Urban Transport VI, C.A. Brebbia & L.J. Sucharov (Editors ...€¦ · Engine (four-cycle) Cylinders Combustion system Maximum engine speed at full load Rated brake power Maximum brake

Natural gas combustion in diesel engine

environment: experiments and computations

C. Mansour, A. Bounif & A. Aris

Universite des Sciences et de la Technologic d'Oran, Algerie

Abstract

The aim of this paper is to investigate the emission and performancecharacteristics of a commercial diesel engine (Deutz FL8 413F) being operatedon natural gas with pilot diesel ignition. In parallel, a numerical simulation of theGas-Diesel (Dual-fuel) engine has been carried with a modified Perfect StirredReactor (PSR-Chemkin ) code. A large number of reaction intermediates wereidentified, and their concentrations were followed for reaction yield rangingfrom low conversion to the formation of the final products. A detailed chemicalkinetic reaction mechanism describing the natural gas oxidation and NO%reduction was built to produce the experimental results. The experimentalemission concentrations are compared with the predictions of the model. Areasonably good prediction of the emission concentrations was obtained bycomputation covering the whole range of the experiments.

1 Introduction and background

The dramatic problem of pollution that is generated by the use of conventionalfuels is growing up seriously. Diesel engine has been identified as being one ofthe sources that are responsible of most toxic pollutants like soot and others.Even if the natural gas is widely utilised in many fields as a source of power likeheaters, bowlers and turbines, its utilisation in vehicles is not generalised. Manypossibilities exist for the utilisation of natural gas in diesel engine. G.A. Karim[1] was the first who has utilised the natural gas in a diesel engine. Heintroduced by his experience the dual-fuel mode. The advantage of this enginetype, which plays on the difference of flammability of two fuels, is that, in caseof lack of gaseous fuel, it is possible to run according the diesel cycle; switchingbeing possible when running and without load variation. The disadvantage of

Urban Transport VI, C.A. Brebbia & L.J. Sucharov (Editors) © 2000 WIT Press, www.witpress.com, ISBN 1-85312-823-6

490 Urban Transport and the Environment for the 21st Century

such an engine is the necessity to have liquid diesel fuel available. Theoretically,the liquid fuel quantity necessary to fire is tiny (less than 1 %), but, it is notpossible to inject with the given pump and injectors assemblies fuel quantitiesvarying of 1 to 100 %. If we do not want the material to be doubled, we must besatisfied by injection of the minimum possible with standard diesel material ,that is the rate of 5 to 25 % of full load and this, whatever be the developedpower. The natural gas requires a high activation energy to ignite. This latter isgiven by a small rate of diesel fuel (25% of the normal rate). The power and thespecific consumption are then lower by 25 to 30% in dual-fuel mode [2,3]. Thisgives us the opportunity to simulate the natural gas combustion under dieselengine conditions, in order to understand how the fuels are burned during thecombustion process. The natural gas that we have experimentally used [4], isconstituted with methane which is the main constituent with 82.82%, 7.38% ofC2H4 and 2.07% of C3H8. In many surveys [6,14,15], the kinetic schemes usingthe chemical kinetics mechanisms are used under constants conditions, likepressure, temperature, residence time. However, very little detailed experimentalresults of motors parameters in dual-fuel mode are available to models. In themotor, all the parameters are continually changing in time, letting the simulationexcessively complicated. A large number of reaction intermediates wereidentified, and their concentration were followed for reaction yield ranging fromlow conversion to the formation of the final products. A detailed chemicalkinetic reaction mechanism describing the natural gas oxidation and NO%reduction was built to produce the present experimental results. The results ofthe experimental investigation like the stoichiometric ratio determined from theoxygen concentration in the exhaust manifold has been used in the simulation.The principal diagnostic parameter was the measured time history of the cylindergas pressure, in the way of verification of the model law (pressure, temperature).

2 Experimental performance

The engine used in this study was a natural-aspirated, eight-cylinder version of theDeutz FL8 413F diesel. The basic engine characteristics of the test engine aresummarised in table I. The natural gas used was introduced into intake air stream.The gas flow is measured by the fine wire anemometer and controlled using amanual, variable area and fine control needle valve. Engine intake air was filteredand measured with the laminar flow element using the pressure measurementsdifference. The diesel fuel flow rate is measured with two volumetric pumpsconnected to the inlet manifold of the injection pump and to the outlet manifold.The volumetric pumps are connected by two photodiode cells to the computer dataacquisition. The engine test facility (LPS 2000) was equipped with the digitalreadout of engine speed, torque and power. A strain gage amplifier was placed inparallel to the dynamometer strain gage and a frequency to -voltage circuit wasplaced in series with the magnetic pick-off the dynamometer.


Urban Transport and the Environment for the 21st Century 491

Table I: Test Engine Specifications

Engine (four-cycle)CylindersCombustion systemMaximum engine speed at full loadRated brake powerMaximum brake torque at 1500 rpmBore x StrokeCompression RatioDisplacement

Diesel FL8 413F8in-VDirect injection2500 rpm230 Ch735 mN125/130 mm18:112761 cc

Figure 1 represents the evolution to full load of the brake power andthe brake torque with different engine speed following the engine norm (DIN70020). One notes a light power and torque loss in dual fuel version with regardto the diesel one , except around the speed of 2400 rpm (cat speed of the fuelinjection pump). This difference could be explained by the response time ofnatural gas system of injection in the admission collector. The brake torque andthe brake power in diesel version are slightly higher than in the dual-fuel mode.This is the results of the interruption of the dual fuel mode when the speed enginerises up to 2450 rpm.

3 rake power (k W ) Brake torque (Mm)

+ Power Dieselx Po\v er dual fuel* Torque Dieselo Torque dual fuel

Engine speed (rpm)

Figure 1: Brake power and brake torque vs engine speed for full load

Figure 2 shows the evolution of the specific power and consumption with theengine speed. One notices that at low engine speed the difference is importantbetween the two version ( loss of power and more elevated consumption), but inthe high regimes the gap becomes less important. The evaluation of the specificconsumption was based on the experimental results of the simulator computerand the consumption results.



Ps (Kw/l) Cs(Kg/Kw/h)12 on 200 on10.80 -i9.60 \8.40 \7.20 J6.00 J4.80 J3.60 J2.40 41.20 \0 00 :

o -o

)r *- ^ ___—-^^

+ Specific power Dieselx Specific power dual-fuelo Specific cosumption dual-fuel* Specific co sumption Diesel

- 160.00

_ 120.00

. 80.00

_ 40.00

1 0.00Engine speed (rpm)

Figure 2 Brake specific fuel consumption and brake specific poweras a function of engine speed for full load

3 Empirical gas-diesel combustion model

The diesel simulation has been carried out with a Wibbe model [8], for threespeeds and two air ratios, in order to determine the influence of one of theseparameters on the maximum temperature in the cylinder during the combustioncycle. The Cylinder gas pressure was measured using a piezoelectric transducermounted in the main combustion chamber, connected with a digital acquisitionsystem (DAS 1401), showing the good representation of the law used in themodel. The different values obtained using the model are in agreement with thereal variation in the engine. The cylinder pressure data measured for this enginefor natural gas fuelling showed (Fig. 3) the peak pressure occurring between 7 a15 cranks angles later (depending on engine load and speed) than thecorresponding peak pressure for diesel fuelling. Figure 3 showed a double humpon the pressure trace, which is explained as combustion of the pilot diesel chargefollowed by the natural gas combustion. The maximum combustion pressure areslightly higher for natural gas fuelling for all engine speeds than the diesel level.The general trend is decreasing pressure and temperature levels with increasingspeeds.

Figure 3: Cylinder pressure data for Diesel and dual fuel gas fuelingfor full loads at 1000 rpm engine speed


Urban Transport and the Environment for the 2 1st Century 493

4 The reaction mechanism

The reaction mechanism of the oxidation of the natural gas used in this presentwork has been partially published previously ([6,14,15,16, 17]). A large numberof reaction intermediates were identified, and their concentration were followedfor reaction yield ranging from low conversion to the formation of the finalproducts. A detailed chemical kinetic reaction mechanism describing the naturalgas oxidation and NO% reduction was built to produce the present experimentalresults. This mechanism includes 573 elementary reactions, most of them beingreversible, among 90 chemical species.In the case of reversible reactions, the rate constants for the reverse reactions arecomputed form the appropriate equilibrium constants and the modifiedArrhenius parameters for the forward reactionsThe kinetic parameters for NO% composition were taken from the Zeldovichmechanism[24]. The reaction that represent the diesel fuel C^H^was taken fromthe scheme of Hautmann and and al.[18].

2.

4. H^+mO^H/

The residence time is correlated with the crank angle of the engine the relation[19]: dcp=6Ndt. The rate of the pilot injection is taken from the experimental test.The minimum temperature required for reasonable diesel engine ignition delay(<2ms) was taken from the results in the survey of Naber and al. [19]. Theauthors have found experimentally that in the dual fuel mode the combustion ofnatural gas begins when the temperature in the cylinder rise to 1 150 °C. Eraserand al [20] noted in their investigation that the effects of pressure on ignitionwere small compared with temperature effects. From the analysis of the resultsof the diesel simulation, the different points that correspond to the initiation ofthe combustion are then determined. The combustion under diesel engineconditions was modelled using the computer codes (chemkin) developed atSandia by Kee and co-workers [21,22]. The transport properties andthermochemical quantities from the Sandia data were also used [23,24]. We haveused the Burcat thermochemical data [25] for the compounds not found in thechemkin data base. The mathematical model used in the way of resolution ofstiff system was previously published [26].

5 Simulation results:Figure 4 presents the pressure, temperature and rate of the heat

production evolution in the cylinder during the closed phases of engine for 1000rpm engine speed with a static advance to the injection system of 16 DV. Onenotices a displacement of the peaks of the maximal sizes after TDC (top dead



centre) with the increase of the engine speed, in spite of the centrifugalcorrection of the injection system for the dynamic advance, which is related tothe engine speed. This remark is due essentially to the induction time of the auto-ignition reaction of the fuel in a very lean mixture (in the first phase, the naturalgas is considered as an inert gas; for the auto-ignition phase of fuel , the airexcess is multiplied by 4 with regard to the initial dilution).

Crank angle (Degree)Figure 4: Cylinder pressure, temperature and rate of heat productionevolution simulation results for 1000 rpm engine speed at full load.

The combustion study has been carried out with good results. It is clear that thediesel fuel burns first giving all the energy required to ignite the lean mixedgases (natural gas and air).The range of pressure and temperature experienced bythe reactants (mixture fuel-gas natural/air) drawn into the cylinder is particularlywide (Fig. 4), and in consequence, the variety of chemical reactions that takeplace is very extensive.

[CO ;CO, ;NOJ [C^ ;C,,H

Time (s)Figure 5: Molecular species profiles obtained form the dual-fuel mode

simulation for 1000 rpm engine speed at full load.



We have defined the characteristic chemical time, of natural gas oxidation fromthe computed concentrations profiles as the time delay during closed residencetime after auto-ignition fuel period estimated by [12]. Nevertheless, for operatingconditions; engine speed higher than 1500 rpm, where such a conversion rate isnot realised at time shorter than the residence time, it is obvious that thechemical and average residence time scales should be considered as identical.Concentration profiles of major products (carbon monoxide, hydrogen, carbondioxide, water, methane, ethylene, ethane, propane and fuel) are deduced from asimulation work taking into account two-time step detailed kinetic model (Fig. 5and 6). At high engine speed, during the first stage of the reaction when the fueland oxygen are consumed at a wide rate, (CjH €2 ) is formed in appreciableamount (the fuel auto-ignition period is less than 0.5 ms, shows in figure 5) arank ordering of the reaction rates indicates that the most important path is theHautmann reaction mechanism. Carbon monoxide appear to be formed with amaximum in relatively great quantity (have been identified in the rank orderCjHL, + Oj 2 CO +2H% ). Intermediate species capable to induce chainbranching, such as hydrogen peroxide and aldehydes have been identified in therank order (HCO + O; -> CO + HO,, HO, + HO, -> (^ + H O,, ) [19], areformed in the first flame reaction zone and disappear later in the normal gasnatural oxidation. Satured or olefmic hydrocarbons, are mainly formed all alongthe reaction evolution. The selected computed mole fractions are presented infigure 5. The predicted results is generally appear clearly that the two steps fuel-gas natural mechanism qualitatively accounts for the observed auto-ignitiondelay towards the experimental pressure evolution in the cylinder. According tothe important role of CO as a typical major product of two-phase flame and ofHzC as a branching agent, a reaction path analysis has been undertaken for thesespecies. Carbon monoxide is mainly formed by reaction involving HCO species(HCO + O; -> CO + HOz,) In the second flame, hydrogen peroxide destructionproceeds principally trough homogeneous decomposition which appears as animportant step to convert HO2 radical in to the more reactive OH radical. Onethere remarks that temperature and pressure in the cylinder become weaker whenthe engine speed decreases (the residence time is shorter than the characteristicchemical time ). At high temperature in range of engine speeds less than 1500rpm (the fuel auto-ignition delay is more than 5.5 ms; auto-ignition appears as asingle-stage flame and the induction period decreases with pressure. Thepredicted NO masse fraction by a reduce mechanism with 7-steps includingthermal NO chemistry, decreases as the engine speed in full load increases. Thisresult can be attributed of the increase of the NO reactivity due to increase areaction pressure and temperature with the residence time.



Moles/cc-sec9.00E-0 1

Figure 6: Rates of production of species (CH4; C2H6; C3H8; C6H10) vs thetime at 1000 rpm engine speed

6 Emission results

Emissions of total hydrocarbons (C HJ, carbon monoxide (CO), carbon dioxide(CCy, oxides of nitrogen (NOJ were measured in the raw exhaust of theengine. The sample stream of raw stream was cooled and dried in a thermostaticsample conditioner before being analysed by the flame ionisation detection ( FIDRS55 used for C^HJ, non dispersive infrared (NDIR Beryl used for CO andCOz), and chemiluminescent (Topaze 2020 used for NOJ instruments. Eachinstrument is equipped with 0-100 mV output. The voltage output is proportionalto the concentration of pollutants in the exhaust. NO% and CO emissionconcentrations measured in the exhaust gas as a function of engine speed for fullload for diesel only and natural gas fuelling are shown in Figure 7. The NO%emissions are affected directly by the change in air dilution to the engine. Thenatural gas-air mixture leaning with increasing engine speed, which wouldinduces a slower flame propagation. NO% concentrations were reduced at highspeed ( greater than about 1500 rpm) for natural gas fuelling. CO emissions wereincreased for natural gas fuelling. The NO* emission characteristic of the gasengine is determined by the air/fuel ratio when the ignition timing is constant. Itis widely known that the concentration of NO% emission reaches maximum inthe vicinity of the excess air ratio around 1.1, and decreases sharply in the rangeof high air / fuel ratio , so called lean mixture ( at the level of high air / fuel ratiocombustion temperature drops and NO% concentration lowers exponentially).The general trend is increasing air-fuel ratio with increasing engine speed fornatural gas fuelling.



|CO] (ppm)

Engine speed (rpm)Figure 7: NO% and CO emissions as a function of speed

for full loads dual-fuel / Diesel

(ppm)12001080 -j960 i840720 J600 -I480 4360240120 -i0

dual-fuel (exp.)dual-fuel (th. NO)dual-fuel (global. NO)

'12'40' '15X0' ' 1 9 2 0 ' ' 2 2 ' 6 0 ' 266 0Engine speed (rpm)

Figure 8: NOx emissions as a function of speed for full loads dual-fuel(simulated results, : thermal NO, O: complete mechanism, +: experimental results,)

5.0E-05

().()E+(X) 7.5E-03

time (s)Figure 9: Profiles of the NO% emissions for two speeds (1000,

2000 rpm) and two equivalence ratios (<()=05,0.7)



The ignition delay time increases at high air /fuel ratio, thus in real engine, thepotential combustion period is limited by the engine speed and the combustionof fuel gas tends to be incomplete. The quench area near the combustionchamber wall increase as air/fuel ratio goes higher. As a result it is supposed thatthe thermal efficiency will go down and unbumed HC and CO will increase (Fig.7). It is expected that the elevated CO concentrations would be accompanied byhigher unbumed hydrocarbons C H . CO2 emissions were decreased for naturalgas fuelling. Ignition of lean natural gas mixtures is difficult to achieve and canresult in incomplete combustion or total misfire. For ignition to be successful,the energy release rate in early stages of ignition must be greater than lossesfrom ignition flame kernel. If not, the flame extinguishes prematurely. For leanmixtures, the energy release per unit volume is less because the fuel charge isdiluted with excess air. Figure 8 shows the measured and the computedconcentration of NO% as a function of engine speed for full load. The computedvalues are generally over-predicted by the Zeldovich mechanism. Nevertheless,we observed (Fig. 8) a good agreement between experimental and computedconcentration profiles reproduced by the global mechanism. The predicted NOmasse fraction by a reduce mechanism with 7-steps including thermal NOchemistry, decreases as the engine speed in full load increases. This result can beattributed of the increase of the NO reactivity due to increase a reaction pressureand temperature with the time ( Fig.9 )

Conclusion

The emission and performance characteristics of a commercial diesel engine(Deutz FL8 413F) being operated on natural gas with pilot diesel ignition wereinvestigated. A computer program has been developed to model theexperimental data using a chemical kinetic reaction mechanism of the Gas-Diesel (Dual-fuel) combustion. The credibility of this work depends on theability of the numerical model to restitute the instantaneous fuel conversion ratesin all the experimental engine conditions, not only for prediction of the amountof unbumed fuel, but also for prediction of the concentration of species whichparticipate in both the combustion and the pollutant species reactions. Theresults of the comparison between computed and experimental results appear inFigures 7 and 8 where we can see that the experimental data are well reproducedby the chemical kinetic reaction mechanism of the natural gas- diesel fueloxidation. Oxides of nitrogen (NOJ emissions were found to be lower at highengine speed under high load. Carbon monoxide (CO) emissions were increasedfor natural gas fuelling while carbon dioxide (COj) were reduced.



Acknowledgement

This research was supported by GNV-Project (SONATRACH L.T.G) which isgratefully acknowledged. We also would like to thank Prof. M. Cathonnet forproviding a copy of the natural gas mechanism and helpful comments on themanuscript.

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Urban Transport VI, C.A. Brebbia & L.J. Sucharov (Editors ...€¦ · Engine (four-cycle) Cylinders Combustion system Maximum engine speed at full load Rated brake power Maximum brake