Due to the skyrocketing of energy price; couple with the increasing concern
about energy shortage and environmental protection, research on improving the
engine fuel economy and reducing exhaust emissions has become imperative in
combustion and engine development (Rousseau et al.,
1999). The development of alternative fuel engine has been the major concern
of the engine community and alternative fuels usually belongs to clean fuel
compare to diesel and gasoline in engine combustion process. The introduction
of the alternative fuel is expected to minimize the fuel shortage and reduces
the exhaust emission from the engine (Ben et al.,
Natural gas is regarded as one the most promising alternatives fuels and probably
one of the cleanest fuel in combustion. The use of natural gas has been realized
in both spark ignition and compression ignition engine. Natural gas comprises
of mixture of different gases where methane is its major component. The combustion
of natural gas produces less emission when compared to that of gasoline and
diesel engine due to its simple chemical structure and absence of fuel evaporation.
The engine possess high anti-knocking capability due to its high octane number
and this allows it to operate at even high compression ratio, leading to further
improvement of both power output and thermal efficiency. However, natural gas
as an engine fuel suffers two major setbacks viz: Slow burning velocity and
poor lean burn capability which often leads to cycle-by-cycle variation (Rousseau
et al., 1999; Ben et al., 1999). Traditionally
to solve these problem, an increase in flow intensity in cylinder is introduced,
unfortunately, this method always increase the combustion temperature and heat
loss to the cylinder wall as well as high NOX emission (Akansu
et al., 2004). To solve these problem of high NOX , EGR
is introduced into the intake system with the aid of regulating valve.
Exhaust Gas Recirculation (EGR) is one of the common ways to control in-cylinder
NOX production and is used on most modern high-speed direct injection
diesel engine. Because NOX formation progresses much faster at high
temperatures, EGR reduces the amount of NOX the combustion generates.
However, large EGR mass fraction may have an unfavorable effect on the engine
performance, fuel economy and HC emission. Proper determination of EGR mass
fraction needs to be determined by comprehensively evaluating engine performance,
fuel economy and NO emission (Das and Watson, 1997;
The most important engine variable that affects NOx emission are fuel/air equivalence ratio, spark timing and burnt gas fraction in charge. The burnt gas fraction in charge depends on the amount of diluents introduction such as EGR. NOX emission can be greatly decreased via lean combustion, retarding the ignition timing and introducing EGR. Most of the previous work concentrated on the method of using lean combustion and retarding ignition timing, while few literature were found in introducing EGR with natural gas to reduce NOX emission.
Amr and saiful investigated the use of EGR in a supercharged natural gas engine
(SI) engine (Ibrahim and Bari, 2010; Zheng
et al., 2004) studied the effects of EGR on the combustion and emission
of the compression ignition engine. Peng et al. (2008)
conducted on experimental study on the effect of Exhaust Gas Recirculation (EGR)
on combustion and emission of the direct injection compression ignition engine.
Their results indicated that the use of percentage EGR dilution in the inlet
mixture decrease the oxygen concentration and consequently, it decreases the
combustion rate significantly. Engine key performance parameter such as brake
torque, brake power decreased with increased of EGR; while brake specific fuel
consumption increased. The increased of EGR dilution in the inlet mixture decreases
both the maximum cylinder temperature and oxygen concentration which leads to
a significant reduction in NO emissions. Allenby et al.
(2001) reported that methane-hydrogen gas engine (Hythane) engine could
tolerates up to 25% EGR, while maintaining a coefficient of variability of indicated
mean effective pressure below 5% (Allenby et al.,
2001; Erjiang et al., 2009). To date, no
literature has ever examined the impact of EGR on the performance and exhaust
emissions of the stratified charge piston DI-CNG engine and explore the applicability
limit of EGR most especially under lean combustion mode.
This study will seek to investigate the effects of different EGR rates on the performance and exhaust emissions of the stratified charge piston direct injection compressed natural gas engine and quantitatively analyze Brake torque, Brake power, Fuel conversion efficiency, Brake Specific Nitric Oxide (BSNO), Brake Specific Unburnt Hydro Carbon (BSUHC), Brake Specific Carbon monoxide (BSCO) emissions and Coefficient of variation of the indicated mean effective pressure (COVIMEP). The study is expected to provide information on natural gas engine-EGR tolerance limit (driveability) and at this point, define the operating conditions for the engine application or operation.
MATERIALS AND METHODS
The specifications of the test engine are listed in Table 1 and Fig. 1 shows the schematic of the experimental set-up. A four-stroke single cylinder research engine was used to investigate the effect of EGR on the direct-injection compressed natural gas engine.
The supplying system of the engine consists of a natural gas, gas valve, gas
flow meter, a pressure regulator. An electronic control unit was used to adjust
the amount of CNG and the excess air ratio into the system, also controlled
by the electronic are injection timing and spark timing. The composition of
the natural gas was given in Table 2 below (Erjiang
et al., 2009).
The recycle exhaust gas was taken from a hole located on the exhaust pipe with the aid of connecting pipe. Cooling of the hot exhaust gas was done by passing it through a water-cooled heat exchanger. A regulating valve was installed: in order to regulate the exhaust gas flow. Percentage increase of EGR in the inlet mixture was done by increasing the amount of the exhaust gas flowing back to the engine intake system. Figure 2 shows the schematic diagram of the EGR control system.
Air, natural gas (from pipeline supply) and cooled exhaust gas were mixed in
the engine intake. A pressure gauge was used to measure the inlet pressure.
The research described in this paper was carried out at wide open throttled
(WOT), with spark timing set at 31.5°CA, injection timing 300°CA (early
injection timing) and four excess air ratio comprises of (
= 0.9, 1.0, 1.1 and 1.2) which represent rich, stoichoimetric, slightly lean
and lean mixture respectively and two engine speed were being utilized viz,
2000 and 3000 r min-1.
|| Engine specification and test bed
|| Composition of natural gas
|| Schematic diagram of the experimental setup
|| Schematic diagram of EGR control system
The excess air ratios are obtained by setting from the Gasmet gas analyzer
which contains oxygen sensor.
The system for the acquisition of in-cylinder pressure is composed of:
||Piezo electric cylinder pressure sensor- AVL QH32D, gain 25.28pc
bar-1 range 0-200bar
||Charge amplifier- AVL3066AO
||Shaft position encoder-AVL364C
||Piezos resistive pressure sensor fixed inside the inlet manifold
The exhaust gas is being measured with the aid of Gasmet gas analyzer which is capable of analyzing up to forty (40) exhaust species.
The reaction mechanisms of natural gas, EGR and air mixture can be represented as follows:
The overall combustion reaction for CH4+EGR+air mixture can be written as:
where, x is the excess air parameter, α is the EGR mole fraction knowing
that the (methane + EGR) mole fraction is equal to 1. The equivalence ratio
(φ) is defined by the fraction of O2 needed to obtain complete
combustion of the fuels. The EGR as it was being utilized comprises of CO2
and H2O only.
METHOD OF CALCULATION
A zero-dimensional thermodynamic model was used to calculate the heat release rate in the study. The model neglects the leakage through the piston rings and thus the energy conservation in the cylinder is written as follows:
From the first law:
The gas-state equation is:
The differential equation of the gas-state equation with crank angle θ is given as:
The heat release rate dQ/dθ can be derived from Eq. 2 and 3 as follows:
Assuming ideal gas, specific gas constant =R=Cp-Cv and therefore the value of specific heat at constant volume can be written as:
Substituting Eq. 6 in 5; then equ. 5 becomes:
More so, the convective heat-transfer rate to the combustion chamber wall can be calculated from the relation:
The mass fraction burned can be calculated based on Rassweiler-Withrow method by using the average value of pressure data of 100 cycles.
RESULTS AND DISCUSSION
The experimental results in this section were obtained from single cylinder
four-stroke direct-injection compressed natural gas engine having 76 mm bore
and 88 mm stroke with a compression ratio of 14:1. The fuel comprises of compressed
natural gas with different fraction of EGR (i.e., CNG-EGR) were being utilized.
The experiment were carried out with EGR rate consists of 0, 10, 20, 30, 40,
50 and 60% at varying excessive air ratios (
= 0.9, 1.0, 1.1 and 1.2) and engine speed of 2000 and 3000rpm at Wide Open Throttle
(WOT). Ignition timing was set to 31.5°CA. A high pressure injector (18
bars) was used on the central direct injection system.
Effects of EGR Rates on performance parameters
Brake torque: Figure 3 shows the variation of brake
torque with EGR rates for different speeds and injection timings at WOT.
|| Brake torque versus EGR rate
In the Fig. 3a it is obvious that the brake torque decreases
as the EGR rates increases. This is largely due to the fact that as EGR rate
is increased, the concentration of air and fuel in the cylinder decreased and
consequently, it reduces the brake torque. The highest value of brake torque
is presented at stoichiometric mixture (
= 1.0) when compared to the other mixtures. The reason being that at stoichiometric
mixture, there is efficient combustion, all the air provided is effectively
utilized for combustion process. In addition, increasing the engine speed from
2000 to 3000 rpm at injection timing of 300°CA as shown in Fig.
3b increases the brake torque due to increase in friction mean effective
pressure at higher speed which decreases the mechanical efficiency of the engine.
Also the brake torque at rich mixture is higher than that at moderately lean
= 1.2) due to the fact that at rich mixture (excess fuel), when the engine speed
is increased, more fuel will be burnt to deliver more torque.
In furtherance, at 300°CA injection timing and engine speed of 2000 rpm
i.e., Fig. 3a, it is obvious that at stoichiometric mixture
the brake torque at both 0 and 60% EGR rates are 23.5 and 22.7 Nm, respectively.
This shows that the brake torque under Wide-Open-Throttle (WOT) and stoichiometric
condition and at 60% EGR rate was about 4% lower than that under normal condition
without EGR (i.e., 0% EGR rate). Also with the same operating conditions, but
at slightly lean mixture (=1.1),
the brake torque at both 0 and 60% EGR rates are 23 and 22.6 Nm. This is pointing
to the fact that at 60% EGR rates brake torque is approximately 2% lower than
that under normal condition without EGR (i.e., 0% EGR rate). While for the moderately
lean mixture (=1.2)
the brake torque respectively for both EGR rates under consideration are 21.5
and 20.5 Nm. This is an indication of about less than 5% reduction in brake
Brake specific fuel consumption: Figure 4 shows the variation of Brake Specific Fuel Consumption (BSFC) at different EGR rates for different speeds and injection timings at WOT. It is clear from the Fig. 4a that BSFC decreases as EGR rates increases. This represent an improvement in fuel consumption and it is due to the following (1) reduction in pumping work as EGR is increased at constant brake load (fuel and air flows remain almost constant, hence intake pressure increases). (2) reduced heat loss to the walls because the burned gas temperature is decreased significantly and (3) a reduction in the degree of dissociation in the high-temperature burned gases which allows more of the fuels chemical energy to be converted to sensible energy near TDC.
More so, BSFC decreases approximately as the mixture is richened (for maximum power) above stoichiometric due to the decreasing combustion efficiency associated with richening mixture. For mixture lean of stoichiometric, the theoretical BSFC decreases linearly as λ increases above stoichiometric. Combustion of mixtures leaner than stoichiometric produces product at lower temperature and with less dissociation of the triatomic molecules, CO2 and H2O. Thus the fraction of the chemical energy of the fuel which is released as sensible energy near TC is greater, hence a greater fraction of the fuels energy is transferred as work to the piston during expansion and the fraction of the fuels available rejected to the exhaust decreases.
In addition, at injection timing of 300°CA, increasing the engine speed
from 2000 to 3000 rpm, it shows that the rich mixture follows the anticipated
pattern of significant BSFC reduction until at about 15% EGR before it start
increase and this increment in BSFC is caused by larger cycle-by-cycle variations.
Deterioration in combustion starts to occur almost immediately on the lean side
of stoichiometric and the fuel consumption worsens for λ≥1.1.
|| Brake specific fuel consumption against EGR rates
Effects of EGR rates on emission parameters
Brake specific nitric oxide emissions (bsNO): Figure
5 shows the variation of brake specific nitric oxide emissions (bsNO) at
different EGR rates for different speeds and injection timings at WOT. From
Fig. 5 and 4a it is obvious that the Brake
Specific Nitric Oxide (BSNO) decreases as EGR rate increases i.e., NO concentration
is decreasing with increasing of EGR rates. This is because as EGR rate is increased,
the burning velocity and combustion maximum temperature will reduce due to the
dilution effect of EGR and also large specific heat capacity of CO2 and
H2O will absorb more released heat and decrease the cylinder gas
temperature. However, the decrease in cylinder gas temperature during combustion
process reduces the formation of NO and eventually results in decreasing the
In addition, the highest NO concentration is presented at excess air ratio
of 1.1 (
= 1.1) which is slightly leaner than the stoichoimetric equivalence ratio. The
oxygen concentration at slightly lean mixture results in higher NO concentration
compared with that at stoichoimetric equivalence ratio. Further increase in
excess air ratio will remarkably decrease the cylinder gas temperature and decrease
More so, lowest NO emissions consistent with good fuel consumption (avoiding the use of rich mixtures) are obtained with a stoichiometric mixture with as much as the engine will tolerate without excessive deterioration in combustion quality.
In addition at 300°CA injection timing and engine speed of 2000 rpm as
shown in Fig. 5(a) it is obvious that at stoichiometric mixture
= 1.0) the concentration of NO emission released at both 0 and 60% EGR rates
are 220 and 100 ppm/kW respectively. This shows that the concentration of NO
emission under Wide-Open-Throttle (WOT) and stoichiometric condition was about
50% lower than that under normal condition without EGR (i.e., 0% EGR rate).While
at slightly lean mixture (
= 1.1) with the same operating conditions, the concentrations of NO emissions
respectively for the EGR rates under consideration are 400 and 300 ppm kW-1.
This shows that at slightly lean mixture there is 25% reduction in concentration
of NO. Also at moderately lean mixture (
=1.2) the concentration of both EGR rates are 330 and 250 ppm kW-1
respectively. This reveals that there is about 24% reduction in the concentration
of NO at moderately lean mixture. Thus it is reasonable to say that large fluctuation
results at stoichiometric mixture as compare to the other mixtures i.e., NO
is more sensitive to EGR and excess air ratio ().
More so, at injection timing of 300°CA, increasing the speed from 2000
to 3000 rpm moderately increases NO concentration. The reason being that the
EGR rate decreases as the speed increases, the effects being greater at lower
inlet manifold pressures (lighter loads). Also the relative importance of heat
transfer per cycle is less as speed increases, which would be expected to increase
NO concentration. More over, by comparing the graphs in Fig. 5a
and b it is quite obvious that there is a large reduction
in NO concentrations at engine speed of 3000 as compare to engine speed at 2000
rpm. The reason might be due to large cycle-to-cycle variation which is caused
by variation in mixture composition or gas motion. However this cycle-to-cycle
variation is measured by coefficient of variation in indicated mean effective
pressure (COVIMEP) and detail about this is given in Fig.
9 for better and more comprehensive understanding.
It is noteworthy to say that the use of rich mixture should be avoided while
controlling the NO emissions in a direct injection compressed natural gas engine
with EGR. Figure 5a and b supports this
||Brake specific nitric oxide emissions (BSNO) versus EGR rates
Good agreement is achieved between these experimental results (Erjiang
et al., 2009).
Brake specific Unburnt Hydro Carbon (bsUHC): The variations of brake
specific Unburnt Hydrocarbon (UHC) emissions with EGR rates are shown in Fig.
6. From Fig. 6 (a) and (b) it is obvious
that at first the increase in HC is modest and is due primarily to decrease
HC burn up due to lower expansion and exhaust stroke temperatures. The HC increases
become more rapid as slow combustion, partial burning and even misfire in turn
occur with increasing frequency. Exception to these, is the rich mixtures (
=0.9) at engine speed of 2000 rpm and injection timing of 300°CA i.e., Fig.
6a. It shows an decreasing trend up to about 25% EGR and then starts to
increases. More so, rich mixture shows that emissions are high. This is primarily
due to lack of oxygen for after burning of any unburnt hydrocarbon that escape
the primary combustion process within the cylinder and exhaust system. HC emission
decrease as the stoichoimetric point is approached; increasing oxygen concentration
and increasing expansion and exhaust stroke temperature results in increasing
HC burn up.
||Brake specific Unburnt Hydrocarbon (bsUHC) versus EGR rates
For moderately lean mixture, HC emission level varies little with excess air ratio. Decreasing fuel concentration and increasing oxygen concentration essentially offset the effect of decreasing bulk gas temperatures. As the lean operating limit of the engine is approached, the combustion quality deteriorates significantly and HC emissions start to rise again due to occurrence of occasional partial burning cycles. For still leaner mixture, HC emission rise more rapidly due to the increasing frequency of partial-burning cycles and even the occurrence of completely misfiring cycle.
Comparing the concentration of HC emission at both engine speeds of 2000 and
3000 rpm respectively and at injection timing of 300°CA shows that higher
concentration of HC occurs at 3000 rpm at rich mixture which is about 1250 ppm
kW-1 i.e., Fig. 6 (b) as compared to the engine
speed at 2000 rpm (i.e., 1100 ppm kW-1) Fig. 6 (a).
This shows that when engine speed is increased from 2000 to 3000 rpm there is
about 12% increment in the concentration of HC emissions.
||Brake specific carbon monoxide emission versus EGR rates
This suggests that if oxygen is available, oxidation of unburned hydrocarbons
both within the cylinder and in the exhaust system will be significantly enhanced
by increases in speed since the expansion stroke and exhaust process gas temperature
increase substantially, due to the reduced significance of heat transfer per
cycle with increasing speed. This more than offsets the reduced residence time
in the cylinder and in the exhaust.
Brake specific carbon monoxide (bsCO): Figure 7 shows
the graphs of brake specific CO emission versus EGR rate for different engine
speeds and injection timings at WOT. From the Fig. 7a and
b it shows that EGR has no significant effect on engine CO
emissions. Highest level of CO emission is presented at rich mixture when compared
to stoichoimetric and lean mixture respectively, because complete oxidation
of the fuel carbon to CO2 is not possible due to insufficient oxygen
(incomplete combustion). For lean mixtures, CO levels are approximately constant
at low level of about 0.5% or less.
Effects of EGR rates on combustion parameters
Maximum cylinder pressure: Figure 8 gives the maximum
cylinder pressure at different EGR rates. The maximum cylinder pressure decreased
as EGR rates increases, this is largely due to the fact that with increase of
EGR mass fraction, the burning velocity and combustion maximum temperature will
reduced due to the dilution effect of EGR and consequent upon this cylinder
maximum pressure will decrease. Also the large specific heat capacity of CO2
and H2O will absorb more released heat and decrease the cylinder
gas temperature and this will give rise to reduction in cylinder maximum pressure.
The highest value of maximum pressure is presented at stoichoimetric in comparison
to lean and rich mixture, this is because there is efficient utilization of
oxygen concentration during combustion process at stoichoimetric as compared
to lean mixture.
In addition, when engine speed is increased from 2000 to 3000 rpm, there is
an increase in cylinder maximum pressure at excess air (
= 1.1) above stoichoimetric and this due to the fact that oxygen concentration
will be higher at excess air or slightly lean condition.
|| Maximum cylinder pressure versus EGR rates
Consequently, the burning velocity and combustion maximum temperature will rise and this will automatically leads to increase in cylinder maximum pressure.
The results equally shows that at the engine speed of 2000 rpm, when EGR exceed 40%, the maximum cylinder gas pressure is almost equal to the maximum motoring pressure (28bar) due to occurrence of misfire and/or the partial burning cycles. Therefore at engine of 2000 rpm and EGR rate of over 40% will not take into account in combustion analysis. When engine speed is increased to 3000 rpm, the increase in turbulence intensity increases the flame propagation speed and the engine can still maintain the normal combustion at EGR rate of 50% or more depending on the combustion strategy (i.e., rich or stoichoimetric or slightly lean or lean) and this is pointing to the fact that increasing the engine speed can extend the tolerated EGR limit. However, the maximum pressure variation depends on both changes in phasing and burning rate.
The magnitude of this variation depends on whether the combustion chamber is faster or slower burning, on average. It is also depend on the cyclic cylinder fuel and air charging variations.
Heat released rate: Figure 9 shows the effects of varying EGR dilution in the inlet mixture on the net heat release rate. As seen in this figure, there are cycle-by-cycle variations in the early stage of flame development (from 0 to few percent heat released rate) and in the major portion of the combustion process-the rapid burning phase-indicated by variation in a maximum burning rate. The maximum heat release decreases as the EGR rates increases. This is largely due to the fact that with the increase of EGR rates, the concentration of air and fuel in the cylinder decreases and this decrease the cylinder gas temperature and consequent upon this; the heat release rate decreases.
Also, considering Fig. 9a-d, starting from
(a) it shows that there is about -0.005 kJ/°CA heat released rate in the
beginning. The negative value implies that the energy is absorbed by the fuel
from the ignition source prior to combustion. However, the energy absorption
gives rise to increase in ignition delay and consequent to this, the combustion
duration increases, that explains the reason why the combustion duration at
stoichiometric (150°CA) is longer as compared to that at moderately lean
mixture (50°CA). In other words, it shows that moderately lean mixture has
a faster burning i.e. shorter combustion duration (due to increase in turbulence
intensity and also on the reaction rate which is dependent on the mixture composition)
as compared to stoichiometric mixture at the same engine speed.
||Heat released rates characteristics for different EGR rates
at 2000, 3000 and 5000 rpm and at different excess air ratios. (a) Heat
released rate @ 2000 rpm @ stoichiometric mixture (b) Heat released rate
@ 2000 epm @ moderately lean mixture (c) Heat released rate @ 3000 rpm @
In addition, at engine speed of 3000 rpm, the combustion duration respectively
for both mixtures (i.e. stoichiometric and moderately lean mixtures are 100
and 75°CA. These results obviously show that the moderately lean mixture
has a shorter combustion (due to faster burning) as compared to stoichiometric
mixture. However, the maximum heat released rate peaks at stoichiometric as
compared to that of lean burn. The reason being that at stoichiometric, all
the air provided is effectively utilized for combustion process (efficient combustion).
At engine speed of 2000 rpm, the maximum heat release rate occurs at stoichiometric mixture which is almost equal to 0.023 kJ/°CA at 20°CA ATDC as compared to moderately lean mixture which is 0.017 kJ/°CA at 18°CA ATDC, while at engine speed of 3000 rpm, the maximum heat released equally occurs at stoichiometric mixture which is approximately equal to 0.025 kJ/°CA at 20°CA as compared to moderately lean mixture which peaks at 0.020 kJ/°CA at about 15°CA ATDC.
In addition, as the mixture become leaner with excess air or more dilute with a higher burned gas fraction from residual gas or EGR, the magnitude of cycle-by-cycle combustion variation increases. Eventually some cycles become sufficiently slow burning that combustion is not completed by the time the exhaust opens, a regime where partial burning occurs in a fraction of the cycle is encountered. For even leaner or more dilute mixtures, the misfire limit is reached. At this point, the mixture in a fraction of the cycle fails to ignite.
Furthermore, heat released rate increases as the speed increases, i.e., increasing
the engine speed from 2000 to 3000 rpm, the maximum heat release rate varies
from 0.023 to 0.025 (stoichiometric mixture) and 0.017 to 0.020 kJ/CA (moderately
lean mixture). This is so, because increase in speed will increase the turbulence
intensity within the cylinder and since the rate of heat-release depends largely
on turbulence intensity and the reaction rate which is dependent on the mixture
composition, hence the heat-release rate will be increased. Figure
9a-d supports this claim.
||Mass fraction burned profile for different EGR rates at 2000,
3000 and 5000 rpm and at different excess air ratios. (a) Mass fraction
burned @ 2000 v rpm @ (stoichoimetric), (b) Mass fraction burned @ 2000
rpm (moderately lean), (c)Mass fraction burned @ 3000 rpm (stoichoimetric),
(d) Mass fraction burned @ 3000 rpm (moderately lean)
While the relative rates are essentially independent of speed, indicating that
combustion rate which depends on fuel-air mixing scale approximately with engine
Mass fraction burned: Figure 10 shows the mass fractioned burned as a function of a crank angle at different EGR rates. It is quite obvious that, increasing the EGR rates, the accumulated mass fraction burned grows up slowly, due to the decrease in burning (flame propagation) velocity. In furtherance, parts of the heat released will be absorbed as a result of introduction of large specific heat capacity gas like CO2 and H2O vapour from the exhaust and decrease the combustion temperature and consequent upon this, flame propagation speed decreased.
It is shown in Fig. 10a, that mass fraction burnt starts
with almost -0.15 at the time of the spark and then starts to increase from
approximately -30° to -15°CA after the spark timing. This crank angle
interval between the spark timing and start of combustion is called flame development
duration or sometimes ignition delay. However, flame development duration reflects
the flame development at early stage and it is related to the ignition delay
which is dependent on the mixture concentration and temperature.
More so, considering Fig. 10a and c i.e.,
increasing the engine speed from 2000 to 3000 rpm at stoichiometric mixture
clearly shows that increasing the engine speed increases the ignition delay
simply because residual gases in the cylinder will increases as a result of
increase in speed.
Following, the flame development duration, the mass fraction burned noticeably
increases with crank angle until it reaches its maximum value where essentially
almost all the fuel chemical energy has been released at basically the end of
combustion process. Figures 10a-d indicates
that the maximum mass fraction burnt, which essentially identifies the end of
combustion occurs later as the percentage of EGR dilution increases.
Determination of EGR-stable operating limit: COVIMEP is one important measure of cyclic variability, derived from pressure data. It is the standard deviation in imep divided by the mean imep and it is usually in percent (%). Mathematically, it is given as:
It is a measure of the cyclic variability in indicated work per cycle. Empirically,
it has been found that vehicle driveability problem usually results when COVIMEP
||Coefficient of variation in indicated mean effective pressure
It is also an important parameter to determine the EGR tolerance limit by the
engine. Based on these, the direct injection compressed natural gas engine-EGR
tolerance limit could therefore be determined by using COVIMEP not
exceeding 10% and this would be applied to Fig. 11 (a) and
(b) to determine the engine-EGR tolerance limit and also
define the engine operating condition at this point.
Figure 11 the result showed that when an engine speed of
2000 rpm, ignition timing of 31.5° CA, injection timing of 300° CA (early
injection timing), 18 bar and 315 k of pressure and temperature respectively
is applied, the EGR tolerance limit by the engine is indicated by the dotted
arrows pointing down for each of the mixture in Fig. 11a
and b, respectively and it is given as: (a) Rich mixture
= 0.9) is 50%, Stoichoimetric (
=1.0) is 60% or more, slightly lean (
=1.1) is 60% or more and Leaner mixture (
=1.2) is 45%.
When the engine speed is increased to 3000 rpm, while other operating condition
remains the same. (b) Rich mixture (
=0.9) is 30%, Stoichoimetric (
=1.0) is 50%, Slightly lean (
=1.1) is 60% or more and Leaner mixture (
=1.2) is 60% or more.
From the foregoing, it is cleared that the engine could still tolerate more
EGR at leaner mixture when engine speed is increased. this is so because an
increase in engine speed will increase the turbulence intensity and this in
turn increase the flame propagation speed and engine will maintain a normal
combustion when egr rates is increased. The experimental result is consistent
with (Erjiang et al., 2009). Also at rich and
stoichoimetric mixture; the engine will tolerate lesser EGR when the engine
speed is increased.
An experimental study of the different effects of EGR rate on the performance and exhaust emissions of the stratified charge piston DI-CNG engine has been conducted and the following are the major contribution to the body of knowledge:
||Engine could still tolerate more EGR, especially when operating
at moderately lean condition (
= 1.2) provided the engine speed is increased and this is largely due to
increase in turbulence intensity which give rise to increase in flame propagation
speed and engine will maintain normal combustion when the EGR rate is increased
||Degradation must be accepted in engine performance and efficiency when
using EGR to reduce NOX emissions
||By carefully optimizing the choice of operating parameters, EGR can improve
the fuel economy in DI-CNG. Figure 4 (a) supports this
||Average indicated mean effective pressure calculated for number
of a cycle (kPa)
||Cylinder gas pressure (kPa)
||Cylinder volume (m3)
||Mass of the cylinder gases ( kg)
||Standard deviation in indicated mean effective pressure (kPa)
||Coefficient of variation in indicated mean effective pressure
||Chemical energy released by combustion (J)
||Heat transfer to chamber wall (J)
||Change in sensible energy (J)
||Piston work (J)
||Mean gas temperature (K)
||Mean wall temperature (K)
||Specific heat at constant volume (J kg k-1)
||Specific heat at constant pressure (J kg k-1)
||Specific heat ratio
||Specific gas constant (J kg k-1)
||Change in crank angle (CA deg)