Combustion in dual-fuelled gas engines. The effect of LCV-gases and detonation stability

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International Flame Research Foundation

Finnish-Swedish Flame Days January 28 -29, 2009

Naantali, Finland

COMBUSTION IN DUAL FUELLED GAS ENGINES The Effect of LCV-Gases and Detonation Sensitivity Arto Sarvi1, Jorma Jokiniemi2,3, Jussi Lyyränen3, Ron Zevenhoven1* 1

Åbo Akademi University, Heat Engineering Laboratory, Biskopsgatan 8, FI-20500,Åbo / Turku Finland 2 University of Kuopio, Department of Environmental Sciences, Fine Particle and Aerosol Technology Laboratory, P.O. Box 1627, FI-70211 Kuopio, Finland 3 VTT Technical Research Centres of Finland, PO Box 1602, FI-02044 Espoo, Finland *Corresponding author: [email protected], +358 2 2153223

ABSTRACT: This work addresses two effects on the maximum power output of a dual fuel gas engine, being the calorific value of low calorific value (LCV) fuel gas and the sensitivity to detonation (knocking). During the adiabatic part of the cycle the LCV gas is compressed to a high pressure and temperature and the characteristics of the gas determine the available maximum power of the engine related to the gas detonation sensitivity. LCV-gas engines are characterised by combustion of gases with net calorific values lower than 30 MJ/m3n, while engine configurations are generally based on a nominal power from combustion of gases with a calorific value of about 30 MJ/m3n (natural gas). For LCV-gases this energy density can be as low as 4 – 10 MJ/m3n, and this requires de-rating of the gas engines operating with nominal power based on natural gas calorific value of 30 MJ/m3n. When the gaseous fuel octane index (OI) is lower than nominal minimum (i.e. 100) the engine volumetric compression ratio shall be reduced by, for example, using a special cylinder head gasket. Besides this, the use of ashless lube oils for gas operation mode permits only fuels with a sulphur content of 1wt. % or less, preventing pre-firing with deposits on the combustion chamber. The LCV gas coming from the gas generator into the combustion cylinder is cooled and filtered with water and therefore, it is important to know the influence of a certain water vapour volume percentage on the LCV calorific value. Depending on temperature, pressure and gas quality, the LCV gas solid particle content before the engine have to be controlled when running under LCV- gas mode. For a dual fuel LCV engine the pilot fuel is light fuel oil (LFO), and the pilot fuel contribution to the main heat energy content is about seven per cent. This paper reports how LCV quality can be related to detonation (knocking) tendency during gas-mode LCV-gas and dual-fuel combustion in large-scale medium speed, turbocharged and air cooled dual fuel engines. Keywords: Gas engine, Dual-Fuel, LCV-gases, Detonation, Knocking

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1. INTRODUCTION Interest in running engines on gaseous fuels has increased recently in view of the projected shortage of liquid fuels, particularly those of good quality and providing low emissions. There have been numerous papers published over the years relating various aspects of the operational and research experiences with gaseous fuel engines generally burning propane or methane. To widen the range of fuels in internal combustion engines LCV gases may provide a promising alternative. An increasingly important application for LCV-gas engines is co-generation, which makes the use of internal combustion engines attractive to provide electricity, cooling, heating and other services. These services have conventionally been provided by the electrical grid and other energy sources. Dual fuel LCV-gas engines are an interesting alternative for co-generation plants. Their relatively good thermal efficiency is sometimes their major advantage and their multi-fuel ability (gas or liquid) gives the end-user the possibility to modify the primary energy sources of the dual fuel engine according to economic variations. The main reason for the moderate success is the high liquid fuel price and environmental aspects compared to other prime movers and power sources with similar power and efficiency. This can be explained from the fact that the power of the gas engines is generally limited by the detonation to a maximum brake mean effective pressure of about 1700 kPa running on LCV-gas. On the other hand, with diesel dual fuel engine operation mixtures leaner than those for the appropriate flammability limit may be used. This is the consequence of diesel injection providing a continuous source of ignition irrespective on how lean the gaseous fuel-air mixture may be (Klat, 1965). Detonation (knocking) normally starts at the periphery of the combustion chamber during combustion in a zone where the mixture of the gas is not already burning. A sudden compression occurs, caused by the rapid expansion of the burning zone and the self-ignition limits may be exceeded. Avoiding this detonation requires the limitation of the compression ratio (CR), i.e. the cycle efficiency, which is proportional to CR, and the increase of the excess air which reduces the volumetric power of the engine and still increases its fuel consumption. As regards to self ignition this also limits the supercharging ratio (Karim, Klat, 1966). One of the problems to solve in LCV-gas combustion is the feeding of a gas which low caloric value (LCV-gas) is up to about ten (10) times lower than for example natural gas into the cylinder. The specific stoichiometric mixture of LCV-gas is nearly constant, ~2.62 MJ/m3 (n), and the gas volume is around 1/3 of the total volume that is introduced into the cylinder, Eq.1. This is an important parameter that has to be found as regards to engine power without de-rating. This also allows the use of unpressurised gas if the airgas mixer (carburettor) is installed before the turbocharger. In the LCV-gas case, the reciprocating dual fuel gas engine is the one of the most attractive choices due to its higher total efficiency and flexible fuel types (Daugas, 1983). This, of course would be expected to have significant consequences on the engine performance, particularly as far as the extent of LCV-gas fuel utilization by the dual fuel gas engine. But, it is to be shown that LCV-gas has sufficiently attractive features that make them desirable fuels for dual fuel internal combustion engines of the compression ignition types at different operational modes.

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2. DUAL FUEL GAS ENGINE The burning of gas in the internal combustion engine is as old as the engine itself since the first reciprocate internal combustion engines all burned gas. The term ‘dual fuel’ is used to denote engine burning gas ignited by small quantity of light fuel oil (LFO) usually in the region 5 – 7 % of the total heat consumption, Fig.1. Dual fuel engine installation has demonstrated that gas burning engines can operate with high efficiency provided that reliable ignition of weak air to fuel mixtures can be achieved. The liquid fuel ignition provided by a diesel fuel injector spray is particularly suited to the ignition of very weak gaseous mixture. This means that efficiencies given by dual fuel engines can be achieved using spark plug ignition without, of course, the cost penalty of liquid fuel oil consumption. But, otherwise the engine flexible operating with two fuels is very positive profit, Tab.1. 3. COMBUSTION – IGNITION AND DETONATION Generally gas fuels have a good octane index which means that, they burn only when (Felt, Steele, 1962). - they are ignited thanks to a hot points - they are mixed to a combustion in given proportion, pressure, temperature and fuel mass on mixture mass for flammability limits, Eq. 2. - the mixture in which there is exactly the necessary quantity of air able to burn all the fuel (stoichiometric) is the one for ignition and firing is the most sensitivity (easiest), Eq.3 for ignition low limit. - flammability limits vary with temperature and pressure, Eq.4 for high limit. - temperature and pressure conditions exist for which a stoichiometric mixture gives self firing (ignition), Eq.5 for low and high limit. - this self firing take place whatever the mixture rate but for temperature and pressure higher and higher when the stoichiometric mixture is more and more distant. This means that they have to be burnt in engines with controlled firing, i.e. in which carburetted mixture is (Karim, 1983): - compressed - fired thanks to an auxiliary device when requested – electric spark or pilot injection of a liquid fuel with a high cetane index shall be higher than or equal to 45, which is self firing in the temperature and pressure conditions at which the carburetted mixture is compressed, Tab.2. 4. CONSEQUENCES The main technical problems of gas combustion are to maintain the mixture inside the flammability limits to make sure that it will be firing, Fig.2. It is necessary to avoid self ignition of the air gas mixture which, if homogeneous, starts simultaneously from every point in the combustion chamber, and makes a very large and instantaneous raise like a deflagration – this is detonation (knocking). To avoid this detonation requires a limitation of the volumetric compression ratio, i.e. the cycle efficiency (which is proportional to it) and the increase of the excess air which

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reduces the volumetric power of the engine and increases its fuel consumption. This self firing also limits the supercharging ratio of the gas engine (Ziner, 1978). Those are the main reasons why up till now gas engines have had a lower efficiency and a more limited supercharging ratio than the corresponding size diesel ones. Therefore, the most part of industrial large-scale diesel engines are still compression engines. As the market of gas engines is very small industrial gas engines are diesels adapted to gas. Furthermore, they are de-rated and thus their efficiency is lowered. They are expensive and uneconomical, and their market is small. Therefore, it is necessary to design gas engines with low fuel consumption and high bmep which means high volumetric compression ratio and high supercharging ratio, but the major problem is detonation (knocking). Parameters that influence detonation are (Fig.3.): - fuel octane index or methane index - temperature and pressure conditions at the end of the compression stroke which depend on: - volumetric compression ratio - compression polytropic coefficient - intake valve closure (Miller-cycle), Fig.4 - mixture temperature and pressure at cylinder inlet - ignition timing - excess of air in mixture: general (homogeneous) and local (stratified) 5.

OCTANE AND METHANE INDEX

The fuel should not ignite by itself, but at higher temperatures and pressures self ignition can occur as defined by the octane index - see also, Tab.3: - iso-octane and heptanes per cent of volume have same detonation limit in the same engine with the same measuring device as the same fuel to number (Karim et al., 1967) - some gaseous fuels have an octane index above 100 but can be increased to 130 to pure methane and LCV-gas over 140, Eq.6. - it is possible to use a methane index, but it is different from the octane index, see Section 12. The octane index of a gas which composition is known can be calculated from tabelised data because the mixture rule is applicable except for a gas containing hydrogen. If hydrogen and carbon monoxide are both present in the gas mixture they have to be considered as only one component which octane index is a function of the ratio of these two LCV-gases. In addition, according to carbon dioxide and nitrogen content an improvement limited to 17 octane index numerical units is found, see Fig.5. When the gaseous fuel octane index is lower than or equal to 100, the engine volumetric compression ratio shall be reduced by 40%, e.g., with the special cylinder head gasket. Nominally dual fuel LCV-gas engines are designed for octane index ≥ 119 and for lower octane index value of 119, the engine power has to be de-rated by one per cent by point below 119 and/or modify engine adjustments. Generally, dual fuel LCV-gas engine operation is possible only when firing a gaseous fuel with an octane index ≥ 95, Tab.4.

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However the mixture (CO + H2) may have a higher octane index (Fig. 6) depending on the H2/CO content ratio. For instance H2 = 4% and CO = 16%, as 16/4 = 80/20 in abscissa gives an octane index of 105 for 25% of fuel. Moreover, the nitrogen and carbon dioxide act as detonation inhibitors. When the concentration is over 5 volpercent the fuel quality is improved (Fig. 6), and the octane indices of H2/CO2 and H2/N2 mixtures, with no other combustible gas are given by Figs. 7 and 8. 6. INLET VALVE CLOSURE INFLUENCE The inlet valve is generally closed after BDC to make profit of the inertia of gas momentum which is inside the inlet pipes, and increases the so-called filling coefficient. According to the mean piston speed this lag varies from 32 to 45 degrees of crankshaft. In this case, the dual fuel diesel engine volumetric compression ratio is higher than the actual SI-engine, and the swept volume which means the effective cylinder volume is smaller than the volumetric one. The same result could be obtained by an intake valve closure at the same crankshaft angle before the BDC (Miller-concept). A compression ignition (diesel) engine should be in a “bad position” with such intake valve timing: temperature after compression is too low for diesel fuel ignition. However, at rated speeds the filling coefficient of the cylinder is lower at the end of the compression stroke and, if the engine is a dual fuel LCV-gas engine, the detonation limit is pushed back from the running conditions, Tab.5. The Miller inlet valve system promises positive advantages for turbocharged dual fuel LCV-gas engines because of the effect of the temperature at the start of compression and combustion on detonation limited output. The Miller system applied to dual fuel LCV-gas engine has been investigated in detail when it comes to variations in the early closing of the inlet valve. According to these results, the increase in the power output of the dual fuel gas engines by about (25-40) % is found, depending on the possible charge (boost) level, the charge temperature, the turbo charging, etc. (Miller, Lieberherr, 1957). In addition, the favourable effects of a combination of exhaust turbo charging and property tuned Miller inlet valve timing on the exhaust gas emissions are achieved. The Miller inlet valve timing is nowadays used in e.g. large-scale diesel engines to reduce NOx emissions, Tab. 6. 7. LCV GAS MIXTURE TEMPERATURE BEFORE CYLINDER When the engine is turbocharged the exhaust gas temperature after the compressor is too high to allow for an acceptable cylinder filling. Therefore, the exhaust gas is cooled by an intercooler. Increasing the air-gas mixture temperature at the end of the compression stroke makes the detonation to occur easier. Decreasing the mixture specific mass and thus also the filling coefficient, if it is pure air, makes the final mixture richer which means that if excess of air becomes closer to the stoichiometric and the detonation sensitivity increases Fig.9. If there is such a mixture that excess of air remains constant but the mass of introduced mixture of fuel in the cylinder (due to volumetric flow limitations) decreases and power loss is the result. On turbocharged dual fuel LCV gas engines booster pressure can be increased but the detonation limit then becomes closer (Daugas, Grosshans, 1985).

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Nevertheless, these two before mentioned phenomena can be used even at low load because there is a too large excess of air or there is a too low intake temperature or the octane index of the fuel is too high, therefore the fuel is firing irregularly, too late or not at all. If the dual fuel LCV gas engine is equipped with an air cooler at compressor outlet the control of this air or air-gas mixture temperature is easy and depends on cooling fluid temperature and it is influence for the thermal balance, Tab.7. 8. CYLINDER INLET PRESSURE The cylinder inlet pressure determines the pressure and temperature at the end of the compression stroke, the higher it is, the closer is the detonation limit. But this pressure also determines the quantity and quality of the mixture which can be introduced in the limited volume of the cylinder, thus the power which can be developed by the engine. For LCV – gas engines the gas supply pressure shall be between 1.5 and 4.0 kPa at full rate (Baxter, 1985). During the design of the gas engine it is necessary to look for the following parameters if optimization of the power of the gas engine with a given octane index gas is desirable. Volumetric (geometric) compression ratio should be the lowest possible but it is limited for dual fuel gas engines by the all weather starting possibilities. This limit can be improved because of an intake valve closure at the BDC. Another parameter is the allowed minimum temperature of the charge air or air-gas mixture at cylinder inlet is determined by site conditions. When the design is optimized power can still be increased thanks to two parameters which can be adjusted during the running: the ignition timing and excess air ratio (ASTM,1971). A firing retard (later) always allows for shifting back the detonation limit but at the cost of the gas fuel specific consumption. In addition, too late ignition may generate combustion which ends in the exhaust manifold or with too large an ignition lag firing a great quantity of fuel gives a rough running. So for each load and for each excess of air there is an optimum ignition time that can only be defined by tests (Ferretti, 1941).Excess air is a parameter which may be adjusted almost endlessly. Moreover, it is easy to determine because the excess air only depends on the exhaust gas temperature at the cycle end. This can be checked with the help of the exhaust gas temperature average at the outlets of the cylinders or at the turbine inlet which is linked to it. The adjustment of excess air is also easy by regulating pressure or a flow rate ratio of gas and air control (Leiker et al., 1971). 9. LUBRICATION OIL FOR GAS ENGINE Medium speed diesel engines commonly burn residual (HFO) fuel oil, which may contain up to 4 wt. % of sulphur. The lubricate oil therefore has to contain an appropriate concentration of alkaline additive, i.e., it has a high Total Base Number (TBN), to neutralise the corrosive combustion acids that may contaminate the oil. In most engines this type low-ash lube oil is needed to minimise the formation of combustion chamber deposits, which could cause pre-ignition. Such engines are not so sensitive to combustion chamber deposits, but it is necessary to change lubricating oil by another consistent in the gas fuel burning mode (Shell, 1980).

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Prior to the shift to gas mode, a change of the lubricating oil for gaseous fuel one is required. Moreover, the engine should be operated at least two hours in diesel mode with diesel oil at load ≥ 0.6 MCR to burn out deposits created by high ash content lubeoils which could generate pre-ignition during the dual fuel mode running. Ashless lubricating oil for is used for gas engine operation only with fuels which sulphur content is ≤ 1 wt. %. Components of ashless lube oils are combined with some other components during the combustion process and some of these are deposited on the combustion chamber walls. As long as the fuel contains sulphur, these deposits are sulphates which are in powder form blown out through the exhaust valves. However, when the engine is running in gas mode, as gas is generally sulphur free, there are carbonates sticking on the cylinder walls and which become and remain incandescent (as a result of ash from lube oil combustion) and this could be reason for pre-firing which means noise and very high combustion pressure (knocking). Generally, over 16 hour’s continuous diesel mode running with a liquid (HFO) fuel sulphur content > 1 %, it is necessary to change lubricating oil for liquid fuel burning mode. 10. LCV GAS HEAT VALUE As can be assumed the gas heat value, heat density kJ/m3 (n), is a very important factor for dual fuel gas engine power. The energy filling the cylinder is related to engine speed, turbo charging air/gas mixture pressure and temperature before cylinder. Therefore, the dual fuel gas engine with lean heat value gas is designed normally for a gas with low heat value (LHV) ≥ 4000 kJ/m3(n) i.e. 1.11 kWh/m3(n). When the gas is coming out of a producer it is most of the time cooled and purified by moistening, and the gas is finally saturated when it reaches the engine. According to the tables of saturation steam pressure of the water as a function of temperature the percentage in capacity and factor by which one should multiply the LHV of the dry gas to have that of a gas saturated with steam can be calculated. For example if the LHV of dry gas is 1.4 kWh/m3 (n) and temperature is 50ºC (323 K) the saturated LHV will be 1.4 * 0.891 = 1.247 kWh/m3 (n), Eq.6. De-rating factor for a stoichiometric air requirement of 0.26 m3/MJ can be amended power calculated by equation, Eq.7. As a result, de-rating factor is 0.88 for 3 MJ/m3 (n) and 0.70 for 2 MJ/m3 (n), etc. The cumulative phenomena with the de-rating for gas LHV and octane index can be determined by the factor calculated by Eqs.7 and 8. 11. LCV GAS PHYSICAL PROPERTIES The physical properties of the gas mixture before the engine have to fulfill the following limits for the pressure, temperature, liquids and solid particles: - LHV gas ≥ 4000 kJ/m3 - temperature ≤ 50ºC - pressure ≥ 4 kPa - liquids, tar micro-mist content ≤ 30 ppm, free of water droplets and tar - solid particles ≤ 3 mg/m3(n) - can be steam saturated -

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12. PILOT FUEL – OCTANE AND METHANE INDEX Normally the engine speed governor controls the gas admission valves, see Fig.1, when running in a dual fuel mode. In most cases the fuel (pilot) pump is usually locked at a constant setting representing certain, about 5-7 %, of the total energy input to engine full load. Detonation is due to the auto-ignition of the charge located away from the ignition combustion zone around the pilot fuel. The difficulties encountered in the calculation of the octane index of liquid fuels from their chemical composition are well known. Octane index was defined for liquids for which vaporization during compression decreases the chamber ambient temperature in function of their vaporization heat. For gaseous fuel this parameter is not a problem nor is their selfignition temperature. Moreover, methane index is based on the comparison of the detonation of a mixture of methane and hydrogen. However, the use of the methane index instead of the octane index is allowable in the following cases (Fig.10 and Eqs. 8, 9, 10, 11): - comparison the engine performance - self - flammability temperature - methane index less than 60 - methane – hydrogen mixture law - methane index versus octane index maximum value 13. CONCLUSIONS During 1980-1985 laboratory test work and research was directed to the development of the dual-fuel LCV gas engine. This work was followed by extensive field tests. Because the development was based on an established production diesel engine, the mechanical problems have been minimal. The use of carburetion before the turbocharger for LCVgas, methane and propane gas by the using Miller-cycle for inlet valve timing profile will further decrease detonation sensitivity with gas mode operation. The use of the carburetion before the turbocharger and the intermediate air-gas cooler led to an excellent performance both LCV- gas, natural and propane gas mode. A reason for it is excellent gas-air mixture control. Comparing engine operating modes it is found that the diesel has the best efficiency closely followed by the DF-engine (Tab.7). ACKNOWLEDGEMENTS Many individuals have contributed to this paper. The authors particularly wish to express their appreciation to Claude F. Daugas, Consultant, 6 rue Gallieni, 78000 Versailles, France, for his support and help in dual-fuel gas engine development.

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NOMENCLATURE AR BDC Bmep BTE C2H4 C3H8 CCAI CG CH4 CI CO CO2 CR CS D DF EBTE FL GC H2 HFO HHV IFWLAR IFWRAR IVC IVCR L LCV LFO LHV MCR MHV MI N2 NG NOx O2 OI P p SI Std T or t TBN THC V VCR η λ ( = α c:)

air ratio bottom dead centre break mean effective pressure break thermal efficiency ethylene propane calculated carbon aromacity index concentration gas component methane cetane index carbon monoxide carbon dioxide compression ratio crankshaft density dual fuel engine break thermal efficiency flammability limits gas component flammability limit hydrogen heavy fuel oil high heat value ignition firing weak limit air ratio ignition firing rich limit air ratio inlet valve closure inlet valve closure retard stoichiometric fuel/air low caloric value light fuel oil low heat value maximum continuous rate mixture heat value methane index nitrogen natural gas nitrogen oxide (= NO+NO2) oxygen octane index power pressure spark ignition standard temperature total base number total heat consumption viscosity volumetric compression ratio efficiency air ratio

Subscripts: l, h, th i, in, g

low, high, theoretical component, inert, gas

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n, s, a o, 3, 4 c, ca

nominal, saturated, actual (total) ambient, before cylinder and turbine combustion, calculated

Equations MHVg =

FLl, h =

FLl =

LHVg * ηc

1 + λ * L th

(1)

100 1 + λ * L th

100 ⎛ CGi ∑ ⎜⎜ GC l ,i ⎝

FLh =

FLl ,h =

(2)

⎞ ⎟ ⎟ ⎠

(3)

100 − ∑ CGin ,i ⎛ CG ∑ ⎜⎜ GC i h ,i ⎝

⎞ ⎟ ⎟ ⎠

100 CG ∑ ( GC i ) l , h ,i

(4)

(5)

LCV s = LCV d * (1 −

ps ) pa

0.6 * LHVg Pca = Pn 1 + 0.36 * LHVg

(6)

(7)

OI = 60 + 0.882 * MI − 0.00182 * MI 2

(8)

CCAI = 976 − 3.4 * CI

(9)

CCAI = D − 140.7 * log log(V + 0.85) − 80.6

(10)

OI = (3760603 / CI )

0.3636

(11)

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References ASTM (1971). D 2699-IP 237. Test for Knock Characteristics of Motor Fuels by the Research Method. BAXTER, I. (1965). Determination and Significance of Gaseous Fuel Octane Numbers. Transaction of the ASME Journal of Engineering for Power. April 1965. DAUGAS, C.F. (1983). Gas Fed Engines Progress. SEMT – VI Meeting, 93202 Saint Denis – Cedex 1, France. DAUGAS, C.F., GROSSHANS, G. (1985). Gas Fed Engines Progress. D5 Conference CIMAC, Oslo, Norway. FELT, A.E., STEELE, W.A. (1962). Combustion Control in Dual Fuel Engines, S.A.E., Trans. Vol. 79 (1962), p 644. FERRETTI, P. (1941). Die Klopffestigkeit einiger Gase (Knock rating of some gases). Kraftstoff 1941 KARIM, E.A., KLAT, S.R. (1966). The Knock and Auto-ignition Characteristics of some Gaseous Fuels and Their Mixtures. J. of the Inst. Of Fuel. Vol. 39, March. 1966, pp 109-119. KARIM, G.A., KLAT, S.R., MOORE, N.P.W. (1967). Knock in Dual-Fuel Engines. Proc. of the Inst. Of Mech. Eng. Vol. 181, March. 1967, p 453-466. KARIM, G.A. (1983). The Dual Fuel Engine of the Compression Ignition Type – Prospects, Problems and Solutions – A Review. For Presentation at the Conference on Compressed Natural Gas as a Motor Fuel. Pittsburgh, June 1983. KLAT, S. R. (1965). Combustion Mechanisms in Dual Fuel Engines. Ph.D. Thesis in Mech. Engg. Of London University. LEIKER, M., CHRISTOPH, K., RANKL, M., CARTELLIERI,W., PFEIFER,U. (1971). The Evaluation of the Anti-knocking Property of Gaseous Fuels by Means of the Methane number and its Practical Application to Gas Engines. A2 Conference CIMAC, Stockholm, Sweden. MILLER, R., LIEBERHERR, H.U. (1957). The Miller Supercharging System for Diesel and Gas Engines operating Conditions. CIMAC-Congress 1957, Zürich, p 787-803. SHELL (1980). A Guide to Shell Lubricants. Prepared by Lubricants Development Division (MKDL) Shell International Petroleum Co. Ltd., Shell Centre, London, February 1980, p 13 – 17. ZINNER, K. (1978). Supercharging of Internal Combustion Engines. Springer – Verlag Berlin, Heidelberg, New York, 1978.

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Figures

Fig.1

Dual – fuel LCV gas engine system

By: PhD T.A. Bradshaw, 1985

Fig. 2 Ignition flame weak-rich limit vs. air ratio

13

= Air Ratio

By: PhD C.F. Daugas, 1984

Fig. 3

Charge air temperature and firing limit – octane index

By: PhD C.F. Daugas, 1984

Fig. 4 Valve - timing diagram vs. Miller inlet valve

14

OI Increase

18 16 14 12 10 8 6 4 2 0

OI_CO2 OI_N2

0

OI Mixture

Fig.5

10

20

30 40 50 CO2 & N2 Inhibitor %

60

70

Octane index vs. CO2 or N2 inhibitors

105 100 95 90 85 80 75 70 65 60 100,0

Fig. 6

4,00

1,50

0,67 Ratio H2/CO

0,25

0,11

0,01

Octane index vs. Ratio H2/CO

OI Mixture

105 100 95 90 85 80 75 70 65 60 100,0

9,00

4,00

2,50 Ratio H2/CO2

Fig. 7 Octane index vs. Ratio H2/CO2

1,50

1,00

0,67

OI Mixture

15

100 95 90 85 80 75 70 65 60 100,0

9,00

4,00

2,50

1,50

1,00

0,67

Ratio H2/N2

Fig. 8 Octane index vs. H2/N2

By: PhD CF Daugas, 1984

Fig. 9

Detonation region weak – rich vs. control parameters

Octane index (OI)

150 130 110 90 70 50 0

20

40

60

80

Methane indx (MI)

Fig. 10 Octane index vs. Methane index

100

120

140

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Tables Table 1

Engine properties

Engine power Engine speed Engine Bmep Bore Stroke VCR

400 - 440 500 – 514 1600 – 1800 400 460 10:1

kW/cyl rpm kPa mm mm -

Table 2 Pilot fuel properties Fuel HHV Fuel LHV Density (15oC) Viscosity (80oC) Cetane Index

45.5 42.8 834 3.32 ~50

MJ/kg MJ/kg kg/m3 mm/s2 -

Table 3 LCV - fuel gas properties Gas type Producer gas Blast furn.gas Coal gas Water gas City gas

H2 % 6 - 15 4 27 49 50

CO % 23 -17 28 7 42 8

CH4 % 0.5 -4 0 48 0.5 29 – 32

C2H4 % < 0.5 0 13 0 4

CO2 % 5-8 8 3 5 2

N2 % rest 60 2 3 7–4

LCV Fuel LHV kJ/m3(n) 3024 - 5714 3 975 28 737 10 750 19 514

Table 4 LCV - fuel gas properties Species & Characteristics O2 H2 CO CH4 C2Hx C3Hy CO2 N2 Octane index LHV kWh/m3(n) Air/fuel m3/m3 Flamm. limits kg/kg Density kg/m3(n) Gas fuel kg/kWh

Gas type 1

Gas type 2

Gas type 3

1.1 5.3 10.2 4.4 4.7 0.2 14.2 59.9 125 1,83 1.61 0.12 – 0.60 1.272 0.697

0 9.4 13.2 4.5 0.2 0.2 16.3 55.4 124 1.41 1.18 0.15 - 0.65 1.235 0.877

0 21.0 14.0 1.3 0 0 16.0 47.7 113 1.24 0.96 0.25 – 0.75 1.113 0.895

Air/Fuel ratio m3/m3 0.61 – 0.24 0.76 7.24 2.21 4.8

Fuel OI Oct.index ~ 120 122 104 98 91

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Table 5

LCV - gas engine heat balance

Load Shaft power Bmep Pilot fuel Pilot + Gas fuel Shaft efficiency Exhaust gas Exh.gas t5 Exh. gas Exh. gas HT water HT water Air flow Air flow t2 Air flow t3 Air flow Air flow Lube-oil Lube-oil Radiation Radiation

Table 6

25 101 419 60 372 27.2 10.9 430 132 35.6 63 17 10.8 63 28 11 2.9 43 12 22 5.8

50 202 839 30 607 33.3 8.21 466 218 36.0 84 14 8.2 103 32 33 5.5 47 7.7 23 3.8

75 303 1258 20 817 37.1 7.01 479 288 35.2 97 12 7.0 143 37 63 7.7 43 6.3 23 2.9

100 404 1678 15 1022 39.5 6.14 496 349 34.1 109 11 6.1 178 45 92 9.0 44 4.4 24 2.3

Intake vale closure influence (See, Fig. 9)

Engine n Air + Gas T3 p3 / po VCR IVCR Compression p Compression T

Table 7 Load THC BTE

% kW/cyl kPa g/kWh kW/cyl % kg/kWh o C kW/cyl % kW/cyl % kg/kWh o C o C kW/cyl % kW/cyl % kW/cyl %

rpm K o CA ABDC MPa K

0 300 1 11.5 0 40 3.05 2.57 797 758

90 1.16 604

500 318 10.2 0 2.57 760

2.9 11.5 0 7.8 815

3.9 10.2 0 8.8 777

40 8.8 845

Engine type thermal efficiency % % %

50 78 38.0

DIESEL 75 100 75 73 39.5 40.0

50 100 29.6

DF-ENGINE 75 100 84 76 35.3 38.7

50 95 31.1

SI-ENGINE 75 100 85 79 34.8 37.6