applied to the real engine is notexpected to provide the same performances for multiple reasons, notonly because of the user-entered Cetane number differing considerablyfrom the diesel Cetane number and the Wiebe representation being anoversimplification (a single injection will bring back the operation ofthe diesel TDI also for what concerns efficiency to what was possiblebefore the introduction of the common rail).

Fig. A1 presents the computed full-load boost pressure, air andfuel flow rates and maximum cylinder pressure, brake mean effectivepressure BMEP, brake power, brake efficiency and brake specific fuelconsumption BSFC vs. engine speed. In terms of gas exchange pro-cesses, complicated by the turbo charger operation, the difference inbetween the measured and the computed flow rates as well as theboost pressure is within a few percents, except than in the low enginespeed region, where the interpolation of the not very accurate mapsof turbine and compressor provides a much less precise result.Other factors may certainly affect the low speed accuracy for theboost pressure as well. Worth of mention is also that the experimen-tally measured large variation of boost pressure with engine speed inthe 1000 to 1500 rpm range is much more difficult to be properlymodeled than the almost flat boost pressure in the 1500 to 4500 rpmrange. Including combustion, the maximum cylinder pressure and thebrake mean effective pressure are well computed, with remarkableexceptions only in the low engine speed region for the reason outlinedbefore. The prediction of BMEP is accurate within a few percents exceptthan in the low engine speed region for gas exchange reasons. In termsof brake specific fuel consumption, the most important parameter forfuel economy and the one more directly linked to the accuracy of thecombustion simulation, the differences are just a few percents all overthe range of engine speeds. This result proves that despite the very sim-plified representation of the complex multi event injection and com-bustion evolution an equivalent single injection of all the fuel and asemi predictive model as the one outlined above may properly predictthe pressure build-up and the heat transfer to the cylinder walls. Differ-ences in between experiments and simulations are the results of the ac-curacy of the many models adopted to describe the many relevantphenomena. Part load results are finally presented. Fig. A2 presentsthe part load brake specific fuel consumption vs. brake mean effectivepressure at engine speeds 4500, 4000, 3500, 3000, 2500, 2000, 1500and 1000 rpm. The brake mean effective pressure is the best loadparameter especially in a lean burn compression ignition enginewhere the load is controlled by the quantity of fuel injected. For agiven speed, the power is proportional to the brakemean effective pres-sure, and the comparison of brake specific fuel consumptions tell usabout the amount of fuel needed to produce the desired output. The dif-ferences in measured and computed brake specific fuel consumptionsare generally a fewpercents over the load range all over the range of en-gine speeds, with differences increasing reducing the loads to show aless accurate modeling of combustion with higher air-to-fuel ratios aswell as to the increased weight of friction mean effective pressure vs.the indicated mean effective pressure resulting from the gas exchangeand combustion simulation. This result confirms the ability of the

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Fig. A1. Measured and computed boost pressure, air flow rate, fuel flow rate, maximum cylinder pressure brake power, brake mean effective pressure BMEP, brake specific fuelconsumption BSFC and brake efficiency vs. engine speed full load (left to right and top to bottom).

106 A. Boretti / Fuel Processing Technology 113 (2013) 97–108

simple combustionmodel to properly predict the pressure build-up andthe heat transfer to the cylinder walls over the full range of speeds andloads. A satisfactory prediction of the pollutant formation, CO, HC and

NOx, as well as of the smoke, soot and combustion noise is out ofreachwith the simplified approach, and other tools are certainly neededto address these issues. Simulations performed for brake specific CO, HC

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Fig. A2.Measured and computed brake specific fuel consumption vs. brake mean effective pressure at engine speeds 4500, 4000, 3500, 3000, 2500, 2000, 1500 and 1000 rpm (left toright and top to bottom).

107A. Boretti / Fuel Processing Technology 113 (2013) 97–108

108 A. Boretti / Fuel Processing Technology 113 (2013) 97–108

and NOx with the simple models available in WAVE do not provideaccurate results over the full speed and load range. Computation ofsmoke and soot as well as of the combustion noise is not even allowedin WAVE. These tools are therefore only accurate engine dynamometerexperiments.

References

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[2] www.fischer-tropsch.org/DOE/DOE_reports/ev45a_2_f3c/ev45a_2_f3c_sec01.pdf,(retrieved February 9, 2011).

[3] J.E. Glidewell, Use of Lpg in Truck Engines, SAE P, 550226, 1955.[4] F. Knecht, F. Papts, M. Signer, New IVECO LPG-Engines for Trucks and Buses, SAE

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Buses and Trucks and their Emission and Cycle Variability Characteristics, SAEP, 2001-01-0144, 2001.

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[11] S. GOTO, et al., Performance and Emissions of a DI diesel engine Operated withLPG and Cetane Enhancing additives, SAE P, 2003-01-1920, 2003.

[12] A. Boretti, Direct Injection and Spark Controlled Jet Ignition to Convert A dieselTruck Engine to LPG, SAE P, 2010-01-1976, 2010.

[13] H. Watson, M. Ali Khan, 4 L Light Duty LPG Engine Evaluated for Heavy DutyApplication, SAE P, 2010-01-1463, 2010.

[14] K.B. Hodgins, P. Ouellette, P. Hung, P.G. Hill, Directly Injected Natural Gas Fuelingof Diesel Engines, SAE P, 961671, 1996.

[15] P.L. Mtui, P.G. Hill, Ignition Delay and Combustion Duration with Natural GasFueling of diesel Engines, SAE P, 961933, 1996.

[16] www.westport-hd.com/technology.php, (retrieved February24, 2011).[17] www.westport.com/core-technologies/fuel-injectors, (retrieved February 24, 2011).[18] www.gtisoft.com/, (retrieved September 29, 2010).[19] www.ricardo.com/What-we-do/Software/Products/WAVE/, (retrieved February

9, 2011).[20] www.afdc.energy.gov/afdc/pdfs/fueltable.pdf, (retrieved February 9, 2011).[21] webbook.nist.gov/chemistry/fluid/, (retrieved February 9, 2011).[22] www.greencarcongress.com/2009/03/high-pressure-d.html, ([Internet]. [Cited 2011

July 20]).[23] H. Eichlseder, C. Spuller, R. Heindl, F. Gerbig, K. Heller, Concepts for Diesel-like

Hydrogen Combustion, MTZ Worldwide Edition, 2010., (www.atzonline.com/index.php;do=show/site=a4e/sid=7569549924dbde0eca7dd0586568901/alloc=3/id=10799 [Internet]. [Cited 2011 May 20]).

[24] H. Eichlseder, C. Spuller, R. Heindl, F. Gerbig, K. Heller, New and InnovativeCombustion Systems for the H2–ICE: Compression Ignition and Combined Processes,SAE P, 2009-01-1421, 2009.

[25] www.volvotrucks.com/trucks/global/en-gb/trucks/new-trucks/volvo_fh/engine-Program/Pages/intro.aspx, (retrieved May 4, 2010).

[26] A. Boretti, Improvements of Truck Fuel Economy using Mechanical RegenerativeBraking, SAE Technical Paper, 2010-01-1980, 2010., http://dx.doi.org/10.4271/2010-01-1980.

[27] A. Boretti, C. Grummisch, 100% LPG Long Haul Truck Conversion — Economy andEnvironmental Benefits, SAE Technical Paper, 2012-01-1983, 2012., http://dx.doi.org/10.4271/2012-01-1983.