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: ron.zevenhoven@abo.fi, +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
2
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.
3
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
4
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.
5
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).
6
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).
7
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
-
8
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.
9
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
10
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)
1 + λ * L th
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)
ps
)
pa
LCV s = LCV d * (1 −
(6)
0.6 * LHVg
Pca
=
Pn 1 + 0.36 * LHVg
(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 )
(11)
0.3636
11
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.
12
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
16
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
17
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