Energy is currently uppermost in everyones minds. Recent price hikes for domestic
gas and electricity, coupled with fluctuating oil prices, concerns about where
our future energy will come from and the impact of energy use on climate change
have all brought energy policy to the top of the political agenda. The alarming
rise in pollution levels in the atmosphere and increased concern for energy
independency are the two major driving factors for seeking alternative fuels.
Among the viable options, hydrogen is the only non-carbonaceous fuel available
on the earth. Therefore, hydrocarbon and carbon monoxide free engine operation
with hydrogen. Hydrogen has been regarded as a future secondary fuel for power
systems due to CO2 and hydrocarbon free operation. Recent drastic
increase in the price of petroleum, rapid increase in emission of green house
gases and very strict environmental legislations are major motivating factors
for usage of hydrogen in fuel cells and internal combustion engines. Hydrogen
as a fuel for internal combustion engine is best suited for spark ignition due
to its high self-ignition temperature (Ganesh et al.,
2008). A lean mixture results in an improvement of fuel economy. More complete
combustion, lower combustion temperature and reduced pollutants (COD,
2001). Normally, cycle by cycle variations occur in the engines operating
with very lean mixtures. But, with hydrogen, these variations are much less
compared to that of engines powered by other hydrocarbon fuels (Kim
et al., 2005). Hydrogen needs very low ignition energy, which ensures
prompt ignition. Hydrogen induction techniques play a very dominant and sensitive
role in determining the performance characteristics of the hydrogen fueled internal
combustion engine (Suwanchotchoung, 2003). Hydrogen fuel
delivery system can be broken down into three main types including the carbureted
injection, port injection and direct injection (COD, 2001).
The port injection fuel delivery system injects hydrogen directly into the
intake manifold at each intake port rather than drawing fuel in at a central
point. Typically, hydrogen is injected into the manifold after the beginning
of the intake stroke (COD, 2001). Hydrogen can be introduced
in the intake manifold either by continuous or timed injection. The former method
produces undesirable combustion problems, less flexible and controllable (Das,
1990). But the latter method, timed port fuel injection (PFI) is a strong
candidate and extensive studies indicated the ability of its adoption (Das
et al., 2000; Das, 2002). This technique is
supported by a considerable set of advantages. It can be easily installed with
simple modification (Lee et al., 1995) and its
cost is low (Li and Karim, 2006). The flow rate of hydrogen
supplied can also be controlled conveniently (Sierens and
Verhelst, 2001). External mixture formation by means of port fuel injection
also has been demonstrated to result in higher engine efficiencies, extended
lean operation, lower cyclic variation and lower NOx production (Yi
et al., 2000; Rottengruber et al., 2004;
Kim et al., 2006). There are consequence of the
higher mixture homogeneity due to longer mixing times for PFI. Furthermore,
external mixture formation provides greater degree of freedom concerning storage
methods (Verhelst et al., 2006). The most serious
problem with PFI is the high possibility of pre-ignition and backfire, especially
with rich mixtures (Kabat and Heffel, 2002; Ganesh
et al., 2008). However, conditions with PFI are much less severe
and the probability for abnormal combustion is reduced because due to its imparts
a better resistance to backfire (COD, 2001). Combustion
anomalies can be suppressed by accurate control of injection timing and elimination
of hot spots on the surface of the combustion as suggested by
Lee et al. (1995). The present study is developed the computational
model for a single cylinder, port injection hydrogen fueled internal combustion
engine. The objective of this study is to investigate the effect of air fuel
ratio and engine speed on the performance characteristics of this engine.
MATERIALS AND METHODS
Engine performance parameters: The brake mean effective pressure
(BMEP) can be defined as the ratio of the brake work per cycle Wb
to the cylinder volume displaced per cycle Vd and it can be
expressed as in Eq. 1:
Equation 1 can be rewrite for the four stroke engine
as in Eq. 2:
where, Pb is the brake power and N is the engine speed.
Brake efficiency (ηb) can be defined as the ratio of
the brake power to the engine fuel energy which is expressed as in Eq.
is the fuel mass flow rate and LHV is the lower heating value of hydrogen.
The brake specific fuel consumption (BSFC) represents the fuel flow rate
per unit brake power output and can be expressed as in Eq.
The volumetric efficiency (ηv) of the engine is defines
as the mass of air supplied through the intake valve during the intake
by comparison with a reference mass, that mass required to perfectly fill
the swept volume under the atmospheric conditions. It can be expressed
as in Eq. 5:
where, ρai is the inlet air density.
Engine model: A single cylinder, four stroke, port injection hydrogen
fueled engine was modeled utilizing the GT-Power software. The injection
of hydrogen was located in the midway of the intake port. The model of
the hydrogen fueled single cylinder four stroke port inject engine is
shown in Fig. 1. Engine specifications for the base
engine are shown in Table 1. The specific values of
input parameters including the AFR, engine speed and injection timing
were defined in the model. The boundary condition of the intake air was
defined first in the entrance of the engine. The air enters through a
bell-mouth orifice to the pipe. The discharge coefficients of the bell-mouth
orifice were set to 1 to ensure the smooth transition as in the real engine.
The pipe of bell-mouth orifice with 0.07 m of diameter and 0.1 m of length
are used in this model. The pipe connects in the intake to the air cleaner
with 0.16 m of diameter and 0.25 m of length was modeled. The air cleaner
pipe identical to the bell-mouth orifice connects to the manifold.
||Hydrogen fueled engine parameters
||Model of single cylinder, four stroke, port injection
hydrogen fueled engine
style manifold was developed from a series of pipes and flow-splits. Firstly,
an attribute heat transfer multiplier is used to account for bends, roughness
and additional surface area and turbulence caused by the valve and stem.
Also, the pressure losses in these ports are included in the discharge
coefficients calculated for the valves. The total volume for each flow-split
was 256 cm3. The flow-splits compose from an intake and two
discharges. The intake draws air from the preceding flow-split. One discharge
supplies air to adjacent intake runner and the other supplies air to the
next flow-split. The last discharge pipe was closed with a cup to prevent
any flow through it because there is no more flow-split. The flow-splits
are connected with each other via., pipes with 0.09 m diameter and 0.92
m length. The junctions between the flow-splits and the intake runners
were modeled with bell-mouth orifices. The discharge coefficients were
also set to 1 to assure smooth transition, because in most manifolds the
transition from the manifold to the runners is very smooth. The intake
runners for the four cylinders were modeled as four identical pipes with
0.04 m diameter and 0.1 m length. Finally the intake runners were linked
to the intake ports which were modeled as pipes with 0.04 m diameter and
0.08 length. The air mass flow rate in the intake port was used for hydrogen
flow rate based on the imposed AFR.
The in-cylinder heat transfer is calculated by a formula which closely
emulates the classical Woschni correlation. Based on this correlation,
the heat transfer coefficient (hc) can be expresses as in Eq.
where, B is the bore in meters, p is the pressure in kPa, T is temperature
in K and w is the average cylinder gas velocity in m sec-1.
|| Temperature of the mail engine parts
||Parameters used in the exhaust environment
The combustion burn rate (Xb) using Wiebe function, can be
expressed as in Eq. 7:
where, θ is the crank angle, θi is the start of
combustion, Δθ the combustion period and a and n are adjustable
parameters. The overall temperature of the head, piston and cylinder for
the engine parts are shown in Table 2. The temperature
of the piston is higher than the cylinder head and cylinder block wall
temperature because this part is not directly cooled by the cooling liquid
or oil. Exhaust system walls temperature was calculated using a model
embodied in each pipe and flow-split. Table 3 shown
the parameters used in the exhaust environment of the model.
RESULTS AND DISCUSSION
It is worthy to mention that one of the most attractive combustive features
for hydrogen fuel, which is its wide range of flammability. A lean mixture
is the amount of fuel less than stoichiometric mixture. This leads to
fairly easy to obtain an engine start. In the present model, hydrogen
was injected into the cylinder within a timing range started just before
IVC (-960 BDC) until TDC (00).
|| Variation of brake mean effective pressure with air
fuel ratio for various engine speeds
The air-fuel ratio
was varied from stoichiometric limit (AFR = 34.33:1 based on mass where
the equivalence ratio φ = 1) to a very lean limit (AFR =171.65 based
on φ = 0.2) and engine speed varied from 2500 to 4500 rpm. Amount
of hydrogen injected in one cycle is approximately 22 mg cycle-1 with injection pulse duration of 4.4 m sec.
Figure 2 shows the effect of air-fuel ratio on the brake
mean effective pressure. BMEP is the good parameter for comparing engines with
regard to design due to its independent on the engine size and speed. The large
engine was always seem to be better when consider the torque, however, speeds
become very important when consider the power (Pulkrabek, 2003).
It is obtained from the results that the decreases of the BMEP with increases
of AFR and speed. It is obvious that the BMEP falls with a non-linear behavior
from the rich condition where AFR is 34.33 to the lean condition where the AFR
is 171.65. The difference of BMEP increases with increases of speed and AFR.
The differences of the BMEP are decrease of 6.682 bar at speed of 4500 rpm while
6.12 bar at speed 2500 rpm for the same range of AFR. This implied that the
engine gives the maximum power (BMEP = 1.275 bar at lower speed 2500 rpm) compared
with the power (BMEP = 0.18 bar) at speed 4500 rpm. Due to dissociation at high
temperature combustion, molecular oxygen is present in the burned gases under
stoichiometric conditions. Thus some additional fuel is added and partially
burned. This increases the temperature and the number of moles of the burned
gases in the cylinder. These effects increase the pressure; those were given
increase power and mean effective pressure (Heywood, 1988).
Figure 3 shows the variation of the brake thermal efficiency
with the air fuel ratio for the selected speeds. The brake power as a
percentage is considered for the intake fuel energy and the fuel energy
are covered the friction losses and heat losses (heat loss to surroundings,
exhaust enthalpy and coolant load).
||Variation of brake thermal efficiency with air fuel
|| Variation of brake specific fuel consumption with air
fuel ratio for different engine speed
It can be observed that the brake
thermal efficiency increases nearby the richest condition (AFR ≅
35) and then decreases with increases of AFR and speed. The operation
within the range of AFR from 49.0428 to 42.91250 (φ = 0.7 to 0.8)
give the maximum values for ηb for all speeds. Maximum
ηb can be seen of 31.8% at speed 2500 rpm compared with
26.8% at speed 4500 rpm. Unaccepted ηb is observed of
2.88% at very lean conditions with AFR of 171.65 (φ = 0.2) for speed
of 4500 rpm while 20.7% at the same conditions with speed of 2500 rpm.
Clearly, engine rotational speed has a major effect in the behavior of
ηb with AFR due to higher speeds lead to higher friction
Figure 4 depicts the behavior of the brake specific fuel
consumption with AFR. The AFR for optimum fuel consumption at a given load depends
on the details of chamber design including compression ratio and mixture preparation
|| Variation of maximum cylinder temperature with airfuel
It varies for a given chamber with the part of throttle load and speed
range (Heywood, 1988). It is clearly seen from Fig.
4 that the higher fuel is consumed at higher speeds and AFR due to the greater
friction losses that can occur at high speeds. It is easy to perceive from the
figure that the increases of BSFC with decreases in the rotational speed and
increases the value of AFR. However, the required minimum BSFC were occurred
within the range of AFR from 38.144 (φ = 0.9) to 49.0428 (φ = 0.7) for
the selected range of speed. At very lean conditions, higher fuel consumption
can be noticed. After AFR of 114.433 (φ = 0.3) the BSFC increase rapidly,
especially for high speed. At very lean conditions with AFR of 171.65 (φ
= 0.2), the BSFC of 144.563 g kW-1 h was observed at 2500 rpm while
1038.85 g kW-1 h for speed of 4500 rpm. The value BSFC at speed 2500
rpm was observed around 2 times at 4000 rpm, however, around 7 times at 4500
rpm. This is because of very lean operation conditions, which can lead to unstable
combustion and more lost power due to a reduction in the volumetric heating
value of the air/hydrogen mixture. This behavior can be more clarified by Fig.
3, where the brake efficiency reduced considerably at very lean operation
Figure 5 shows the behavior AFR can affect the maximum
temperature inside the cylinder. In general, lower temperatures are required
due to the reduction of pollutants. It is clearly demonstrated that the
decrease of maximum cylinder temperature with increases of AFR. The effect
of the engine speed on maximum cylinder temperatures with AFR are not
significant. At stoichiometric operating conditions (AFR = 34.33), the
maximum cylinder temperature of 2752.83 K was recorded and it is dropped
to 1350K at AFR of 171.65 (φ = 0.2). This lower temperature inhibits
to the formation of NOx pollutants.
|| Effect of volumetric efficiency with the rotational
speed for different equivalence ratio
In fact this feature is
one of the major motivations toward hydrogen fuel.
Figure 6 shows the variation of the volumetric efficiency
with the engine speed. The volumetric efficient increase with increases of AFR.
In general, it is desirable to have maximum volumetric efficiency for engine.
The importance of volumetric efficiency is more critical for hydrogen engines
because of the hydrogen fuel displaces large amount of the incoming air due
to its low density (0.0824 kg m-3 at 25°C and 1 atm). This reduces
the volumetric efficiency to high extent. The stoichiometric mixture of hydrogen
and air consists of approximately 30% hydrogen by volume, whereas a stoichiometric
mixture of fully vaporized gasoline and air consists of approximately 2% gasoline
by volume (White, 2006). Therefore, the low volumetric
efficiency for hydrogen engine is expected compared with gasoline engine works
with the same operating conditions and physical dimension. This lower volumetric
efficiency is apparent in Fig. 6. Leaner mixture gives the
higher volumetric efficiency. The maximum volumetric efficiency was observed
79.4% at lean conditions with AFR = 171.65 (φ = 0.2) while 62.4% at stoichiometric
Higher speeds lead to higher volumetric efficiency because of the higher speeds
give higher vacuum at the port and consequent larger air flow rate. Further
increase in engine speed leads toward the maximum value of ηv.
The volumetric efficiency increases with increases of engine at certain limit
then decreases. For equivalences ratio until below 0.6, the maximum ηv
was recorded at 4200 rpm. For equivalence ratio below 0.6, the maximum ηv
was recorded at 3800 rpm.
|| Variation of combustion duration with engine speed
for different equivalence ratio
At further higher engine speeds beyond these values,
the flow during at least part of the intake process becomes chocked. Once chocked
occurs, further increase of speed does not increase the flow rate significantly,
thus, the volumetric efficiency decreases sharply. This sharp decrease happens
due to the charge heating in the manifold and higher friction flow losses. In
fact that the several solution were suggested to solved this problem (Nagalingam et al., 1983; Furuhama and Fukuma, 1986; Lynch, 1983) suggested and carried out tests with pressure
boosting systems for hydrogen engine.
Figure 7 shows the combustion duration with the engine speed
for different equivalence ratio. The hydrogen combustion velocity (1.85 m sec-1)
is rapid compared with that of gasoline (0.37-0.43 m sec–).
Therefore, the short combustion duration is expected. It is well established
that the duration of combustion in crank angle (degree) increases slowly with
increases of speed for gasoline and diesel engines (Heywood,
1988). This fact is also true for hydrogen engine (Fig. 7).
The fluctuation shown in Fig. 7 is very small, however, it
is enlarged in the Fig. 7 with a very high scale. The fluctuation
take place within the range of 0.0248°CA.
The performance characteristics of single cylinder port injection hydrogen
fueled engine were investigated. The following conclusions are drawn:
||At very lean conditions with low engine speeds, acceptable BMEP
can be reached, while it is unacceptable for higher speeds. Lean operation
leads to small values of BMEP compared with rich conditions
||Maximum brake thermal efficiency can be reached at mixture composition
in the range of (φ = 0.7 to 0.9) and it decreases dramatically
at leaner conditions.
||The desired minimum BSFC occurs within a mixture composition range
of (φ = 0.7-0.9). The operation with very lean condition (φ<0.2)
and high engine speeds (> 4500) consumes unacceptable amounts of
||Lean operation conditions results in lower maximum cylinder temperature.
A reduction of around 1400K can be gained if the engine works properly
at (φ = 0.2) instead of stoichiometric operation
||The low values of volumetric efficiency seem a serious challenge
for the hydrogen engine and further studied are required
The authors would like to express their deep gratitude to Universiti
Malaysia Pahang (UMP) for provided the laboratory facilities and financial