Thermodynamic analysis of combustion and pollutants formation in a wood-gas spark-ignited heavy-duty engine

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 8 ( 2 0 1 3 ) 1 2 4 4 6 e1 2 4 6 4

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Thermodynamic analysis of combustion and pollutants formation in a wood-gas spark-ignited heavy-duty engine R.G. Papagiannakis a,*, T.C. Zannis b a

Propulsion & Thermal Systems Laboratory, Thermodynamic & Propulsion Systems Section, Aeronautical Sciences Department, Hellenic Air Force Academy (HAFA), Dekelia Air Force Base, Military Post 1010, Dekelia, Attiki, Greece b Laboratory of Marine Internal Combustion Engines, Naval Architecture & Marine Engineering Section, Hellenic Naval Academy (HNA), Hatzikiriakio, 185 39 Piraeus, Greece

article info

abstract

Article history:

Awareness of limitations of petroleum based liquid fuels as for example used in spark-ignited

Received 16 May 2013

heavy-duty engines for power generation, has led engineers to propose various solutions

Received in revised form

such as the use of alternative/renewable energy sources. Wood-gas is an alternative gaseous

26 June 2013

fuel generated from the gasification of wood, which could be used as a full supplement fuel in

Accepted 1 July 2013

conventional heavy-duty spark-ignited engines fuelled with natural gas. Previous related

Available online 9 August 2013

research studies have shown that the main disadvantage of the wood-gas combustion is its negative impact on brake engine efficiency compared to the normal natural gas operation,

Keywords:

while NO and CO emissions are also increased. Compression ratio and spark timing are some

Wood-gas

of the engine parameters, which influence significantly the combustion mechanism inside

Hydrogen

the combustion chamber of a wood-gas powered spark-ignited engine. In order to examine

Carbon monoxide

the effect of these parameters on the performance and exhaust emissions of a heavy-duty,

Nitric oxide

turbocharged, spark-ignited engine fuelled with wood-gas, a theoretical investigation is

Spark-ignition engine

conducted in this work by using a numerical simulation. The results concern engine performance characteristics, NO and CO emissions for various engine operating conditions (i.e. air to fuel excess ratios), by using a comprehensive two-zone phenomenological model. The predictive ability of the thermodynamic model was tested against experimental measurements, which were obtained from the operation of a multi-cylinder, four-stroke, turbocharged, spark-ignited engine fuelled with wood-gas fuel at various loads. The experimental results are found to be in good agreement with the respective computed ones obtained from the simulation model. The main objective of the comparative assessment shown in the present work is to record and comparatively evaluate the relative impact of each one of the above mentioned parameters (compression ratio and spark timing) on the engine performance characteristics and emitted pollutants. Furthermore, an effort is made to determine the optimum combinations between these parameters, since at high engine load conditions their simultaneous increase may lead in undesirable results concerning the engine performance characteristics. The conclusions from the present investigation are valuable for the use of wood-gas as a full supplement energy source in conventional, natural gas fuelled, heavy-duty, spark-ignited engines used for electric power generation. Copyright ª 2013, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights reserved.

* Corresponding author. Tel.: þ30 210 8193806; fax: þ30 210 807 4606. E-mail addresses: [email protected], [email protected] (R.G. Papagiannakis), [email protected] (T.C. Zannis). 0360-3199/$ e see front matter Copyright ª 2013, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.ijhydene.2013.07.007

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Nomenclature A c h k L m _ m N p Pe Q R Re s T t U u0 V

2

surface, m specific heat capacity under constant pressure, J kg1 K1 specific enthalpy, J kg1 thermal conductivity, W m1 K1 Taylor micro-scale, m mass, kg mass flow rate, kg s1 engine speed, rpm pressure, Pa brake power output, W heat, J gas constant, J kg1 K1 (also radius, m) Reynolds number piston stroke, m temperature, K time, s internal energy, J turbulence intensity, m s1 volume, m3

Greek symbols Da Crank angle step, degrees volumetric efficiency, % hV exhaust gas temperature,  C qg l air to fuel excess ratio r density, kg m3 s induction or characteristic time, s 4 fuel/air equivalence ratio Subscripts b burning cyl cylinder F end of flame propagation f fuel, flame g gas [ laminar

1.

Introduction

With the increasing public interest in energy supply and the environment, attention has focused on the development of ecological and efficient combustion technologies. Nowadays, around 80% of the world primary energy is satisfied by fossil fuels [1,2]. It has become common belief that today’s main resources of energy, such as the conventional petroleum based liquid fuels, will become scarce within the next generation. Awareness of limitations of fossil fuels reserves and the fact that burning of fossil fuels has a major contribution to the greenhouse gases emission, has led to a growing interest in the use of bio-energy and other renewable energy sources. Biomass offers flexibility of fuel supply due to the range and diversity of fuels that can be produced. Moreover, it can be converted into liquid (bio-oil) or gaseous (biogas) fuels, which can be used on energy systems increasing the energy available for economic development without contributing to the

mix p st u w

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mixture flame kernel stoichiometric unburned wall

Superscripts comb combustion ref reference tot total Abbreviations AFR air to fuel ratio (by mass) ANSA after normal spark advance BMEP brake mean effective pressure BSFC brake specific fuel consumption BTDC before top dead centre CA crank angles methane CH4 CO carbon monoxide carbon dioxide CO2 CR compression ratio deg degrees HD heavy-duty IC internal combustion IVC inlet valve closure NCR normal compression ratio NEO normal engine operating conditions NO nitric oxide nitrogen oxides NOx NSA normal spark advance OH hydroxyl ROHR rate of heat release rpm revolutions per minute SI spark-ignition sulfur dioxide SO2 SOI start of ignition T/C turbocharged TDC top dead centre

greenhouse effect, while most of the biomass fuels are characterized by friendly environmental attributes such as low levels of sulfur and NOx emissions [3,4]. Many internal combustion engines, usually converted from commercial compression-ignition [5e8] or sparkignition engines [9e12], have been fuelled with biomass fuels (bio-fuels) for use in electric power generation [13,14]. Landfill gas [15,16], sewage gas [17], and syn-gas [18e21] are alternative gaseous bio-fuels that can be used as full supplement energy sources in commercial compression- or sparkignition engines used for electric power generation. Landfill gas and sewage gas are both the by-products of anaerobic decomposition of organic matter and are primarily composed of methane (w50e70%) and carbon dioxide (w50e30%). Landfill gas is produced in sanitary landfills, whereas the digester gas at sewage treatment plants [15,16]. On the other hand, syn-gas is the product of the partial combustion (a thermo-chemical process, gasification and pyrolysis) of

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biomass in a gasifier (i.e. device in which solid biomass materials, such as wood, agricultural and agro-industrial wastes are converted to gases) [22e25], with its main components being: H2 (w20%), CO (w20%), N2, CO2 and CH4 [3,26e28]. Among the various types of syn-gases, wood-gas [29,30] offers a promising opportunity for sustainable development in the electricity generation sector, especially in areas where the price of petroleum based fuels is high, or where supplies are unreliable. Furthermore, using wood for electrical power generation produces zero net gain of carbon dioxide and other greenhouse gases, essentially eliminates sulfur dioxide (SO2) emission, increases the energy security by using indigenous fuel, contributes to the industrial and forest economy, and improves the environment by using wastes and residues [31,32]. From the thermodynamic point of view, combining heat and electricity production using wood gasification leads to major advantage [33e35]. One of the key components of woodgas is hydrogen, which has very clean burning characteristics, high flame propagation speed, and wide flammability limits. Moreover, it has a laminar combustion speed roughly eight times that of natural gas [38e40], and hence increased hydrogen content in a gaseous fuel that reduces the combustion duration and thereby improves the quality of fuel utilization. A major point of interest is that the presence of hydrogen in the wood-gas extends the lean limit of gas operation without entering the lean misfire region. Lean mixture combustion has great potential for attaining higher thermal efficiency and lower emissions. In particular, lean mixture combustion can produce low (and even extremely low) NOx emission levels with only a slight increase in hydrocarbon emissions [36e40]. On the other hand, wood-gas is a low energy-density fuel, and the power degradation is extensive compared to high energydensity fuels such as natural gas [36e40]. A number of internal combustion (IC) engines, mainly with spark-ignition (SI) system, have recently been developed, which were fuelled with alternative gaseous fuels such as biogas, natural gas and hydrogen [41e56]. A substantial research (experimental and theoretical) on biogas powered spark-ignited engines has concentrated on the extent of syngas-fuelling [36e41] and its effect on performance and emissions [42e47]. Despite the extensive literature concerning the economic feasibility and environmental impact of syn-gas powered SI engines, just very few studies can be found for internal combustion engines fuelled with wood-gas. Moreover, the information is very general and does not contain any details regarding the impact of wood-gas on engine performance parameters and pollutant emissions. In addition, there is still more to be learnt about the relative impact of each one of some critical engine parameters, such as compression ratio, spark advance and boost pressure, on the performance characteristics and environmental behaviour of a heavy-duty (HD), spark-ignited engine, where the conventional gaseous fuel (i.e. natural gas) is fully substituted by wood-gas. These are the primary objectives of a research project, funded by the European Committee [57], in which the authors were participated in the past. The authors contribution on the aforementioned project was concentrated on the development of a simulation model for the prediction of performance characteristics and pollutant emissions and, moreover, on the use of the validated model to examine the effect of various engine

parameters (compression ratio, spark timing, etc.) on the performance characteristics and environmental behaviour of a heavy-duty, spark-ignited, natural gas engine fuelled with wood-gas, instead of the conventional gaseous fuel [33,57]. Thus, the primary objective of the present work is to examine the effect of compression ratio and spark timing on the performance and exhaust emissions of a multi-cylinder, spark-ignited engine, operating with wood-gas as the fuel, by using a two-zone phenomenological combustion model. For the current investigation, the simulation model has been properly modified and improved substantially to describe more accurately the complicated wood-gas combustion process in a spark-ignition engine environment. The simulation results comprised of the in-cylinder pressure and heat release rate histories, the maximum combustion pressure, the duration of combustion, the brake power and the brake specific fuel consumption, as well as the calculated brake specific NO and CO emissions for various engine loading conditions (i.e. various air to fuel excess ratios), at 1500 rpm engine speed. The model predictions corresponding to engine operation with normal compression ratio (i.e. NCR) and normal spark advance (i.e. NSA) are validated against corresponding experimental values obtained from a multi-cylinder, turbocharged (T/C), water-cooled, spark-ignited engine, which is fuelled with wood-gas instead of natural gas at various engine loading points. The engine was supplied from a wood-gas producer, inside which the solid woody material is converted to gaseous fuel through a thermo-chemical process. From the comparison of computed and experimental findings, it is shown that the simulation model predicts adequately the engine performance and pollutant emissions trends with engine load under normal compression ratio (NCR) and normal spark advance (NSA) operating modes. Taking into account the character of the specific work, where a comparative assessment is conducted by using the simulation results concerning the relative impacts of the examined parameters on engine performance characteristics and emitted pollutants, it is revealed that the developed model could be used safely to examine the effect of each one of the examined engine parameters on engine performance and pollutant emissions. In the present work, examining the theoretical results, important information is derived revealing the applicability of each one of the examined techniques on an existing heavyduty, SI engine operating under wood-gas fuel mode. Moreover, a comparative assessment has been conducted, by using the simulation results, concerning the relative impact of each one of the examined strategy on engine performance characteristics and emitted pollutants. From the theoretical findings, important information is derived revealing both the applicability of each one of the examined techniques on an existing spark-ignition engine operating with wood-gas, and also the effect of each technique on engine performance and pollutant emissions. Consequently, the information derived from the present work is extremely valuable regarding the implementation of one of the two strategies for improving the efficiency and the environmental behaviour of an existing heavy-duty spark-ignited engine fuelled with wood-gas, without detrimental repercussions on the constructional endurance of the engine.

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2. Description of the engine simulation model 2.1.

Physical overview

The engine simulation model used in the present study is a phenomenological two-zone one, which has been successfully used in the past to simulate the operational and environmental performance of SI engines burning gaseous fuels [34,57]. The model examines the part of engine cycle corresponding to the time interval that both inlet and exhaust valves are closed (closed cycle). After IVC and during the compression stroke, the in-cylinder charge is treated as a uniform mixture of air and gaseous fuel (i.e. wood-gas in this case), which has been premixed during the intake stroke. Uniformity of temperature, pressure and composition is assumed for the cylinder mixture at each crank angle during the compression stroke (single-zone assumption). The combustion process is treated as a two-stage procedure comprised of the ignition process and the stable flame propagation process [49,50]. As shown in Fig. 1(a), the ignition process consists of the formulation of the flame kernel due to spark electric discharge and its unsteady flame propagation. Commencement of combustion is considered when a finite volume of the flame kernel exceeds 0.001 times the total incylinder volume [58,59]. As can be seen in Fig. 1(b), upon combustion initiation, a two-zone phenomenological approach is adopted for the treatment of the combustion process. According to this concept, one zone consists of air and gaseous fuel mixture and is called “unburned” and the other zone named “burning” consists of combustion generated gases and potentially excess air depending on the air to fuel ratio. In each zone prediction of a uniform temperature is performed at each crank angle by employing the first law of thermodynamics and the ideal gas law and by assuming uniform composition. Heat exchange between zones is neglected. The two zones are separated by the flame kernel front, which is assumed to have a spherical shape.

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Since the laminar flame thickness under engine conditions is infinitesimal [60e64], the flame front is treated as negligibly thin. The amount of gaseous fuel and air entraining the burning zone is determined by the volume change rate of the spreading spherical flame front. The speed of the flame front propagation into the unburned zone is predicted by taking into consideration the turbulent flame propagation mechanism. Hence, the fuel burning rate inside the propagating flame kernel depends on the turbulent flame speed.

2.2.

Mathematical treatment

2.2.1.

Derivation of conservation and state equations

The combustion simulation model is based on the application of mass and energy conservation and ideal gas laws in each zone separately and for the entire cylinder charge. The implementation of first law of thermodynamics (energy conservation) and ideal gas state equation in each zone at each crank angle in the following form [62,64]: dUj ¼ dQ j  pdVj þ pVj ¼ mj Rj T j

X

 dmj hj ð j ¼ u; bÞ

ð j ¼ u; bÞ

(1) (2)

results in the derivation of two first-order differential equations describing the rate change of unburned zone Tu and burning zone Tb temperatures. The application of the same principles leads to the formulation of a third first-order differential equation describing the rate change of cylinder pressure p. The system of the three aforesaid ordinary differential equations is numerically solved using a non-stiff predictorecorrector method. Eq. (1) contains total values of internal energy and enthalpy, which means that enthalpies of formation of all species are included and hence, the heat release is taken into account implicitly [61]. In addition, the mass exchange between unburned and burning zone at each time step dt is: dmu ¼ dmb

Fig. 1 e Schematic views of the (a) flame kernel formation and (b) spark-ignited two-zone combustion concept.

(3)

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where dmb is the mass entrained into the burning zone and dmu the mass reduction of the unburned zone.

2.2.2.

Ignition delay

An ignition delay period is considered between the spark ignition and the initiation of combustion, which takes place when the volume of the generated flame kernel becomes higher than 0.1% of the total cylinder volume. At the nominal ignition time, the mixture is considered to burn adiabatically under constant volume inside the flame kernel. For this reason, an internal energy balance is performed for the mixture inside the kernel. The initial estimation of the flame kernel temperature Tp is made using the following relation [61]: Tp ¼ Tu þ 25004 for 4  1 Tp ¼ Tu þ 25004  700ð4  1Þ

for 4 > 1

(4)

where 4 is the mixture fuel/air equivalence ratio. Determination of the final value of temperature Tp is conducted through an iterative procedure in order the specific internal energy of the products becomes equal to the pertinent value of. The volume of the flame kernel is calculated as: Vp ¼ ð2=3ÞpR3p

(5)

where Rp is the radius of the flame front from the spark plug, which is obtained from: Rp ¼ St

Da 360$Ns

(6)

where St is the turbulent flame speed, Da is the time interval in crank angle degrees and Ns is the engine speed in rotations per second. For each time step after the spark-plug discharge, if the volume of the flame kernel is less than the 0.1% of the total cylinder volume, the mixture is considered not to have been burned yet and the compression continues to the next crank angle step. Then, the ignition delay is calculated from the following expressions taking into consideration the flame kernel volume, the engine speed and the turbulent flame speed:  1=3 Rp ¼ 3Vp =2p

and Da ¼

Rp 360$Ns St

(7)

When the kernel volume becomes higher than 0.1% of the total cylinder volume, the combustion phase initiates and the cylinder chamber is subdivided into a burned (suffix ‘b’) and an unburned (suffix ‘u’) zones, each of them having its own temperature and pressure values. The pressure jump arisen during this stage is attributed to the expansion of the flame kernel as a result of the spark ignition. Thus, at this step, variation between the pressures of the unburned and burning zone occurs. However, as already mentioned, in-cylinder pressure uniformity in space is adopted by the simulation model. Thus, an equilibration of the zone pressures is considered, which is treated as an adiabatic constant volume process.

2.2.3.

Definition of the burning zone

Assuming that the geometry of the flame front formed inside the cylinder of an SI engine has a spherical shape [62], the determination of the instantaneous flame geometry after

spark discharge is achieved by employing a model developed by Annand [60]. According to this model, the net outer surface area of the spherical burning zone Af is calculated by employing the instantaneous flame radius from the spark plug and the chamber height in a Simpson’s integration scheme. As the spherical burning zone spreads into the unburned one, a homogeneous mixture of air and gaseous fuel is entering into the burning zone. The mass entrainment rate into the burning zone is calculated from the volume change rate of the zone as follows: dmb ¼ ru dVbtot ¼ ru Af St

2.2.4.

(8)

Flame propagation mechanism

Knowing that the combustion process in SI engines is turbulent it is logical to assume that the combustion propagation of gaseous fueleair mixture inside cylinder will depend heavily on the turbulent flame velocity St, which is defined as follows: St ¼ S[ þ u0

(9) 0

where S[ is the laminar flame velocity and u is the turbulence intensity, which is given by the following relation [59,65e67]:  1=3 u0 ¼ C7 $ðN$s=30Þ$ ru =rSOI u

(10)

is the density of the unburned where C7 is a constant and rSOI u mixture at the start of ignition. The laminar flame speed is given by the following correlation:  S[ ¼ Sref [ $

Tu Tref

#bt at " p $ pref

(11)

where Tref ¼ 298 K, pref ¼ 1 atm, while Sref [ , at, bt are constants depending on the local fuel/air equivalence ratio 4. These constants can be calculated as follows: ¼ C1 þ C2 ð4  4m Þ2 Sref [

at ¼ C3  C4 ð4  1Þ

bt ¼ C5 þ C6 ð4  1Þ

(12)

where C1.C6 are constants and 4m is the fuel/air equivalence ratio at which Sref [ has the maximum value. Hence, as observed, the laminar flame velocity and through this, the turbulent flame speed depends on the fuel/air equivalence ratio, the temperature of the unburned gas and the cylinder pressure.

2.2.5.

Combustion

According to a well-known turbulent entrainment model [58,62,65], the combustion rate of the gaseous fuel/air mixture in the burning zone is calculated using the characteristic reaction time sb as follows: ¼ dmcomb b

mb  mcomb b sb

(13)

where mb is the mass entrained by the flame front, and mcomb b is the mass consumed during the combustion period. The characteristic reaction time is calculated by taking into account the Taylor micro-scale L, as follows: sb ¼ L=S[ The Taylor micro-scale is obtained from the formula:

(14)

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1=3   1=3 SOI L ¼ C8 $ Vcyl $ rSOI u =ru

(15)

SOI is the instantaneous cylinder where C8 is a constant and Vcyl volume at the start of ignition. As the combustion process continues the volume of products increases at a rate determined by the turbulent flame velocity whereas, the volume of the unburned zone decreases by the same amount respectively. Theoretically, combustion should be terminated when Vb  0. However, according to the combustion model, there is a small amount of mixture that has not been consumed yet when Vb  0. Hence, the combustion rate of the remaining unconsumed mixture during the final burning phase is calculated as follows:

¼ mb;F $exp dmcomb b

 tF  t sb

where mb,F is the unburned mixture inside the burning zone and suffix F denotes the conditions at the end of the flame propagation.

2.2.6.

Knock

Engine knock is a spark-ignited engine phenomenon, where the compressed gaseous fueleair mixture inside the cylinder ignites before it has been reached by the propagating flame front [64]. A rather simplified but efficient approach to model knock in SI engines is to use the knock integral and apply it to the unburned zone of the mixture. Thus, in the proposed model, the autoignition of the mixture occurs when [62,64]: Ztk 0

with the cylinder wall surfaces, a bulk average temperature of both zones is used as follows [34,69,70]: Tg ¼

mu cu Tu þ mb cb Tb mu cu þ mb cb

 dQ j dQ mj cj Tj ¼ ð j ¼ u; bÞ dt mu cu Tu þ mb cb Tb dt

(17)

where t is the elapsed time from the start of the mixture compression process, tk is the time of autoignition and s is the induction time at the instantaneous temperature and pressure conditions. In the simulation model, the induction time is modelled by an Arrhenius type equation, which relates it to the cylinder pressure p and the temperature of the unburned zone Tu as follows [68]: s¼

Ckn 2 $expð3800=T Þ Ckn u 1 $p

(18)

kn where Ckn 1 and C2 are experimental constants.

2.2.7.

Heat transfer

The heat exchange between in-cylinder gaseous mixture and cylinder walls is primarily ascribed to the convection and after combustion initiation, to the radiation mechanisms. In the present simulation model, the heat exchange rate between incylinder working medium and cylinder walls is calculated using the well-known formula developed by Annand [63], which incorporates the two aforesaid heat transfer mechanisms:     dQ Reb  ¼ A ak Tg  Tw þ c T4g  T4w dt D

(19)

where a, b and c are constants and k is the thermal conductivity of the in-cylinder gaseous mixture. During the combustion phase, since the burning zone is not fully in contact

(21)

Equilibrium chemistry of combustion products

As known, under high temperatures occurring in IC engines, the products of the perfect combustion are heavily dissociated producing more species, which are considered to be in chemical equilibrium. According to Heywood [62], chemical equilibrium concept for combustion products is considered valid for the expansion stroke except from the exhaust process, which however, is not examined by the present simulation model. The chemical equilibrium species considered are: H2O, H2, OH, H, N2, NO, N, CO2, CO, O2 and O. In the present model, the molar concentration of the aforementioned 11 species is calculated in each crank angle by employing an iterative procedure proposed by Vickland et al. [66] using the products of the perfect combustion as initial estimates.

2.2.9. 1 dt ¼ 1 s

(20)

where cu and cb is the specific heat capacities of the unburned and the burning mixture respectively. Thus, the total heat exchange rate dQ/dt is distributed between the two zones according to their mass, temperature and specific heat capacity as follows:

2.2.8. (16)

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Formation of nitric oxide and carbon monoxide

The formation of nitric oxide (NO) is controlled by chemical kinetics based on the reactions of the well-known extended Zeldovich mechanism [67]. According to this mechanism, the following reactions, with their related forward reaction rate constants, are considered in the present model: NþNO4N2 þO k1f ðm3 =kmolsÞ ¼ 1:61010 NþO2 4NOþO k2f ðm3 =kmolsÞ ¼ 6:41010 Tb expð3125=Tb Þ NþOH4NOþH k3f ðm3 =kmolsÞ ¼ 4:21010 (22) As reported [62,65], the carbon monoxide (CO) formed by the combustion process is oxidized to carbon dioxide at a rate that is relatively slow compared to the carbon monoxide formation rate. The two kinetically-controlled reactions, with their related forward reaction rate constants, are the following ones [34,65,69,70]: COþHO4CO2 þH k1f ðm3 =kmolsÞ ¼ 6:761010 expðTb =1102Þ CO2 þO4COþO2 k2f ðm3 =kmolsÞ ¼ 2:511010 expð22;041=Tb Þ (23)

2.3. Model modifications conducted to simulate combustion of wood-gas fuel Having given by the engine manufacturer the actual chemical composition of the wood-gas on a volume basis, it was calculated the number of carbon, hydrogen and oxygen atoms of the fuel, which, then, was used to derive a representative chemical formula for the wood-gas fuel. In addition, the known net heating value of the wood-gas fuel was used to estimate its enthalpy of formation through an

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iterative procedure. Hence, the influence of wood-gas heating value on the prediction of burning zone temperature, cylinder pressure and heat release rate was taken into account through the involvement of gaseous fuel enthalpy of formation in the first law of thermodynamics. Moreover, the derivation of the wood-gas carbon/hydrogen/oxygen atom analogy allowed the simulation of combustion chemistry of the wood-gas/air mixture and the determination of the chemical equilibrium species, NO and CO formation rates as previously described.

3.

Model validation

To verify the ability of the model to predict, apart from overall engine performance, the main characteristics of wood-gas combustion process in a spark-ignited environment, an extended theoretical and experimental investigation has been conducted in the past [33,34,57]. Specifically, to calibrate and also to evaluate the predictive ability of the proposed model, the results are used from an experimental investigation, conducted by the engine manufacturer, on a “GE Jenbacher J320GS” multi-cylinder, turbocharged, sparkignited engine fuelled with wood-gas. These results are also used as a basis to evaluate the theoretical findings concerning the effect of compression ratio and spark advance, on engine performance and pollutant emissions characteristics of a multi-cylinder, spark-ignited, heavy-duty engine “GE Jenbacher J316GS” fuelled also with wood-gas. The basic data related to the simulated engines are presented in Table 1. Observing the data given in Table 1, it is revealed that the engine “J316GS” has the same operational and geometrical data with the engine “J320GS”, with the only difference between them being the number of cylinders. Moreover, it must be stated here that for both types of engine, the normal (constant) speed is kept constant at 1500 rpm, since both engines are used as electric power generators (actually in a co-generation mode). The wood-gas used in the present investigation is a mixture of methane (CH4), carbon monoxide (CO), hydrogen (H2), nitrogen (N2) and carbon dioxide (CO2). The actual chemical composition of the wood-gas was provided by the engine manufacturer and is given in Table 2. It must be stated here that the net calorific value and the stoichiometric air to fuel ratio, which are also given in Table 2, have been calculated by taking into account the known composition of the fuel. Observing Table 2, it is also shown that the type of the wood-gas used consists of about 50% combustible gases, mainly carbon monoxide and hydrogen, while the rest is noncombustible gases. As already mentioned, all experimental measurements involved in this study were obtained from the engine manufacturer [33] and they refer to the operation of the “J320GS” engine fuelled with an alternative gaseous fuel (i.e. wood-gas), with normal value of spark timing (i.e. normal spark advance e NSA) and normal value of compression ratio (i.e. NCR). Measurements were taken at four different engine operating points corresponding to 40%, 65%, 85% and 100% of full engine load, at 1500 rpm engine speed. For each engine operating point, the following engine parameters were measured:

Table 1 e Basic geometrical and operational data of the test engines. Engine model Intake and combustion type Number of cylinders Arrangement of cylinders Bore Stroke Compression ratio Engine displacement volume Connecting rod length Inlet valve diameter Outlet valve diameter Mean piston speed

J320GS

J316GS

SI e Turbocharged

SI e Turbocharged

20 V 700

16 V 700

0.135 m 0.170 m 11:1 48.7 l

0.135 m 0.170 m 11:1 48.7 l

0.32 m 0.055 m 0.051 m 8.5 m/s

0.32 m 0.055 m 0.051 m 8.5 m/s

 Air to fuel excess ratio (i.e. lambda) (l), which is defined as actual air to fuel ratio divided by its stoichiometric value.  Inlet charge conditions (i.e. pressure ( pIVC) and temperature (TIVC)).  Spark advance (NSA).  Volumetric efficiency (hv).  Cylinder pressure diagram.  Brake power output (Pe).  Brake specific fuel consumption (BSFC).  Exhaust gas temperature (qg). _ f ).  Fuel mass flow rate (m  Specific NO and CO emissions. The engine manufacturer provided [33] the aforementioned experimental data under a European research program [57], which are summarized in Table 3. For validating the ability of the developed two-zone model to predict engine performance, a comparison is given in the present work between measured and calculated cylinder pressure diagrams. Fig. 2 shows the comparison between computed and experimental pressure traces under “woodgas” operating mode, for an engine speed of 1500 rpm and at 100% of full engine load. As shown, there is a good agreement between computed and measured traces, which is promising for the use of the proposed model to predict the performance characteristics of a spark-ignited engine fuelled with wood-gas. The model has been calibrated using the available experimental results taken at a speed of 1500 rpm and for 100% of full engine load, which are depicted in Table 3. It must be emphasized here that the values of the present model’s constants are held constant for the entire test cases considered in the present study.

Table 2 e Basic characteristics of the fuel used. CH4 (%v/v)

CO (%v/v)

0.5 21 Net calorific value : AFRst:

H2 (%v/v) 24

N2 (%v/v)

CO2 (%v/v)

54.5 19.30 (MJ/kg) 5.948 (kg/kg)

0

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Table 3 e Experimental data for the operation of engine “J320GS” with wood-gas. Load (%) PIVC (bar) TIVC (K) L () mf (kg/h) 40 65 85 100

1.45 2.02 2.35 2.70

365 362 363 366

1.56 1.68 1.76 1.83

247 389 483 552

NSA Pe (kW) qg (0C) BSFC (g/kW h) NO (g/kW h) CO (g/kW h) (deg CA BTDC) 35 30 27 25

Moreover, the experimental heat release rate was obtained from the analysis of the corresponding experimental cylinder pressure trace by using the methodology already presented in Refs. [69,70]. The comparison between experimental and calculated values of some of the most important engine performance characteristics, i.e. brake power output and brake mean effective pressure (BMEP) as well as brake specific fuel consumption (BSFC) and brake thermal efficiency, at various engine loading points, is given in Fig. 3(a)e(d). Observing the results given in Fig. 3(a) and (b), a very good coincidence is shown between experimental and calculated values, thus revealing the ability of the specific model to predict adequately the cylinder pressure diagram under various engine loading conditions. Observing Fig. 3(c) and (d), it is shown that for all engine loading points examined the proposed model predicts with relatively good accuracy the experimental values. Specifically, examining the BSFC variation with engine load, it is observed that the increase of engine load results in a decrease of the brake specific fuel consumption as the combustion of the gaseous fuel becomes more efficient. It must be stated here that the experimental brake specific fuel consumption is estimated from the measured brake power output and the measured mass flow rate of the gaseous fuel. In addition, it must be underlined that

310 518 670 790

796 750 720 698

21.7 15.6 12.08 8.54

206 127 108 94

the unusually high BSFC value is the result of the very low heating value of the gaseous fuel used. The variation of measured and calculated specific NO and CO concentration with engine load at 1500 rpm engine speed is given in Fig. 4(a) and (b). Examining these figures, it is observed that the model predicts with adequate accuracy the trends of specific NO and CO emissions with engine load. The predicted absolute values are close to the respective measured ones, except for the case of low engine load conditions. Moreover, for each loading point examined, the computed values are underestimated compared to the experimental ones. Nonetheless this is usual for a two-zone model, as it under-predicts the burning zone temperature, a fact seriously affecting the formation mechanisms of both NO and CO emissions. Consequently, from the comparison between the experimental and theoretical results obtained from the “J320GS” engine, it is revealed that despite the differences observed between the measured and calculated absolute values, which are primarily attributed to the inherent limitations of the phenomenological two-zone model, the simulation model manages to predict with adequate accuracy the trend of engine performance characteristics and pollutant emissions with the engine load variation (i.e. change of inlet charge conditions and air to fuel excess ratio). Hence, the specific phenomenological model can be used to perform a parametric study concerning the effect of the compression ratio and spark timing on the performance and pollutant emissions of the engine “J316GS”, for both of which the operating and geometrical characteristics are completely similar with those of the engine “J320GS”.

4.

Fig. 2 e Comparison between experimental and computed cylinder pressure traces of the “J320GS” engine operating with wood-gas, at 100% of full engine load and 1500 rpm engine speed.

568 609 610 605

Test cases examined

In the present work, an effort is made to use the two-zone combustion model for theoretically investigating the relative impact of compression ratio and spark timing, on the performance characteristics and exhaust emissions of a sixteencylinder, turbocharged, heavy-duty, spark-ignited engine “J316GS” fuelled with wood-gas. Thus, at two different engine loads examined (i.e. two different air to fuel excess ratios), corresponding to 65% and 100% of full engine load and at 1500 rpm engine speed, it was initially investigated the effect of compression ratio on the performance characteristics and pollutant emissions of a heavy-duty, spark-ignited engine, running under normal spark advance (i.e. NSA) operating conditions. Thus, under NSA operating mode, the compression ratio was increased by 9%, 18% and 27% with respect to the normal compression ratio (i.e. NCR ¼ 11). Furthermore, it

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Fig. 3 e Comparison between calculated and experimental values of brake power output (a), brake mean effective pressure (b), brake specific fuel consumption (c), and brake thermal efficiency (d), of the “J320GS” engine operating with wood-gas at 1500 rpm and at various loads.

was examined the effect of ignition timing on the performance characteristics and pollutant emissions of the aforementioned engine, running under normal compression ratio (i.e. NCR). Thus, under NCR operating mode, spark timing was changed by 2 crank angle after the normal start of ignition

(i.e. 2 CA ANSA) and by 2 and 4 crank angle before normal start of ignition (i.e. 2 and 4 CA ANSA). The aforementioned CR and spark advance values were selected after a suggestion made by engine manufacturer [33] under a European research program. The purpose of this program was the

Fig. 4 e Comparison between calculated and experimental values of specific NO emissions (a), and specific CO emissions (b), emitted by the “J320GS” engine operating with wood-gas at 1500 rpm and at various loads.

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utilization of gaseous fuel produced by wood gasification in HD SI engines, which are used for decentralized electric power generation. Hence, part of this investigation was the engine mapping in terms of endurance, efficiency and gaseous emissions under compression ratios and spark advances higher than the normal ones of the production SI engine used in the program. It must be emphasized here that for each engine loading point examined (i.e. 65% and 100% of full engine load), the test cases corresponding to the engine operation with normal spark advance (i.e. NSA) and normal compression ratio (i.e. NCR ¼ 11), are referred to as “Normal Engine Operating” points (i.e. NEO). Some of the experimental data given in Table 3 (i.e. inlet charge conditions ( pIVC, TIVC), air to fuel excess ratio (l), normal spark advance (NSA) and volumetric efficiency (hv)), were used as input values to the simulation program. Furthermore, for each value of (l), the total amount of the inducted mixture has to be kept constant since the engine runs at constant engine speed, i.e. 1500 rpm. Thus, for each normal engine operating point examined, the gaseous fuel supply is obtained from the formula: _ fuel ¼ m

_ mix;IVC m l$AFRfuel st þ 1

(24)

_ mix;IVC ) is the mass flow rate of the gaseous mixture where (m inside the combustion chamber at the initiation of compression stroke, and (AFRfuel st ) is the stoichiometric air to fuel ratio that is calculated from the known composition of the fuel examined.

5.

Results and discussion

In this section, the predictive capabilities of the phenomenological model are explored. The predicted effects of compression ratio and of spark timing are examined, on some basic performance characteristics and pollutant emissions of a turbocharged, heavy-duty, spark-ignited engine “J316GS” fuelled with wood-gas, for two engine operating points corresponding to 65% and 100% of full engine load, at 1500 rpm engine speed.

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5.1. Effect of compression ratio and spark advance on engine performance characteristics 5.1.1. Effect on cylinder pressure and net heat release rate histories Fig. 5(a) and (b) provides the predicted cylinder pressure and net heat release traces for two compression ratios (i.e. CR ¼ 11 and CR ¼ 13), at 65% and 100% of full engine load and 1500 rpm engine speed, with normal value of spark timing (NSA). The cylinder pressure and net heat release rate traces estimated with CR ¼ 11 correspond to normal engine operating mode (NEO). Observing Fig. 5(a) and (b), it can be seen that the compression ratio affects the cylinder pressure history. Thus, for the same engine load conditions, during the compression stroke, the cylinder pressure under NEO operating mode diverges from the respective values observed with increased compression ratio. The difference becomes more evident during the last stages of compression stroke. After the initiation of combustion and for both loads examined, the rate of cylinder pressure rise with increased compression ratio during the initial stage of the combustion process becomes higher, while the peak of the cylinder pressure occurs slightly earlier compared to the respective values observed under NEO condition. This could be attributed to the slightly sharper gaseous fuel burning rate, since the increase of the cylinder charge temperature results in higher flame propagation velocity. As far as the effect of compression ratio on the heat release rate curve is concerned, it is revealed that compression ratio affects very slightly the burning rate of the alternative gaseous fuel. Examining the heat release rate curves shown in Fig. 5(a) and (b), it is revealed that, for all test cases examined herein, the shape of the heat release rates is almost the same with a very low acceleration of the initiation of combustion, without prolonging or shortening seriously the duration of combustion. Thus, the phasing of the heat release traces with increased compression ratios shifts more into the compression stroke as compared to the respective one observed under NEO operating mode. The latter could be attributed to the slight improvement of the gaseous fuel combustion quality, especially during the initial stages of the combustion process,

Fig. 5 e Calculated cylinder pressure and heat release traces for various compression ratios with normal spark timing at 1500 rpm and 65% load (a), and 1500 rpm and 100% load (b).

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which is caused primarily by the slight improvement of the cylinder charge temperature that contributes to the existence and the fast spread of the flame front surrounding the burning zone. Fig. 6(a) and (b) provides the predicted cylinder pressure and heat release traces for various spark timings, with normal compression ratio (NCR) at 65% and 100% of full engine load and 1500 rpm engine speed. The cylinder pressure and heat release traces estimated with normal value of spark timing (NSA) correspond to normal engine operating mode (NEO). At 65% of full engine load conditions, normal spark timing is defined to 145 CA while at 100% load the respective normal value is defined to 154 CA. Observing these figures it is revealed that for both loads, ignition timing affects also cylinder pressure history, especially during the initial stages of the combustion process. Thus, advancing spark timing increases the rate of the cylinder pressure rise during the initial stage of the combustion process. This is because the gaseous fuel combustion is accelerated, revealing thus the increase of the cylinder charge temperature, which in association with the higher energy release rates, contributes to the increase of the rate of the cylinder pressure rise observed during the specific phase. As far as the effect of spark timing on the heat release rate is concerned, it is revealed that the ignition timing does not affect seriously the shape of the heat release rate without prolonging seriously the duration of combustion. However at the same time, by advancing the ignition timing combustion initiates earlier while the rate of heat release rise during the initial stages of combustion becomes slightly higher. This is the result of the increased fuel mass burned before top dead centre (TDC). The latter contributes significantly to the existence and the fast spread of the flame front surrounding the burning zone. On the other hand, by retarding ignition timing, the rate of heat release rise observed during the initial stages of combustion process becomes slightly lower compared to the one observed at NSA operating mode. This is the result of the slight deterioration of the gaseous fuel combustion quality, which is owed primarily to the fact that the initiation of combustion is delayed. Thus, the charge temperature becomes lower, which results to higher amounts of gaseous fuel burned after the TDC position.

5.1.2.

Effect on maximum cylinder pressure

Since the maximum cylinder pressure is a critical parameter affecting the mechanical strength of engine structure, the study of the effect of both engine parameters examined herein (i.e. compression ratio and spark advance) is of particular interest. Consequently, in this section theoretical predictions are presented about the effect of each parameter on maximum cylinder pressure, as well as theoretical predictions for the combined effect of the two parameters on the maximum cylinder pressure. Thus, at each engine loading point examined, the variation of the calculated maximum cylinder pressure with compression ratio and spark timing is given in Fig. 7(a) and (b), respectively. Moreover, for both engine loading points, the variation of the maximum cylinder pressure versus spark timing at both normal and increased compression ratios is shown in Fig. 7(c). By examining Fig. 7(a), it is revealed that for each loading point, by keeping constant the normal spark timing, as engine compression ratio increases maximum cylinder pressure increases considerably. This is attributed to both the reduction of ignition delay period of the gaseous fuel and the considerable enhancement of the gaseous fuel combustion rate, especially during the initial stages of combustion, due to the acceleration of the flame front. The effect becomes more intense at part load conditions where the air to fuel excess ratio of the mixture is closer to the stoichiometric value, as compared to the one the mixture has at full load conditions, which favours the flame propagation mechanism. On the other hand, examining Fig. 7(b), by increasing spark advance by 2 or 4 crank angle relative to normal spark timing, keeping constant the normal compression ratio, a slight increase of the maximum cylinder pressure is also observed. Comparing the results depicted in Fig. 7(a) and (b), it is observed that for the same load the rate of maximum cylinder pressure increase achieved by accelerating the ignition timing appears to be milder, as compared to the respective rate observed with the increased compression ratio. Finally, observing Fig. 7(c), it is revealed that at 65% of full load engine performance could be further improved with a limited increase of compression ratio in combination with a mild acceleration of ignition timing by 4 crank angle, without creating a serious danger for the engine structure. On the

Fig. 6 e Calculated cylinder pressure and heat release traces for various spark timings with normal compression ratio at 1500 rpm and 65% load (a), and 1500 rpm and 100% load (b).

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Fig. 7 e Calculated maximum cylinder pressure at 65% and 100% loads and 1500 rpm engine speed for various compression ratios (a), various spark timings (b), and various combinations of spark timings and compression ratios (c).

other hand, at 100% of full engine load conditions, the aforementioned change of both parameters could lead to a significant problem with regard to mechanical strength of the engine, since a significant increase in maximum cylinder pressure is then observed.

5.1.3.

Effect on brake specific fuel consumption

It is stated here that for each test case examined, brake specific fuel consumption (BSFC) is estimated from the calculated brake power output and the calculated mass flow rate of the alternative gaseous fuel (i.e. wood-gas) and its lower heating value [34,57]. Theoretical predictions about the effect of each parameter on BSFC are given in the specific section. Thus, at each engine loading point examined, the variation of the calculated BSFC with compression ratio and spark timing is given in Fig. 8(a) and (b) respectively. Moreover, in Fig. 8(c) results are given concerning the variation of the BSFC versus spark timing at both normal and increased compression ratios. Observing Fig. 8(a), it is revealed that for both loads, the increase of compression ratio affects positively (i.e. reduce) brake specific fuel consumption. This improvement may be attributed to the amelioration of the gaseous fuel burning rate. The latter can be attributed to the increase of the cylinder charge

conditions (i.e. pressure and temperature) at the end of the compression phase and during the initial stages of combustion, which favour the flame propagation mechanism. The effect becomes more evident at part load conditions, since the lower air to fuel excess ratio combined with the increased cylinder charge conditions affect more intense the gaseous fuel combustion quality, as compared to the one observed at full load. Examining Fig. 8(b), it is observed that the ignition timing has a nonessential effect on engine efficiency. Specifically, at part load with normal compression ratio, an increase of spark advance leads to a slight increase of BSFC, as compared to the one observed with NSA timing. This may be attributed to the slight decrease of the brake power output observed at the specific test cases. Furthermore, at 100% of full load conditions, the change of ignition timing does not seem to affect seriously brake specific fuel consumption. This may be attributed to the negligible variation of the brake power output observed. However, by comparing the results between both engine loading points, it is revealed that at part load ignition advance seems to have a more intense effect on BSFC, as compared to the respective one observed at full load conditions. This is due to the fact that the lower air to fuel excess ratio plays a significant role on the flame propagation

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Fig. 8 e Calculated brake specific fuel consumption at 65% and 100% loads and 1500 rpm engine speed for various compression ratios (a), various spark timings (b), and various combinations of spark timings and compression ratios (c).

mechanism, thus contributing to the slight improvement of the gaseous fuel combustion quality observed at 65% of full load conditions, a fact that influences the variation trend of BSFC versus spark timing. However, observing Fig. 8(c), it is revealed that at full load conditions the engine efficiency achieved with NSA timing could be further improved if the acceleration of the ignition timing was combined with increased compression ratio. At part load conditions, the increased compression ratio in conjunction with advanced ignition timing seems to contribute very slightly to the improvement of the engine efficiency.

5.2. Effect of compression ratio and spark advance on exhaust NO and CO emissions 5.2.1.

Effect on specific NO emissions

The variation of the calculated specific NO emission as a function of compression ratio and spark advance relative to NSA timing, at 1500 rpm engine speed under part (i.e. 65%) and full (i.e. 100%) load conditions, is shown in Fig. 9(a) and (b), respectively. As known [62,64,65,67], NO formation

mechanism is predominantly controlled by the cylinder charge temperature and the local oxygen excess ratio. Thus, Fig. 9(c) and (d) depicts the effect of compression ratio and (e) and (f) illustrates the impact of spark timing on the calculated burning zone temperature at 1500 rpm engine speed under part (i.e. 65%) and full (i.e. 100%) load conditions. By examining Fig. 9(a), it is observed that for both engine loads examined, the increase of the compression ratio using normal spark advance timing results in a slight increase of the specific NO emissions. This may be attributed to the higher charge temperature caused by the increased compression ratio (Fig. 9(c) and (d)). Furthermore, at 65% of full load the effect of the increased compression ratio seems to be more intense, as compared to the one observed at full load conditions. At the same compression ratio the increase of engine load leads to an increase of the oxygen availability in the cylinder charge, which enhances the NO formation mechanism. However, at the same time, the increase of engine load affects negatively (i.e. decrease) the cylinder charge temperature (Fig. 9(c) and (d)), since the mixture strength is moved away from stoichiometry, a fact restraining the NO formation mechanism. Consequently,

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Fig. 9 e a and b. Calculated specific NO emissions at 65% and 100% loads and 1500 rpm engine speed for various compression ratios (a), and various spark timings (b). (c and d). Calculated burning zone temperature at 65% and 100% loads and 1500 rpm engine speed for various compression ratios. (e and f). Calculated burning zone temperature at 65% and 100% loads and 1500 rpm engine speed for various spark timings. (g). Calculated specific NO emissions at 65% and 100% loads and 1500 rpm engine speed, for various combinations of spark timings and compression ratios. by taking into account that NO formation mechanism seems to be more sensitive to the charge temperature than to the air to fuel excess ratio and also that at higher air to fuel excess ratios more brake power output is achieved, the difference in specific

NO emissions observed in Fig. 9(a) between the two engine loading points examined herein is justifiable. As far as the effect of ignition timing on specific NO emissions is concerned (Fig. 9(b)), it is revealed that for both

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engine loads examined, the acceleration of the ignition timing relative to the normal value results in an increase of specific NO emissions. Comparing the results shown in Fig. 9(a) and (b), it is observed that the aforementioned effect seems to be more intense as compared to the one observed with increasing compression ratio. As ignition timing increases relative to NSA timing, combustion starts earlier relative to the top dead centre (TDC) position. Thus, the gaseous fuel combustion rate is affected positively, especially during the initial stages of combustion, thus justifying the increase of the maximum cylinder temperature depicted in Fig. 9(e) and (f), a fact that favours significantly the NO formation mechanism. Inferentially, both strategies examined herein promote the NO formation rate. From the results given in Fig. 9(g), it is revealed that for both loads the specific NO emissions seem to be more sensitive to the acceleration of the spark timing rather to the increase of the compression ratio. Thus, for both loads, the acceleration of ignition timing at increased compression ratio results in a drastic increase of NO emissions, which does not seem to be restrained in a noteworthy way by the simultaneous increase of the compression ratio.

5.2.2.

Effect on specific CO emissions

The variation of the calculated specific CO emission with compression ratio and spark timing, at 1500 rpm engine speed under part (i.e. 65%) and full (i.e. 100%) engine loading conditions, is given in Fig. 10(a) and (b), respectively. As known [62,64,65], CO formation rate depends on the air to fuel excess ratio, the unburned gaseous fuel availability, and the cylinder charge temperature. The latter two parameters control the rate of fuel decomposition and oxidation. By examining Fig. 10(a), it is observed that for both loads examined the increase of compression ratio using the normal value of the spark timing (NSA) results in a slight decrease of the specific CO emissions. This is due to the fact that for the same load, increase of compression ratio results to an increase of the cylinder charge temperature, which eventually enhances the CO formation mechanism. However, at the same time, it affects also positively the CO oxidation rate due to the increased time interval during the expansion phase, for which high temperatures persist in the cylinder. Consequently, by taking into account that the CO oxidation rate seems to be affected by cylinder charge temperature more intensely than the respective formation one and also that at higher compression ratios more brake power output is

Fig. 10 e Calculated specific CO emissions at 65% and 100% loads and 1500 rpm engine speed for various compression ratios (a), various spark timings (b), and various combinations of spark timings and compression ratios (c).

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achieved, the slight decrease of the specific CO emissions observed with an increase of the compression ratio is justifiable. Furthermore, it is observed that for the same compression ratio the increase of engine load results to a decrease of the specific CO emissions. This may be attributed both to the lower cylinder charge temperature, which restrains both CO formation and oxidation rates, and also to the increased brake power output. Observing Fig. 10(b), it is revealed that for both engine load conditions with normal compression ratio, the acceleration of spark timing relative to NSA one affects negatively (i.e. increase) the specific CO emissions. As ignition timing is accelerated relative to NSA point, combustion starts earlier relative to the top dead centre (TDC) position. This results to an increase of the maximum cylinder temperature, which favours the CO formation mechanism. However, at the same time, spark advance results to a decreased time interval during the expansion phase for which high temperature persists in the cylinder, a fact influencing negatively the CO oxidation rate and thus allowing less mass of the formed CO to be completely oxidized. However, for the same spark timing, specific CO emissions observed at 100% of full engine load are slightly lower as compared to the respective ones estimated at 65% of full engine load. Finally, observing the results depicted in Fig. 10(c), it is revealed that for each engine operating point the effect of spark timing on specific CO emission seems to be more intense, as compared to the respective effect caused by the increase of compression ratio. At high engine load, the simultaneous increase of both parameters may lead to a slight decrease of CO emissions, which tend to converge to the respective one observed under normal engine operating mode (NEO).

emissions (i.e. specific NO and CO), the peak cylinder pressure/ BSFC and specific exhaust NO/specific exhaust CO trade-off curves are given in Fig. 11(a) and (b) respectively. Predictions are presented for the “J316GS” engine operating with wood-gas at 1500 rpm and at 65% and 100% of full engine load considering two compression ratios (i.e. CR ¼ 11(NCR) and 13) and three spark timings (i.e. NSA, 2 CA ANSA and 4 CA ANSA). According to Fig. 11(a) and (b), the increase of compression ratio in conjunction with the advancement of spark timing results in the increase of both peak cylinder pressure and BSFC and in the increase of both specific NO and specific CO emissions at 65% load with reference to its NEO point (i.e. NCR ¼ 11 and NSA). Again having as reference the pertinent NEO point at 100% load, increase of peak cylinder pressure and specific NO emissions and reduction of BSFC and CO emissions is observed in Fig. 11(a) and (b) when increasing compression ratio and advancing spark timing simultaneously. The reasons behind the observed effects of compression ratio and spark advance on peak cylinder pressure, BSFC, NO and CO emissions have already been discussed in previous sections in this study. Hence, it can be concluded that the simultaneous increase of compression ratio and spark advance in an HD SI engine operating with wood-gas has negative impact on engine endurance due to severe increase of peak cylinder pressure and on specific NO emissions at both part and full engine load. Potentiality for BSFC improvement and exhaust CO curtailment with increasing compression ratio and spark advance is envisaged only at full load conditions.

5.3. Effect of compression ratio and spark advance on peak cylinder pressure/BSFC and specific NO/specific CO trade-offs

In the present work, an existing two-zone phenomenological model is used to examine the effect of the compression ratio and the spark timing, on the performance characteristics and pollutant emissions of a heavy-duty, turbocharged, sparkignited engine fuelled with wood-gas. A good coincidence between calculated and measured values under normal engine operating mode (i.e. normal compression ratio and

In order to have a complete overview of the relative impact of compression ratio and spark timing on engine endurance (i.e. peak cylinder pressure), efficiency (i.e. BSFC) and gaseous

6.

Summary and conclusions

Fig. 11 e Combined effect of compression ratio and spark timing on (a) peak cylinder pressure/BSFC and (b) specific exhaust NO/specific exhaust CO trade-off. Predictions were generated for 65% and 100% loads and 1500 rpm engine speed.

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normal spark advance timing) was observed, for both performance characteristics and pollutant emissions. Specifically, the model predicts with reasonable accuracy the absolute values but, most importantly, it predicts very well the trends of the combustion and pollutants formation mechanisms with engine load. Acknowledging the predictive ability of the two-zone combustion model, this was used to examine the effects of compression ratio and spark advance on cylinder pressure and heat release traces, brake power output, maximum cylinder pressure, brake specific fuel consumption, and specific NO and CO emissions. From the evaluation of the theoretical findings, the following conclusions can be summarized: For normal spark timing, the increase of compression ratio up to 27% could lead to a sensible increase of the brake power output, which could be accompanied by a slight improvement of the brake specific fuel consumption. The effect becomes more evident for air to fuel excess ratios that are closer to the stoichiometric value. However, at the same time, it leads to an increase of maximum cylinder pressure. The latter may have detrimental repercussions on the constructional endurance of the engine, as it may be proven to be harmful to the engine mechanical strength. Moreover, the increase of the compression ratio could lead to a slight increase of the specific NO emissions. The effect seems to be more sensible at lower air to fuel excess ratios, where the maximum appearing increase is up to 10% higher relative to the value observed under NEO operating mode. However, the specific methodology appears to affect positively (i.e. decrease) the specific CO emissions. Thus, for both air to fuel excess ratios examined, the increase of compression ratio up to 27% could lead to a slight decrease (i.e. up to 15% relative to the value observed under NEO operating mode) of the emitted CO. For normal compression ratio, the acceleration of spark timing by 4 CA relative to NSA timing could lead to an almost negligible variation of the brake power output, which could be accompanied by a very slight increase of the maximum cylinder pressure. However, at the same time, advancing the spark timing affects negatively (i.e. increase) the brake specific fuel consumption. The aforementioned negative effect seems to be more intense at part load conditions (i.e. lower air fuel excess ratios). As far as the effect of spark timing on specific NO and CO emissions is concerned, it is revealed that the acceleration of combustion initiation affects positively (i.e. increase) the specific NO and CO emissions. However, the negative impact of spark advancing on specific CO emissions could be curtailed by increasing the compression ratio. In general, the increase of compression ratio accompanied with advanced spark timing in an HD SI engine operating with wood-gas could be a promising solution for improving engine efficiency, while simultaneously reducing CO emissions only at full load conditions. However, this potentiality is limited considering that the simultaneous increase of both CR and spark advance results in a severe increase of peak cylinder pressure and in an increase of NO emissions. The issue of increased peak cylinder pressure at full load conditions is crucial for the engine operational lifetime since engine operation with cylinder pressures higher than the one of the NEO

point at full load for long time will probably cause severe damage to the engine mechanical strength.

Acknowledgement The authors express their gratitude to the European Union for funding the research program “LIFT-OFF” under which the present investigation was conducted, and also to the GE Jenbacher Company for coordinating it. We must also thank GE Jenbacher Company for supplying us with valuable experimental data.

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