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A SIMULATION TOOL FOR EVALUATING GAS COMPOSITION EFFECTS ON ENGINE PERFORMANCE

C. Caillol and G. Berardi

Ecole Supérieure d’Ingénieurs de Marseille - Centre de Mécanique Energétique France

G. Brecq, M. Ramspacher and P. Meunier Gaz de France - Direction de la Recherche

France

ABSTRACT

In the context of an interconnected gas network with varying natural gas sources, the variation in fuel quality is a specific question for gas engines. Fluctuating fuel composition may particularly affect the combustion quality at lean operating limit conditions. The investigation of the effects of gas fuel composition on the propagation of a turbulent premixed flame in a combustion chamber is then of fundamental interest with regard to the control of natural gas engines.

In this context, Gaz de France has launched the development of a simulation tool to

analyze the influence of gas composition on engine performance and pollutant emission levels. The predictive model is based on a two-zone thermodynamic approach. The premixed flame propagation, in the heat release rate, is controlled by the turbulent mixing process. The outputs from the combustion rate model are compared with measurements from a single cylinder spark ignition engine and from a 10 liter displacement, six cylinder gas engine. For all operating conditions, simulated cylinder pressure histories agree well with the experimental data. The results show that engine turbulence characteristics can be described by a single model parameter and that gas composition effects are mainly related to the laminar burning velocity.

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INTRODUCTION Lean burning of natural gas in spark ignition (SI) engines has the potential to combine

low exhaust gas emissions with high power output and good efficiency. In view of meeting the more stringent emission regulations that will take effect in the near future [1], natural gas is an attractive alternative to conventional energy such as gasoline and diesel fuel, widely used for transportation applications. Mainly composed by methane, natural gas has one of the lowest carbon contents among hydrocarbons, resulting in a potential of CO2 emission 23% lower than that of Diesel fuels. Its high resistance to auto-ignition allows to set high compression ratios, lowering the difference of brake thermal efficiency between Diesel and SI engines. With wider flammability limits, compared to gasoline fuel, it makes possible to operate at leaner air-fuel ratios with a reduction in NOx emissions.

The stability of fuel specifications is an important condition for engine manufacturers

to achieve the best possible tradeoff between high level of power, low consumption, low emissions and the prevention of knock. The natural gas composition varies with the feedstock location. Contrary to conventional liquid fuels, it is a non refined product : its specifications can not be controlled precisely and only depend on the origin of the fuel. Variations in fuel quality can bring about significant changes in the combustion characteristics. With the need to specify an acceptable gas quality standard for safe operation of on-board compressed natural gas (CNG) storage systems and the acceptable operation of CNG engines and vehicles, the Gas Research Institute carried out a major survey of the variation of natural gas composition across the U.S. [2]. The SAE Recommended Practice J1616 [3] summarizes the results of this study, examines the more important physical and chemical characteristics of CNG fuel and describes pertinent test methods to evaluate fuel properties. In order to maintain and preserve some kind of energetic independence, governments and gas trading companies tend, as far as possible, to multiply the number of natural gas sources. Consequently, the types of natural gases (and the possible mixtures between them) being possible to encounter on gas supply networks are more and more numerous (especially on interconnected grids), increasing the range of fuel quality possible to use in the engine. Because the application of natural gas to engine represents up to now a small part of its utilization, no centralized process can be applied to the whole quantity of the gas supplied to the consumer and, on the other side, no process is sufficiently affordable to be applied directly in the fuel station.

Although vehicle manufacturers are strong advocates of fuel standardization, there are

solutions for dealing with fuel quality variation associated with natural gas, in particular fuel flexible engine control. A recent joint research program was conducted by companies active in the automobile and natural gas industries to collect information on the operating behavior of natural gas vehicles with changing gas properties in view of developing strategies for the adaptation of the engine to gas properties [4]. Implementing fuel flexible engine control demands not only an adequate sensor technology but also the knowledge

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of the effects of gas composition on key combustion properties : brake thermal efficiency, pollutant formation and knock onset. Part of this knowledge is condensed in the so-called fuel quality indices such as Methane Number, Stoechiometric Air/Fuel Ratio, etc. This knowledge is very useful to define what process could be applied to the gas, or what specifications have to be regulated. However no simulation tool provides full information on gas engine performance depending on the gas composition.

The purpose of this study is then to produce a predictive tool able to analyze gas

composition effects on engine power and consumption and on pollutant emission levels. This paper mainly focuses on the modeling of the cylinder pressure which is the first information to control.

NUMERICAL MODEL

Two Zone Thermodynamic Model

The predictive model developed in this work is mainly based on the mathematical

formulation of the conventional two-zone approach proposed by Ferguson [5]. This zero-dimensional thermodynamic model assumes that at anytime of the cycle, the cylinder volume is divided into two zones, corresponding to burned and unburned gas fields. The two zone are always separated by an infinitesimally thin flame front with a hemispherical shape. The thermodynamic state is defined in each zone by the mean parameters and the specific heat of each gas (formulation from the table of the thermodynamic properties JANAF). The burned gases are assumed to be in chemical equilibrium (each element of the fuel-air mixture which combusted mixed instantly with the previously burned gases), and both zones are assumed having the same uniform cylinder pressure. The model takes as input data :

• engine geometry, • operating conditions (ignition timing, air-fuel ratio, gas composition, etc.), • ambient conditions, • in-cylinder conditions at inlet valve closure.

The region in the combustion chamber is treated as a control volume and the

governing equations are the mass and energy conservation equations and equations of state. A formulation of the burned mass fraction bx is used to describe the spatial dynamics of the combustion process. The energy equation for an open system bounded by combustion chamber walls is :

ll

mdu dm dV dQm u p hd d d dθ θ θ θ ω

+ = − + −&

(1)

where θ is the crank angle, ω the angular speed, m the total mass in the cylinder, p the gas pressure, V the cylinder volume, Q the heat transfert and u the specific internal

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energy. The unburned mixture and burned mixture zones are each treated as separate open systems. Hence, the specific internal energy and volume are given by :

( )1b b b uUu x u x um

= = + − (2)

( )1b b b uVv x v x vm

= = + − (3)

where U is the internal energy, V the total cylinder volume, uu and uv are the specific internal energy and volume of unburned gases at temperature uT , bu and bv are the specific internal energy and volume of burned gases at temperature bT . The thermodynamic gas properties as function of pressure, temperature, and equivalence ratio for the mixture of air, fuel and residual gas fraction, and for the mixture of burned gases at equilibrium are obtained by using the method proposed by Olikara and Borman [6]. This equilibrium constants based method solves for chemical equilibrium compositions, specific heats, internal energies, enthalpies and entropies. Combustion products composition considers 10 chemical species including the overall complete combustion products : CO2, H2O, N2 and O2, and species from dissociation reactions : CO, H2, H, O, OH and NO.

On the right-hand side of equation (1), the heat transfer into the thermodynamic

system is expressed in terms of the heat loss from the burned and unburned gas :

b uQ QdQdθ ω

+= −

& & (4)

Heat transfer between gas and cylinder walls in spark ignition engine can be assumed to be convective :

( ),b g b b b wQ h A T T= −& (5)

( ),u g u u u wQ h A T T= −& (6)

where the wall temperature wT is assumed constant and uniform and equals to 400 K [7], bA and uA are the combustion chamber wall areas in contact with burned and unburned

gases respectively. The surface areas are computed as function of cylinder bore b and instantaneous combustion chamber volume V with the following relations :

2

1 242b bb VA x

bπ

= +

(7)

( )2

1 24 12u bb VA x

bπ

= + −

(8)

Equations (7) and (8) assume that the fraction of the cylinder area exposed to burned gas is proportional to the square root of the burned mass fraction which reflects the fact

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that burned gas occupies a larger volume fraction than the unburned gas [5]. The transfer coefficients are estimated with the Hohenberg correlation [8] :

0.06 0.8 0.4 0.8, 1 2( )g b b b ph C V p T v C− −= + (9)

0.06 0.8 0.4 0.8, 1 2( )g u u u ph C V p T v C− −= + (10)

where the constants 1C and 2C are experimentally adjusted, bV and uV are the volume of the burned and unburned gases respectively, and pv is the mean piston speed.

The last term in equation (1) represents the energy flow due to blow-by. The enthalpy

loss is expressed as :

( )2 21l b b b uh x h x h= + − (11)

where bh is the specific enthalpy of burned gases, uh is the specific enthalpy of unburned gases, and which indicates that more leaking is due to the unburned gas compared with the burned gas in the early stage of combustion. The instantaneous leakage rate lm& is assumed to be always out of the cylinder and is computed with the blow-by constant

b lC m m= & dependent upon ring design. The model described with equation (1) through equation (11) constitutes a set of

ordinary differential equations to predict the cylinder pressure, burned and unburned temperatures of gases, work done, heat loss and enthalpy lost. The numerical integration of this system, with crank angle as the independent variable, is obtained by using a fifth-order Runge-Kutta method. Closure for this system is achieved through a combustion model describing the evolution of the burned mass fraction.

Combustion Model

The purpose of incorporating a combustion model is to establish a formulation of the

combustion rate related to the operating conditions and engine parameters. The model is based on the approach developed by Chmela et al. [9] in order to increase the predictive accuracy and to reduce the tuning efforts when applying the numerical model to different engines and various operating conditions. This phenomenological analysis makes use of the eddy dissipation concept (EDC) developed by Magnussen and Hjertager [10] for prediction of gaseous combustion reactions in turbulent flows. The EDC model is based on the assumption that the chemical reaction rates are fast compared to the mixing. Thus, the reaction rate is determined by the rate of intermixing of reactants and products, i.e., by the rate of the dissipation of turbulent eddies. Accordingly, in the case of lean premixed conditions, the mean rate of combustion of fuel can be expressed by :

F EDC FC ckεω = −& (12)

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where EDCC is a dimensionless constant, Fc is the local mean fuel concentration, ε is the rate of dissipation of turbulent kinetic energy, and k is the turbulent kinetic energy. Considering the local mass of fuel Fm , the rate of heat release by combustion in the reaction zone can be related to the mean rate of fuel disappearance. This leads to the following expression :

C EDC LHV FQ C Q mkε

=& (13)

where LHVQ is the lower heating value. The reactants are contained in a reaction zone of thickness Rδ which propagates

through the cylinder volume with the turbulent flame speed TS . The flame front is assumed to have a hemispherical shape and the local mass of fuel in the reaction zone is assumed to decay with the combustion progress, which is described by the burned mass fraction bx . The local mass of fuel has then the following form [11] :

( ) 2 22 11

bF b R Tm x S t

rρπ δ

φ= −

+ (14)

where φ is the equivalence ratio, r is the stoichiometric ratio of the mass of air to the mass of fuel, bρ is the burned gas density, and t is the time elapsed since the start of combustion.

For a complete closure in equation (14), the turbulent flame speed is correlated to the

laminar burning velocity LS and to the turbulence intensity u′ . In this work, the model proposed by Zimont and Lipatnikov [12], which is valid in the thickened flame regime, is adopted :

1 2

1 4 1 4Pr ReTT

L L

S uAS S

′=

(15)

where A is a model constant, ReT is the turbulent Reynolds number, and Pr is the Prandtl number. The laminar flame speed of the fuel components is evaluated with an analytical approximation formula based on the asymptotic description of premixed flame.

The rate of dissipation of turbulent kinetic energy ε is expressed by :

3 2

DkCL

ε = (16)

where DC is a dimensionless constant and L is the integral length scale. Relating the rate of heat release to the combustion rate bx&, and combining equations

(13), (14), (15) and (16) with crank angle as the independent variable, leads to the following equation :

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( )2

0 1b bT L b

dx C S xd m

ρ θ θθ ω ω

− = −

(17)

where 0θ is the crank angle at the start of combustion. All parameters related to turbulence in the combustion rate formulation are condensed into a single model parameter TC . This parameter is assumed to be a constant independent of the crank angle all-over the flame propagation phase, and is expressed as follows :

1 2

2 22 Pr Re3T EDC D T R

kC A C CL

π δ =

(18)

The combustion law is used between two crank angle limits : • the start of combustion angle (spark timing + ignition delay), • the crank angle corresponding to a 99% burned mass fraction.

DESCRIPTION OF FACILITIES

Experiments have been designed to allow comparison of engine performance between

various gas compositions fueling of different test engines over a range of operating conditions. Two four-stroke SI gas engines (with very different configurations), have been used for the testing to evaluate the validity of the combustion model.

Single Cylinder Engine

The experimental setup consists of a single cylinder spark ignition KOHLER

Command Pro 6 engine coupled to a direct current motor/generator dynamometer. The engine was naturally aspired and its specifications are provided in Table 1. Tests were conducted under wide-open throttle and full load conditions at a constant speed (2000 rev/min) drive control with the dynamometer.

Table 1 : Engine specifications

Ignition type Spark ignition Number of cylinders & valves Single cylinder with overhead valve Bore x Stroke 67.0 x 51.0 mm Volumetric capacity 204 cm3 Compression ratio 8.5:1

The data acquisition equipment allows simultaneous measurement of steady-state

engine parameters. These included air and natural gas flow rates, engine speed and torque. The air-fuel equivalence ratio was controlled by adjusting the natural gas flow rate with a rotameter and the air flow rate was measured with an orifice flowmeter.

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A separate high-speed data acquisition system enables measurements of in-cylinder pressure as a function of crank angle. The fast acquisition system included an encoder, an AVL 12QP300cvk piezo-electric pressure transducer installed through the cylinder head, and a transient recorder model 6010-20 manufactured by Krenz Electronics. The rotary encoder was coupled to the engine crankshaft and triggered the data acquisition system to sample data at every one crank angle rotation. The top dead center was the reference point from which data was sampled. For each operating equivalence ratio, an average cylinder gas pressure was computed at every crank angle from 300 consecutive recorded cycles.

Three gaseous fuels were used for experimental data in this work : gas G20 (pure

methane), gas C2 (methane and ethane) and gas C3 (methane and propane). Gas properties and composition are listed in Table 2.

Table 2. Gas specifications

G20 C2 C3 Methane CH4 (vol. %) 100 90 93 Ethane C2H6 (vol. %) 10 Propane C3H8 (vol. %) 7 Molecular weight (g/mol) 16.04 17.45 18.01 LHV (MJ/kg) 50.010 49.559 49.371

Six Cylinder Engine

The second engine tested is a six-cylinder turbo-charged natural-gas-fuelled spark-

ignition engine. Adapted from a Diesel version, this industrial engine is mainly employed for CNG city bus applications. Table 3 gives its main technical characteristics.

Table 3. Standard specifications of the RVI engine

Manufacturer RENAULT VI Ignition type Spark ignition Number of cylinders & valves 6 cylinder with 2 valves per cylinder Bore x Stroke 120.0 x 145.0 mm Volumetric capacity 9834 cm3 Compression ratio 11:1 (± 0.5) Maximal couple 1000 Nm (1100 rev/min) Maximal power 186 kW (2100 rev/min)

The test-bench is composed of a magnetic engine brake (Borghi & Savieri FE 350 S)

designed to receive low and heavy duty engines. The main specification of this test bench is to have an online gas synthesizer from which natural gases of any composition can be

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produced. The cooling system of the test bench includes two heat exchangers (operating with water under pressure). The air flow temperature is maintained constant and equal to 50°C. The exhaust gases are analyzed by a SIEMENS Ultramat analyzer (semiautomatic).

Regarding the thermal behavior of the engine, 16 thermocouples with platinum

thermometer probe are arranged at the hotter locations on the engine and at the inlets/outlets of the gas/air/water flows (I/O cooling system, intake and exhaust pipes, etc.).

The test pilot used is the Morphee software (developed by D2T Inc.) which is devoted

to the panel command and the low-frequency acquisition. The high-frequency cylinder pressure acquisition system is composed of a

piezoelectric sensor directly mounted in the cylinder head (Kistler 6052B1, range of 0-250 bar : operating temperature to 400 °C, frequency 130 kHz). This sensor is used with an adequate amplifier system. Signal acquisition and processing are done by the Osiris software (D2T Inc.). The signal requires a pressure reference (pressure in the inlet manifold) to provide the exact level of the cylinder pressure.

For this first campaign on natural gas effects, the engine torque and the speed are

maintained constant during experiments (see Table 4).

Table 4. Engine speed and load

Running speed 1500 rev/min Engine torque 750 Nm PME 9.6 bar Power output 118 kW

A matrix of six pure gases (CH4, C2H6, C3H8, C4H10, CO2, N2) is used to simulate

various natural gas compositions accounting for the main constituents of natural gases in the ranges commonly encountered in Europe. In order to reduce the costs of the experiments, the gas composition is varied by different adjunction of pure gases directly to network gas. The compositions (shown in Table 5) are determined from the mass flow rates and checked by gas chromatography.

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Table 5. Gas compositions (chromatography)

Network gas with No adjunction 5% C2H6 + 2% C3H8 10% C2H6 4% C3H8 7.5% N2

CH4 (% vol) 89.89 81.53 78.8 86.39 83.39 C2H6 (% vol) 5.57 11.91 16.96 5.71 5.46 C3H8 (% vol) 1.1 3.16 0.97 3.81 1.15 N2 (% vol) 1.72 1.83 1.84 1.55 8.4 CO2 (% vol) 1.06 1.31 1.18 1.16 0.95 MV (kg/m3) 0.80 0.86 0.87 0.85 0.84 PCI (Wh/m3) 10417 11152 11242 11068 9799

The gas mixture is based on network gas, which already contains significant amounts

of ethane, propane, and nitrogen. Consequently, a same kind of gas can vary depending on the variation encountered on the network. However, every gas composition has been accurately measured. The RVI gas engines are employed with lean mixtures. Therefore, four levels of equivalence ratio are selected between two limits (lean misfiring and too high temperatures at the exhaust) : 0.66, 0.70, 0.77 and 0.85. The pressure in the inlet manifold is adjusted to counterbalance the modification of the equivalence ratio (pressure is decreased with an increase of equivalence ratio) in order to maintain a constant power (i.e. a rather constant gas flow rate). RESULTS AND DISCUSSION

In the numerical model, two parameters are required to simulate the combustion rate :

the crank angle at the start of combustion 0θ and the turbulence factor TC . They need to be adjusted in order to produce a numerical cylinder pressure fitting with experimental data. In this section, the cases corresponding to the two gas engines are presented, where each is aiming at calibrating the model to the specific engine geometry and for each operating condition.

Model Validation Using the Single Cylinder Engine

In order to evaluate the validity of the model, a comparison is made between the

present model predictions and the measurements obtained with the single cylinder gas engine. The calibration of the model is obtained by adjusting the turbulence parameter TC in order to maximize agreement between simulated and measured cylinder pressure history at a specified engine operating condition. For all operating points, the delay between the start of combustion and the spark timing has been set to 3 crank angle (CA).

Examples of comparison between predicted and experimental cylinder pressure

histories for the three gaseous fuels and for four different equivalence ratio cases ranging from 0.77φ = to 1.01 are given in Figure 1. The reported data correspond to the pressure

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curves for crank angle range between inlet valve closing and exhaust valve opening. For all operating conditions, comparisons show that simulated results using the numerical model agree well with the experimental data. Thus, the combustion rate formulation, assuming that the model parameter TC is a constant independent of the crank angle, is validated. The present model overpredicts cylinder pressure in the expansion phase for cases (c) and (d). This is due to the fact that the blow-by coefficient bC was kept constant over the entire combustion duration and for all operating conditions.

(a) Gas C2 and 0.77φ = (b) Gas G20 and 0.83φ =

(c) Gas C3 and 0.89φ = (d) Gas C2 and 1.01φ =

Figure 1. Predicted and measured pressure traces for different gas composition and operating equivalence ratio

The optimized parameter TC versus equivalence ratio is shown in Figure 2 for the

three gas compositions used in this work. It can be seen that, for a given fuel-air ratio, TC is mostly independent of the gas composition since the difference between the values of the turbulence parameter corresponding to each gas is always lower than 4%. The regular grows of TC in function of the equivalence ratio represents the thickening of the reaction zone Rδ when the fuel-air ratio is increased. Near stoichiometric mixtures, the limit of

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validity of the laminar flame speed formulation for the range of equivalence ratio is approached, resulting in an inaccurate evaluation of TC .

Figure 2. Model parameter TC as a function of equivalence ratio for the three compositions of fuels

Figure 3 shows a comparison between the experimental cylinder pressure trace and the model predictions computed for the limits of the range of the turbulence parameter, corresponding to the extreme values TC =1.45·103 and TC =1.60·103, in the case of gas C2 and equivalence ratio 0.83φ = . We can notice a 4% discrepancy in peak pressure between the predicted results for TC =1.60·103 and the experimental data. These results suggest that engine turbulence characteristics of the single cylinder engine can be described by a mean value of the turbulence parameter and that gas composition effects are mainly related to the laminar burning velocity.

Figure 3. Predicted pressure traces for different TC , for gas C2 and 0.83φ =

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Model Calibration Using the Six Cylinder Engine

Crank angle at the start of combustion. Spark ignited engines require an external

energy source to start the combustion. This is accomplished by discharging electrical energy through the anode and the cathode of the spark plug. The gas ionization caused by the electrical discharge initiates a flame kernel located inside and around the spark gap. The flame kernel then grows into a flame front that begins to propagate within the combustion chamber. The duration between the start of combustion (SOC) and the spark timing (ST) is called the ignition delay δ .

This delay is generally defines as the time required to ignite 1 to 5 % of the mass of

the gases [13]. During this period, the heat released by the combustion is supposed to be zero. After the SOC, the present numerical model is applied. In this study, the ignition delay, indirectly required by the combustion model, needed to be uncommonly short (between 0.5 to 1% of the burn rate), since the combustion model can be successfully applied even in the early stages of the combustion. It has been empirically set to the constant value of 3 CA for all the operating points.

Turbulence factor. The values of TC are adapted in order to fit with the experimental data of cylinder pressure. A large field of values has been scanned to obtain similar curves between the start of combustion angle and the end of combustion (burned mass fraction upper than 99%). The two strokes of compression and expansion are independent of the TC factor. The influence of this factor can be seen in Figure 4 (with

1T T optimalC C= =6.4.108, 2 1T TC C= + 7.5%, 3 1T TC C= + 15%). A 15% variation on TC corresponds to a 2~3 bar difference on the cylinder pressure.

Figure 4. Predicted pressure traces for several values of TC ,

for network gas + 5% C2H6 + 2% C3H8 and equivalence ratio 0.85φ =

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Predictive Model. In order to make the numerical model predictive, the turbulence factor TC must be given as a function of the operating conditions. TC is mainly influenced by the turbulence level and the nature of the air/fuel mixture (including the gas composition). The correlation for TC must then account for these specifications and several parameters have been selected. With regard to the level of turbulence, only the inlet manifold pressure collP varied in our experiments. For the air/fuel mixture, many indices have been tested .

Results obtained for the cylinder pressure can be seen in Figure 5 :

Figure 5. Comparison between experimentation and simulation

for network gas + 4% C3H8 and equivalence ratio 0.85φ =

In order to quantify the differences between optimized and correlated values of TC , two indicators have been studied (see Table 6) :

• the maximum pressure (bar) : Pmax, • the crank angle of maximum pressure (CA) : alpha.

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Table 6. Indicators studied for the cylinder pressure

Pmax exp (bar)

alpha exp (CA)

differences Pmax (%)

differences alpha (%)

net gas 77,8 14,0 0,6 -1,010% C2 80,0 14,0 -0,7 -1,04% C3 78,9 13,0 6,2 -2,0

5% C2 & 2% C3 78,1 15,0 -3,0 -3,0

net gas 66,4 18,0 6,5 -2,010% C2 69,2 15,0 4,9 0,04% C3 69,6 17,0 2,3 -2,0

5% C2 & 2% C3 70,3 15,0 -13,8 4,0

net gas 66,6 17,0 2,8 0,010% C2 69,9 16,0 5,5 -2,04% C3 69,1 16,0 4,5 -1,0

5% C2 & 2% C3 70,7 17,0 -5,1 -2,0

net gas 65,1 18,0 3,6 -1,010% C2 68,9 17,0 -0,3 -2,04% C3 67,7 18,0 1,2 -2,0

5% C2 & 2% C3 71,0 17,0 -3,5 -2,0

PHI=0,85

PHI=0,66

PHI=0,7

PHI=0,77

The maximum cylinder pressure is slightly overvalued (+0.7% on average), as for the

angle of Pmax, it is generally underevaluated (-1.2 CA). This panel of results is satisfactory and validates the model for this field of use. The prediction of the cylinder pressure cycle allows the computation of the output power and the thermal efficiencies in the next version of the model.

GAS COMPOSITION EFFECTS

The burn rate obtained from the combustion model, applied in the case of the six

cylinder engine, can be analyzed for each equivalence ratio in order to underline the effect of the gas composition (see Figure 6).

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Equivalence ratio = 0,66

0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

0,9

1

0 5 10 15 20 25 30 35 40 45 50 55

crank angle after spark timing

burn

ed fr

actio

n

+4% C3H8

+10% C2H6

network gas

+5% C2H6 +2% C3H8

Equivalence ratio = 0,70

0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

0,9

1

0 5 10 15 20 25 30 35 40 45 50 55

crank angle after spark timing

burn

ed fr

actio

n

+4% C3H8

+10% C2H6

network gas

+5% C2H6 +2% C3H8

Equivalence ratio = 0,77

0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

0,9

1

0 5 10 15 20 25 30 35 40 45 50 55

crank angle after spark timing

burn

ed fr

actio

n

+4% C3H8

+10% C2H6

network gas

+5% C2H6 +2% C3H8

Equivalence ratio = 0,85

0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

0,9

1

0 5 10 15 20 25 30 35 40 45 50 55

crank angle after spark timing

burn

ed fr

actio

n

+4% C3H8

+10% C2H6

network gas

+5% C2H6 +2% C3H8

Figure 6. Evolution of the burn rate

On the basis of this curves, four characteristic angles have been studied: CA005,

CA05, CA50, CA95 that respectively correspond to burn rates of 0.5%, 5 %, 50% and 95%.

The combustion duration is defined by CA95-CA05. The evolution is plotted on

Figure 7.

Figure 7. Evolution of the combustion duration

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An adjunction of hydrocarbons (ethane and propane which lower the methane

number) shorten the combustion duration, contrary to inert, like N2, which increases the methane number and the combustion duration. From Figure 7, a change of +10 MN points corresponds to +2CA on the combustion duration. The crank angles corresponding to the end of combustion (CA90, CA95 and CA99) are also increased with the methane number.

In order to compare the effects of the equivalence ratio and the gas composition, we

have studied (Figure 8) the CA50. This indicator is widely used by engine manufacturers to analyze the combustion process. One can notice that a same difference on CA50 (around to 2CA) can be obtained by either a variation of 10 MN points or a variation of 0.05-0.07 point on the equivalence ratio.

Figure 8. Evolution of the middle combustion angle

CONCLUSIONS The major conclusions to be drawn from the present study are : 1. A new approach to the modeling of natural gas effects on engine performance,

promoted by Gaz de France, has been presented. It is based on a predicative model using a phenomenological approach to reproduce the combustion process.

2. From preliminary academic works on a research single cylinder engine, a first

evolution of a model has been developed, tested and optimized. Gaz de France’s experimental facilities and theoretical knowledge have been utilized to extend the model towards industrial applications. The model has been successfully applied to simulate a city-bus gas engine operating with a wide variety of natural gases.

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3. A satisfactory reproduction of the cylinder pressure has been obtained for most of

the gases and for all operating points corresponding to the two gas engines : slight overestimation of the cylinder pressure and a short interval on the angle of the peak pressure.

4. The laminar flame speed was found to play a major role in the prediction of the

effects of the gas composition. More study must be carried out to evaluate more precisely the possibility to characterize the impact of the gas through the laminar flame speed or another index directly linked to it (such as the combustion potential).

5. The use of the model as an analysis tool enables to compare the relative influence

of Methane Number against equivalence ratio variations regarding the CA50: a variation of 10 MN points corresponds to a variation of 0.05-0.07 point on the equivalence ratio.

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Society of automotive engineers, 1994. [4] H.J. Schollmeyer, R. Wegener. Effect of gas properties on the operation of natural

gas vehicles. 2001 International Gas Research Conference, Amsterdam, 2001. [5] C.R. Ferguson. Internal Combustion Engines. Wiley, New York, 1986. [6] C. Olikara, G.L. Borman. A computer program for calculating properties of

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