INTRODUCTION
Fossil fuels represent the major energy sources in the world. Unfortunately,
a lot of harms accompanied the use of these sources. Among the gaseous fossil
fuels, natural gas, which consists of about 90% methane, is gaining acceptance
because of its low emission levels and alternative energy source concerns (Arici,
1993). One of the important alternative fuels is hydrogen. Hydrogen is a
very efficient and clean fuel. Its combustion produces no greenhouse gases,
no ozone layer depleting chemicals and little or no acid rain ingredients and
pollution. Hydrogen, produced from the renewable energy (solar, wind, biomass,
tidal etc.) sources, would result in a permanent energy system which would never
have to be changed (Kahraman et al., 2007; Rahman
et al., 2009a, b). Hydrogen internal combustion
engine (H_{2}ICE) is a technology available today and economically viable
in the nearterm. This technology demonstrated efficiencies in excess of today’s
gasoline engines and operate relatively cleanly (Nox is the only emission pollutant)
(Boretti et al., 2007). Increases efficiencies,
high power density and reduce emissions are the main objectives for internal
combustion engine development (White et al., 2006).
One of the major parameter that effective in the improvement of performance
and emission regulation in the ICE is the amount of heat loss during the combustion
process. Therefore, the accurate model describes the heat transfer phenomena
for H_{2}ICE, which gives important information that required to improve
the simulation of engine performance.
The heat is convected from the intake port wall to the mixture charge. Heat
transfer mainly caused by forced convection, which is controlled by turbulent
charge movement and the temperature gradient of the mixture charge to the wall.
Whilst surveying the correlations for use in model of quasisteady heat transfer
in the piping system of internal combustion engine. Schubert
et al. (2005) noticed that the correlations are mostly based on the
similarity theory developed by Nuβelt:
where, α, d and k_{f} are heat transfer coefficient, cylinder diameter and fluid thermal conductivity, respectively.
The classical, steady correlations in the form of Nu = CRe^{m} or Nu
= CRe^{m}Pr^{n} are widely used for estimating the convective
heat transfer coefficient in the intake manifolds of engines (Dittus
and Boelter, 1930; Bauer et al., 1998; Depcik
and Assanis, 2002; Shayler et al., 1996)
because of a correlation type and the easytouse as well as unsteady heat transfer
model is not available. The constants C, m and n are adjusted to match the experimental
data to account for unsteady heat transfer enhancements, surface deposits and
surface roughness. Steady state correlations constants are listed in Table
1.
The amount of heat transfer is converted from the intake port wall to mixture charge calculates according to the Newton's law of cooling as Eq. 2:
where, Q, A, T_{g} and T_{w} are amount of heat transfer, heat transfer area, gas temperature and wall temperature, respectively.
In addition, frequencies based on valve events as well as pipe lengths, drastically
alter the flow patterns and change the heat transfer relationship (Depcik
and Assanis, 2002). Besides, type of the working gas another parameter will
give a significant effect on the heat transfer process. These correlations provide
reasonable agreement with experimental data in fullydeveloped steady pipe flows
and acceptable agreement with timeresolved experimental data in unsteady flows
and slow velocity variation under the quasisteady assumption. However, these
correlations can produce large errors in both phase and magnitude (Zeng
and Assanis, 2004) for highly unsteady flows with rapid velocity variation.
Experimental results published by different researchers show that the unsteady
flow effect in the engine intake manifold enhances heat transfer by 50 to 100%
over the prediction of the steady pipe flow correlation presented by Dittus
and Boelter (1930). At different engine speed and load conditions, the unsteadiness
of the flow condition is different.
Table 1: 
Steady state correlations constants 

Therefore, the constants (C and m) are usually optimized only for one operating
condition of a given engine and hence compromised for other conditions.
The physical properties of hydrogen fuel differ significantly from those fossil
fuels (Li and Karim, 2006; Verhelst
and Sierens, 2001). This provides the impetus for the authors to verify
the heat transfer process inside intake port for hydrogen fueled engine and
specify the differences with methane. Thus, the heat transfer process will be
taken into account for the present study to demonstrate the ability of the heat
transfer correlations which basically found for intake port with hydrocarbons
fueled engine to represent heat transfer process inside intake port with hydrogen
fueled engine as well as features detection of heat transfer phenomenon for
the intake port with the new alternative fuel.
MATERIALS AND METHODS
This study was conducted at Automotive Excellence Centre, Faculty of Mechanical Engineering, Universiti Malaysia Pahang, Kuantan. The duration of the project is April 2009 to August 2010.
Engine model: A single cylinder four stroke spark ignition port injection
with two valves (one inlet and one exhaust) model was developed utilizing GTPower
software. The injection of fuel was located in the midway of the intake port.
The computational model of single cylinder hydrogen fueled engine is shown in
Fig. 1. The schematic diagram for the model is demonstrated
in Fig. 2. The engine specifications are shown in Table
1. The specific engine characteristics are used to make the model A is shown
in Table 2. It is important to indicate that the intake and
exhaust ports of the engine cylinder are modeled geometrically with pipes. The
characteristics of the intake port of engine are shown in Table
3. The intake and exhaust ports of the engine cylinder are modeled geometrically
with pipes and the air enters through a bellmouth orifice to the pipe.
Table 2: 
Engine specifications for model (A) 

Table 3: 
Intake port characteristics 


Fig. 1: 
Model of single cylinder four stroke, port injection engine 

Fig. 2: 
Schematic diagram for the model 
The discharge coefficients of the bellmouth orifice were set to 1 to ensure
the smooth transition. The diameter and length of bellmouth orifice pipe are
0.07 and 0.1 m, respectively and it is connected to intake air cleaner with
0.16 m of diameter and 0.25 m of length. A log style manifold was developed
from a series of pipes and flowsplits. The total volume of each flowsplit
was 256 cm^{3}. The flowsplits compose from an intake and two discharges.
The intake draws air from the preceding flowsplit. One discharge supplies air
to adjacent intake runner and other supplies air to the next flowsplit. The
last discharge pipe was closed with a cup. The flowsplits are connected with
each other through pipes with 0.09 m diameter and 0.92 m length. Intake port
wall temperature value was assumed according to previous investigation results
(Bauer, 1997). The junctions between the flowsplits and
the intake runners were modeled with bellmouth orifice. The intake runners
were linked to the intake ports with 0.04 m diameter and 0.08 m length. The
temperature of the piston is higher than the cylinder head and cylinder block
wall temperature. Heat transfer multiplier is used to take into account for
bends, additional surface area and turbulence caused by the valve and stem.
The pressure losses are included in the discharge coefficients calculated for
the valves and no additional pressure losses were used because of wall roughness
(pressure losses have been estimated using dependency on Reynolds number only).
The exhaust runners were modeled as rounded pipes with 0.03 m inlet diameter
and 800 bending angle for runners 14 and 40° bending angle of runners 2
and 3. Runners 14 and runners 2 and 3 are connected before enter in a flowsplit
with 169.646 cm^{3} volume. Conservation of momentum is solved in 3dimentional
flowsplits even though the flow based on a onedimensional version of the NavierStokes
equation. Finally a pipe with 0.06 m diameter and 0.15 m length connects the
last flowsplit to the environment. Exhaust walls temperature was calculated
using a model embodied in each pipe and flowsplit. The air mass flow rate in
intake port was used for fuel flow rate based on the imposed equivalence ratio
(φ). The specific values of input parameters including the equivalence
ratio and engine speed were specified in the model. The boundary condition of
the intake air was defined first in the entrance of the engine. This object
describes end environment boundary conditions of pressure, temperature and composition.
The air enters through this object to the pipes. This object describes an orifice
placed between any two flow components and its parts represent the plane connecting
two flow components. The orifice diameter is set equal to the smaller of the
adjacent component diameter on the either side of the orifice. While the orifice
forward and reverse discharge coefficients are automatically calculated using
the geometry of the mating flow components and orifice diameter, assuming that
all transitions are sharpedged. The hydrogen and methane have been used to
represent as a fuel in current study for revealing the difference between of
these fuels in term of heat transfer characteristics inside intake port.
Model governing equations: One dimensional gas dynamics model is used
for representation of flow and heat transfer in the components of the engine
model. Engine performance can be studied by analyzing the mass, momentum and
energy flows between individual engine components and the heat and work transfers
within each component. Simulation of 1D flow involves the solution of the conservation
equations including the mass, momentum and energy in the direction of the mean
flow as Eq. 35, respectively:
where A, e,V and H are the cross sectional area, specific internal energy, element volume and enthalpy.
All properties for the charge have been computed by solving the above governing equations simultaneously based on explicit technique.
To complete the simulation model other additional formulas beside of the main governing equations are used for calculations of the pressure loss coefficient, heat transfer and friction coefficient. The pressure loss coefficient (C_{pl}) is defined by Eq. 6:
where, p_{1} and p_{2 }are the inlet and outlet pressure, respectively, ρ the density and u_{1} the inlet velocity. The heat transfer from the internal fluids to the pipe and flow split walls is dependent on the heat transfer coefficient, the predicted fluid temperature and the internal wall temperature. The heat transfer coefficient is defined as Eq. 4:
where, U_{eff}, C_{f}, C_{p} and Pr are the effective velocity outside boundary layer, friction coefficient, specific heat and Prandtl number, respectively. The effective velocity at any axial location inside intake port was considered as the rootmeansquare value of the instantaneous velocity.The friction coefficient for smooth walls can be expressed as Eq. 8:
where, Re_{D} and D are Reynolds number, pipe diameter and roughness height, respectively.
The Prandtl number can be expresses as Eq. 9:
where, υ and λ are the kinematic viscosity and thermal diffusivity, respectively. The amount of heat rate which is transferred from the inlet charge inside the intake port to its walls calculates according to the formula of Newton's law of cooling (Eq. 2).
RESULTS
Steady state gas flow and heat transfer simulations for the incylinder of four stroke port injection spark ignition engine model fueled. The hydrogen and methane is considered as a fuel with two operation parameters namely equivalence ratio (φ) and engine speed. The equivalence ratio was varied from stoichiometric limit (φ = 1.0) to a very lean limit (φ = 0.2). The equivalence ratio was change with interval of 0.2. The engine speed varied from 2000 rpm to 5000 rpm with interval of 1000 rpm. The heat transfer coefficient was used as an indicator for heat transfer characteristics to reveal the difference between hydrogen and methane fuels.
Table 4: 
Specifications of the engines models 

*NA: Not available 
Model validation: In the present study were approved by adopting experimental
results from two previous works. General assessment for the model performance
is dedicated in the first validation while the second validation is devoted
for revealing the extent and reliability of model results compared with previous
existing correlation for intake port heat transfer. The experimental results
obtained from Lee et al. (1995) were used for
purpose of first validation in this study. Engine specifications of Lee
et al. (1995) and present single cylinder port injection engine model
(B) are shown in Table 4. The same engine model which described
in Fig. 1 was used for the purpose of first validation (taking
into account the difference in the engines dimensions). Engine speed and equivalence
ratio were fixed at 1500 rpm and (φ = 0.5), respectively in this comparison
to be coincident with Lee et al. (1995) results.
The incylinder pressure traces for the baseline model (B) and experimental
previous published results (Lee et al., 1995)
are shown in Fig. 3. It can be seen that incylinder pressure
trace are very good match for compression stoke and acceptable trends for expansion
strokes while large deviation was obtained for combustion period due to the
delay in the combustion for experimental as in claim's of Lee
et al. (1995), beside the difference between the some engine configuration
conditions that is not mentioned by Lee et al. (1995).
However, considerable coincident between the present model (B) and experimental
results can be recognized in spite of the mentioned model differences. To demonstrate
the effectiveness of the adopted model for the present study model (A), direct
comparison with model (B) in term of incylinder pressure traces was done as
shown in Fig. 4. The difference between two models is due
to the difference in dimensions and compression ratio between of two models.
Correlation which introduced by Depcik and Assanis (2002)
is used for the purpose of model verification specifically for the intake port
heat transfer in present study.

Fig. 3: 
Comparison between published experimental results (Lee
et al., 1995) and present single cylinder port injection engine
model (B) based on incylinder pressure traces 

Fig. 4: 
Comparison between models (A and B) based on incylinder pressure
traces 
This correlation was proposed by using a least square curvefit of all available
experimental data to get a general relationship which describe a dimensionless
heat transfer coefficient Nu with Reynolds number expressed as Eq.
10:
Direct comparison between the acquired results from the engine model (A) and
the getting results from empirical correlation for hydrogen and methane fuel
is represented. Variation of heat transfer coefficient with engine speed for
hydrogen and methane with stoichiometric mixture is revealed in Fig.
5. It’s clear that the correlation performance for describing of methane
fuel give a good agreement with engine model results, where the deviation is
11% within correlation limit and 16% outside this limit, if taking into consideration
both of the deviation and limitation for the original correlation compared with
the original experimental results which used for fitting.
Where, correlation have an rvalue (deviation factor) of 0.846 and applicable
with range Reynolds number (1500<Re<40000) (Depcik
and Assanis, 2002) which correspond to engine speed values equal to (1500<RPM<4000)
for methane and (1500<RPM<4500) for hydrogen. The same trends achieved
for the hydrogen fueled engine model. Through this comparison can be determined
the extent and reliability model adopted in the present study. While, the correlation
results was under prediction for the performance of model (A) in case of hydrogen
fuel with deviation equal to 28%. This result is expected because the correlation
was derived mainly for hydrocarbon fuels.
Heat transfer coefficient for intake port: Comparison between hydrogen and methane in term of heat transfer coefficient and their behavior with changes of engine speed and equivalence ratio (φ) represents as indicator used to reveal the characteristics of steady state heat transfer inside the intake port for port injection H_{2}ICE. Direct comparison between hydrogen and methane in term the variation of heat transfer coefficient with engine speed is described in Fig. 6. The heat transfer coefficient is increasing as engine speed increase for methane and hydrogen fuels with keeping the highest values for methane fuel.

Fig. 6: 
Variation of heat transfer coefficient with engine speed for
equivalence ratio φ = 1.0 

Fig. 7: 
Variation of heat transfer coefficient on for methane fuel
against speed and equivalence ratio 
Effect of equivalence ratio (φ) on heat transfer coefficient with variation
of engine speed is shown in Fig. 7 and 8
for methane and hydrogen, respectively. The difference between methane and hydrogen
behavior is very clear in term of equivalence ratio (φ). In case of methane
there are no effect (or negligible) for equivalence ratio (φ) on the values
and behavior of heat transfer coefficient. As a result, it is expected that
there will be no any impact for this variable on the overall process of heat
transfer.

Fig. 8: 
Variation of heat transfer coefficient for hydrogen fuel with
speed and equivalence ratio 
On the contrary, it can be seen the impact of this factor in case of hydrogen.
It decreases the equivalence ratio (φ) values heat transfer increase due
to disappear of the blockage phenomenon.
DISCUSSION
The effects of engine speed and equivalence ratio on heat transfer coefficient
are illuminated in case of hydrogen and methane fuels in the previous section.
The behavior of heat transfer coefficient is found to be governed by the blockage
and forced convection effects. Forced convection effects are related to engine
speed variation while the blockage effects related to equivalence ratio variation.
It can be seen that increases of heat transfer coefficient with engine speed
increases due to increasing the driving force for the heat transfer process
(forced convection) for hydrogen. The similar trends were observed for methane
fuel however methane is given greater values about 11% for engine speed compare
with hydrogen fuel. The similarity of heat transfer coefficient behavior with
variation of engine speed is expected because there is no relation between engine
speed and fuel properties. Density of methane fuel is greater than hydrogen
as well as the diffusion coefficient for methane is lesser than hydrogen, hence
the blockage effect for hydrogen fuel is greater than methane so that the present
of forced convection with methane is more strength due to the high restriction
for the charge flow in case of hydrogen fuel. Therefore, the heat transfer effect
is more efficient in case methane fuel. The results are coincident with previous
results which are acquired by Dittus and Boelter (1930),
Bauer et al. (1998), Depcik
and Assanis (2002) and Shayler et al. (1996)
with some difference in heat transfer coefficient because of the variation for
adjusting constants which are function for experimental conditions.
The variation of equivalence ratio (φ) in case of methane fuel has a very minor influence on heat transfer coefficient inside intake port. On the other hand, in case of hydrogen fuel, by decreasing the equivalence ratio (φ) lead to enhancement of heat transfer characteristics due to the deterioration for the blockage effect consequently. The flow of gas becomes more fluently that means the effect of the forced convection is more strength. This difference in heat transfer coefficient behavior with change of the equivalence ratio due to the large difference in physical properties in terms of density and diffusivity. Rate of increase in heat transfer coefficient comparison with stoichiometric case for hydrogen fuel are: 4 and 8% for the equivalence ratio of 0.6 and 0.2, respectively.
CONCLUSION
The present study considers the comparison in heat transfer characteristics inside the intake port for port injection engine fueled with hydrogen and methane, respectively. The foregoing results indicates that heat transfer coefficient in the intake port is changed with variation of engine speed and equivalence ratio due to the dissimilarity of fuels properties. Comparison between hydrogen and methane in term of heat transfer coefficient and their behavior with change of engine speed and equivalence ratio are clarified that hydrogen is more dependable on equivalence ratio, whilst both of them have the same trend with engine speed variation. The blockage effect was affecting the heat transfer process dominantly in case of hydrogen fuel, due to the low density and high diffusion velocity for hydrogen in comparison with methane.
ACKNOWLEDGMENT
The authors would like to thank Universiti Malaysia Pahang for provides laboratory facilities and financial support under project No. RDU0903093 and Doctoral Scholarship Scheme (GRS 090121).